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Applying condensing-temperature control in air-cooled reciprocating water chillers for energy efficiency
Applied Energy 72 (2002) 565–581 www.elsevier.com/locate/apenergy
Applying condensing-temperature control in air-cooled reciprocating water chillers for energy efficiency K.T. Chan*, F.W. Yu Department of Building Services Engineering, The Hong Kong Polytechnic University, Hong Kong, China Received 1 May 2002; received in revised form 17 June 2002; accepted 22 June 2002
Abstract This paper reports on the modelling and findings of the energy performance of an aircooled reciprocating multiple-chiller plant under the conventional head pressure control and the new condensing-temperature control in a subtropical climate. The simulation model was validated using the operating data of an existing chiller plant. As noted from this existing aircooled reciprocating chiller plant, there was a substantial efficiency drop at part-load resulting from the head pressure control. If operating at variable lower condensing-temperatures based on the established operating mode of the condenser fans and compressors, it is shown that the chiller consumption can be maintained below 2 kW/refrigeration ton throughout the entire range of outdoor temperature and part-load conditions, giving an average efficiency of 1.08 kW/refrigeration ton. The energy imposition due to cycling on more condenser fans can be compensated by the reduced compressor consumption. Potential energy savings of 18.2 and 29% in the annual chiller consumption are achievable by applying the condensing-temperature control to two existing chiller plants studied. This supports the need to develop the condensing-temperature control as an improvement to the conventional head pressure control. # 2002 Elsevier Science Ltd. All rights reserved. Keywords: Air-cooled chillers; Condensing temperature control; Head pressure control; Energy efficiency
1. Introduction Air-cooled reciprocating chillers have been widely installed in local commercial buildings for providing space cooling and dehumidification. The operations of chiller plants have a significant impact on electricity consumption and maximum power * Corresponding author. Fax: +852-27746146. E-mail address: [email protected] (K.T. Chan). 0306-2619/02/$ - see front matter # 2002 Elsevier Science Ltd. All rights reserved. PII: S0306-2619(02)00053-3
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Nomenclature Symbols AU overall heat-transfer coefficient (kW K-1) CR compression ratio Cpa specific heat-capacity of air (kJ kg1 K-1) Cpw specific heat-capacity of water (kJ kg1 K-1) Ecc power input to operating compressors (kW) Ecf power input to operating condenser fans (kW) Ech overall power input to a chiller (kW) Fcc derating factor for the compressor efficiency hi specific enthalpy of refrigerant at state i kW/refr. ton chiller operating efficiency in kW per refrigeration ton mr refrigerant mass flow rate per compressor (kg s1) mw chilled water mass flow rate (kg s1) n index of reversible polytropic expansion process Ncc number of operating compressors P saturated pressure of refrigeration circuit (kPa) PLR chiller part-load ratio (given by Qcl/Qcr) Qcd total heat rejection (kW) Qcl required chiller cooling output (kW or tons of refrigeration) Qcr rated chiller cooling capacity (kW or tons of refrigeration) qrf refrigeration effect (kJ kg1) T saturated temperature of refrigeration ( C) Tcdae entering condenser air temperature or outdoor air temperature ( C) Tcdal leaving condenser air temperature ( C) Tchwr return chilled water temperature ( C) Tchws supply chilled water temperature ( C) Va airflow provided by operating condenser fans (m3 s1) Vp piston displacement of each compressor (m3 s1) vr refrigerant specific volume at compressor suction pressure (m3 kg1) win isentropic work input to the compressor (kJ kg1) " heat-exchange effectiveness isen isentropic efficiency cc compressor efficiency v volumetric efficiency of a reciprocating compressor a air density (kg m3)
Subscripts cd ev tot
condenser evaporator total
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demand. The chiller coefficient of performance (COP) of two existing chiller plants was less than 2.5 most of the time [1], which did not meet with the minimum COP of 2.7 stipulated in the local energy code for energy efficiency of air-conditioning installations [2]. This poor chiller efficiency was associated with oversized equipment [3], inappropriate matching of operating chillers with the cooling-load demand, and the chillers always worked at a high condensing-temperature based on the design outdoor condition of 33 C (often referred to as head pressure control) [1]. The local weather data [4] indicated that the outdoor temperatures are below 25 C half of the time in a year. There is a considerable scope for lowering the condensing temperature and thereby improving the chiller efficiency [1,5]. An experimental study by Smith and King [5] on the reduction in condenser pressure lift for a reciprocating chiller rated at 35 kW cooling capacity illustrated that a 10% reduction in the overall chiller consumption was achieved by driving the condenser fan harder at all outdoor temperatures below 25 C. Nevertheless, no detailed algorithms have been reported for variable condensing-temperature. The trade-off between increased condenser fan power and decreased compressor power, and the long-term energy saving by lowering the condensing temperature based on the weather and part load conditions has not been assessed. Many simulation models for chiller systems were documented [6–12]. There were a few models developed for air-cooled reciprocating chillers: among those was a reciprocating chiller model by Leverenz and Bergan [10] for hourly energy analysis using a ‘‘black box’’ concept. Such a concept directly relates the input and output parameters of the chiller by a set of empirical equations without considering the operating balance among the components of the evaporator, compressor and condenser. Moreover, an optimisation study cannot be carried out using the ‘‘black box’’ models, as both the controllable and controlled variables are absent. Other reciprocating chiller models deal with a set of energy-balance equations without considering the capacity control of the reciprocating compressors influencing the power input [11]. A mechanistic model taking into account the thermodynamics and operating balance within the components of an air-cooled reciprocating chiller is developed here for the study of the chiller performance under the conventional head pressure control and the condensing temperature control. Necessary calibration of the model is made according to the manufacturer’s specification and the operating data from two existing chillerplants. This paper presents the validated chiller model, the chiller’s operating parameters and the analysis of chiller efficiency. Energy saving potential by lowering the condensing temperature at part load conditions for the two existing chiller plants is assessed.
2. Model description 2.1. Characteristics of the modelled chiller An air-cooled reciprocating chiller, using refrigerant R-22 (chlorodifluoromethane) with rated capacity of 199 tons of refrigeration (i.e. 700 kW, 1 refrigeration ton=3.52 kW), was studied. The modelled chiller is commonly used in multiple-chiller
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Fig. 1. Refrigeration cycle of the modelled chiller.
systems for small to medium sized buildings requiring cooling capacity in the range of 300–1000 tons of refrigeration (1056–3520 kW). In general, individual chillers with different or identical sizes in a multiple-chiller plant operate in parallel to meet the varying cooling load demand. The chiller has six steps of capacity control from full load down to a minimum of 16.7% by staging the six compressors, and contains two refrigerant circuits to enhance the reliability and standby capacity. In order to meet the required heat rejection under the head pressure control, twelve steps of control on the heat rejection airflow rate are achieved by cycling the twelve condenser fans. 2.2. Refrigeration cycle and assumptions for the model Fig. 1 shows the refrigeration cycle for the modelled chiller. In the refrigeration cycle, heat loss or gain from the surroundings is ignored. Pressure drops in the refrigerant pipelines are neglected. It is assumed that the refrigerant throttling process through the expansion valve is adiabatic. The compression process is polytropic and the actual compression work is the isentropic compression work input, win divided by the isentropic efficiency, isen. The degrees of sub-cooling and superheat are maintained at 6.5 and 5 C respectively under all loading conditions. 2.3. Evaporator model The chiller modelled contains a shell and tube evaporator. Energy balance for the heat flow between the refrigerant and water at the evaporator can be expressed as: Qcl ¼ mw Cpw ðTchwr Tchws Þ
ð1Þ
¼ Ncc mr qrf
ð2Þ
¼ mw Cpw "ev ðTchwr Tev Þ
ð3Þ
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The mass flow rate of chilled water, mw and the supply chilled-water temperature, Tchws are maintained at 27.4 kg s1 and 7 C respectively based on the manufacturer’s recommendation. The return chilled-water temperature, Tchwr will vary with the cooling capacity, Qcl in the range of 7.9–12.5 C. Referring to Fig. 1, the refrigeration effect, qrf in Eq. (2) is given by: qrf ¼ h1 h4
ð4Þ
In determining the evaporating temperature, Tev, the classical heat-exchange effectiveness method is applied using Eq. (3). The heat-exchange effectiveness at the
Fig. 2. Flow chart for parameter evaluation.
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waterside, "ev is determined by Eq. (5). The overall heat-transfer coefficient, AUev is described by the empirical relation in Eq. (6) [12], where c1, c2 and c3 are constants evaluated using the chiller’s performance data. AUev "ev ¼ 1 exp ð5Þ mw Cpw AUev ¼
c1 m0:8 w
1 þ c2 Q0:745 þ c3 cl
ð6Þ
For a constant supply chilled-water temperature, Tchws, the evaporating temperature, Tev varies from 5.3 to 5.8 C as calculated by Eq. (6) based on the anticipated range of return chilled-water temperature. The evaporating pressure, Pev and hence the specific enthalpy of refrigerant at compressor suction state, h1 can be determined using the state equations of the refrigerant R-22 [13]. The variation of the refrigeration effect, qrf, is small under the head pressure control but is substantial under the condensing temperature control with varying condensing temperatures. The refrigeration effect increases with the reduction in the condensing temperature at a constant evaporating temperature. For a certain cooling capacity, the number of operating compressors, Ncc, can be determined by Eq. (2). The refrigerant mass flow rate is calculated by Eq. (8) and the refrigeration effect is evaluated from the refrigerant’s state equations. 2.4. Compressor model The compressor power input, refrigerant mass flow rate per compressor at the compression process and the volumetric efficiency [14] are given by Eqs. (7)–(9): Ecc ¼ Ncc mr
mr ¼
win isen cc
Vp v vr
v ¼ 0:94 0:015 CR
ð7Þ
ð8Þ ð9Þ
where CR is the compression ratio: CR ¼
Pcd Pev
ð10Þ
The piston displacement of a compressor, Vp, in Eq. (8) relating to the cylinder diameter, stroke of the piston and number of cylinder per compressor, is regarded as a constant for a constant-speed compressor. It can be evaluated using the manufacturer’s data at the full-load condition. The specific volume of refrigerant at the
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suction of the compressor, vr was determined from the evaporating pressure and the degree of superheat using the refrigerant state equations. The compression ratio varies with the condensing pressure lift to a small extent under the head pressure control, but it becomes a dependent variable changing with the condensing pressure under the condensing-temperature control. Based on the refrigeration cycle in Fig. 1, the isentropic work input by the compressor, win in Eq. (7), is given by: n n- 1 CR n 1 win ¼ Pev vr ð11Þ n1 The isentropic compression index, n, is taken to be 1.2 for the semi-hermetic compressor motors by refrigerant cooling. The specific enthalpy of the refrigerant at state 2 can be evaluated using Eq. (12): h2 ¼ h1 þ
win isen
ð12Þ
where the isentropic efficiency, isen, can be calculated using the empirical equation in relation to the compression ratio, CR, [15]: isen ¼ 0:73 þ 0:1 CR 0:026 CR 2 þ 0:0017 CR 3
ð13Þ
Since there is no mathematical relation governing the compressor’s efficiency, cc, in the first trial, it is approximated by the volumetric efficiency with an additional 10% to offset the power loss due to the mechanical friction in the compressor [16]: cc ¼ 1:1 v
ð14Þ
Verification of the approximation is carried out subsequently by comparing the simulation result with the operating data. 2.5. Condenser model The chiller is installed with an air-cooled condenser of finned coils integrated with sub-cooling coils. The total heat rejection, Qcd, is the sum of refrigeration load, Qcl, and compressor power input, Ecc. It can be calculated by the following energy balance equations: Qcd ¼ Qcl þ Ecc
ð15Þ
¼ Ncc mr ðh2 h3 Þ
ð16Þ
¼ Va a Cpa ðTcdal Tcdae Þ
ð17Þ
¼ Va a Cpa "cd ðTcd Tcdae Þ
ð18Þ
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The evaluation of the condenser’s effectiveness, "cd, together with AUcd is analogous to that of the evaporator. Modifications to the two variable terms with empirical coefficients c4 and c5 of AUcd are made to account for the effects of the airflow on the airside heat-transfer coefficient and the refrigerant mass flow on the condensing refrigerant heat-transfer coefficient [17]. AUcd "cd ¼ 1 exp Va a Cpa AUcd ¼
1 0:8 þ c c4 V0:5 þ c 5 mr 6 a
ð19Þ
ð20Þ
The power input to the operating condenser fans, Ecf, is calculated by: Ecf ¼
Va Ecf;tot Va;tot
ð21Þ
The required heat rejection, Qcd, for a certain cooling capacity is readily calculated upon the compressor power input is evaluated from the compressor model. The calculated Qcd can then be used to find out the required condenser airflow, Va, leaving condenser air temperature, Tcdal from Eqs. (17) and (18), the number of operating condenser fans and hence the fan power, Ecf, in Eq. (21).
3. Method of study 3.1. Parameter evaluation Having established the equations for the chiller components, Fig. 2 shows the flow chart for parameter evaluation. The cooling capacity, Qcl, the entering condenser’s air temperature, Tcdae, and the condensing temperature, Tcd, constitute the independent input parameters to the model components. The input parameters of Tchws and mw are regarded as constant for simplifying the chiller dynamics. Since the energybalance equations in the condenser model are implicit functions, the parameters Tcd, Pcd and Va have to be solved by an iterative approach. The main output parameters are the operating status of compressors, Ncc, condenser fans, Ncf, and their corresponding power consumption’s, Ecc and Ecf, for calculating the chiller operating efficiency in terms of kW/refrigeration ton, as given in the following section. 3.2. Operating efficiency Electricity consumption of an air-cooled chiller, Ech, is given by: Ech ¼ Ecc þ Ecf
ð22Þ
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The chiller operating efficiency can be expressed as: kW=refr:ton ¼
Ech Qcl
ð23Þ
It is desirable to maintain the chiller efficiency at a low value under the entire weather-load conditions. The chiller efficiency curve, defined as a plot between the kW/refrigeration ton and the chiller part load ratio at a particular outdoor temperature, can be used for realizing the model and assessing the cooling energy. 3.3. Head pressure (H-P) control Under the head pressure control, the condensing pressure is principally set at 1943 kPa corresponding to the condensing temperature of 50 C based on the design condition of 33 C outdoor temperature. However, some variation of the condensing pressure and temperature states will occur while equating the energy equations at the condenser side due to the step change in the condenser airflow and the interdependence of the condensing temperatures with the leaving and entering condenser air temperatures. Readjustment to the condensing refrigerant states is then made throughout the capacity control range in the parameter evaluation. The required condenser airflow is first obtained by predicting the number of operating condenser fans, so as to meet the required heat rejection and to maintain the condensing refrigerant settings. The condensing pressure and temperature are then rebalanced by the associated energy equations using an iterative procedure. It is expected that more condenser fans will be cycled off at decreasing outdoor temperatures during the mild weather. 3.4. Condensing temperature (C-T) control The condensing temperature and the condensing pressure are controlled parameters instead of controllable parameters as under the head pressure control. The maximum value of the input parameters Ncf and Va are determined, subject to the lowest condensing temperature limit of 27 C for proper compressor lubrication [18], for the evaluation of the output parameters Tcd and Pcd. The output parameters are solved iteratively by substituting back to the energy equations in the condenser model. Under the local weather conditions, the condensing temperatures can be varied from 28 to 50 C and the condensing pressures are controlled from 1131 to 1943 kPa accordingly.
4. Discussion of results 4.1. Model validation Model validation can be accomplished by calibrating the empirical model equations based on a comparison of simulation results with the manufacturer specification and
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the operating data of existing chiller plants. The manufacturers usually specify the chiller efficiency in terms of Energy Efficiency Ratio (EER) or COP according to the specific weather-load condition stipulated in the ARI Standard 550/590 [19]. The insitu chiller efficiency was evaluated from the operating data taken from two chiller plants in two local hotels, namely Hotel A and Hotel B. The two chiller plants consist of four chiller-pump pairs, with each chiller having the capacity of 185 and 199 tons of refrigeration (650 and 700 kW) for the Hotels A and B respectively. The size and steps of capacity control of individual chillers in both chiller plants are comparable to the modelled chiller. Fig. 3(a) shows the efficiency curves of the modelled chiller without adjustment to the modelling equations, and of the existing chiller plants at outdoor temperature of 33 C together with the manufacturer efficiency curve rated at the ARI Standard. Different characteristics of chiller efficiency were observed from the three scenarios. The manufacturer’s curve demonstrates an improved efficiency when the chiller part load decreases, which obviously cannot represent the actual chiller operation under the head pressure control. Using the manufacturer curve to determine the chiller combinations for the multiple-chiller systems would give rise to overstated chiller efficiency at part-load conditions and oversized equipment. Basically, the manufacturers state the chiller part-load efficiency for the integrated part-load value (IPLV) based on the weather-load conditions for a single chiller under the ARI Standard [19]. Such weather-load conditions may not reflect the actual load conditions by individual chillers in the multiple-chiller plant under the local weather. The efficiency curves by the simulation and the in-situ data exhibit a similar characteristic that the chiller efficiency is depressing at decreasing part loads, but the degree of depression is larger in the existing chillers. This probably associates with energy losses imposed on the compressors leading to extra compressor consumption at small part-load ratios. Apart from the deviation in the chiller efficiency at partload ratios below 0.8, the model can replicate the chiller efficiency of the in-situ chillers. This finding supports the need to readjust the model equation for the compressor efficiency to account for the extra energy loss at small part-load ratios.
Fig. 3. Chiller efficiency curves at part load.
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A derating factor, Fcc relating the chiller part-load ratio, PLR in the form of Eq. (24) is therefore introduced to the Eq. (14) to calibrate the compressor efficiency. Eq. (14) then becomes: cc ¼ 1:1v Fcc Fcc ¼ a2 PLR 2 þ a1 PLR þ a0
ð14aÞ ð24Þ
where the constants a0, a1 and a2 are determined to be 0.32, 0.14 and 0.55 respectively characterizing the energy losses by the compressor of the existing chillers. Fig. 3(b) shows that the efficiency curve of the calibrated model is adequate to represent the in-situ chillers. An analysis of the chiller efficiency and the operating parameters influencing the energy consumption within the chiller components under the condensing temperature control can then be carried out with the calibrated model. 4.2. Chiller efficiency and parametric analysis A set of chiller-efficiency curves for the head pressure and condensing temperature controls at outdoor temperatures corresponding to ARI standard rating is shown in Fig. 4. The maximum chiller efficiency occurs at full load condition irrespective of the outdoor temperature under both control methods. This suggests that the lead chillers in the multiple-chiller plants should be fully loaded before putting the lag chiller into operation for maximum efficiency. In assessing the chiller efficiency, a typical efficiency curve can be used under the head pressure control, whereas under the condensing temperature control, a family of efficiency curves at different outdoor temperatures should be employed. Under the condensing temperature control, the chiller efficiency can be improved by reducing the condensing temperatures at
Fig. 4. Chiller efficiency curves under head pressure (H-P) control and condensing temperature (C-T) controls. (Tcdae: entering condenser air temperature).
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Fig. 5. Number of operating compressors under head pressure and condensing temperature controls.
low outdoor temperatures as indicated by the downward shift of the efficiency curves, and the improvement is more significant at low-load conditions. Based on the local outdoor conditions, giving the entering condenser air-temperature from 33 to 12.8 C, the efficiency of the modelled chiller varies widely from 1.08 to 3.37 kW/refrigeration ton under the head pressure control. When using the condensing temperature control, the efficiency is significantly improved and varies in the range from 0.68 to 2.95 kW/refrigeration ton. It is possible, in a plant with multiple chillers, to obtain the chiller efficiency at above the minimum COP requirement of 2.7 (i.e., less than 1.3 kW/refrigeration ton) for most of the operating conditions. However, the minimum COP requirement cannot be complied with when the chillers operate under the head pressure control. This is also observed in the recorded data of the existing chillers in hotels A and B, where the COP is below 2.7 most of the time. In Fig. 5, the number of operating compressors is compared between the head pressure control and the condensing temperature control at various part-load ratios. There are six compressors in the modelled chiller. The number of operating compressors can be reduced by one under the condensing temperature control when producing the same cooling load as under the head pressure control. Hence, by adopting condensing-temperature control, the compressor power will be reduced. Fig. 6 shows the operating modes of the condenser fans at different weather-load conditions. The number of operating condenser fans increases with the outdoor temperature, irrespective of the part-load ratios. Under the head pressure control, all the condenser fans are running only at the design load and outdoor condition, whereas fewer fans will be energized at other conditions because smaller airflows can satisfy the heat-rejection load. For the condensing-temperature control, more condenser fans will be operated to induce larger airflow for suppressing the condensing pressure. Maximum number of fans may be switched on at both full-load and partload conditions. More fan power is consumed under this circumstance. Nevertheless, such power imposition can be compensated by the energy saving from the
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Fig. 6. Required number of condenser fans at different outdoor temperatures and part-load ratios.
Fig. 7. Energy consumption components by compressors, Ecc and condenser fans, Ecf under head pressure (H-P) and condensing-temperature (C-T) controls.
reduction in the number of operating compressors and the smaller compressor work input at lower condensing temperatures. A significant reduction in the compressor power is observed under the condensingtemperature control compared to the conventional head pressure control, as indicated by the trend lines in Fig. 7. The energy saving from the compressors outweighs the additional energy for running more condenser fans throughout the loading range. Substantial net energy saving is expected as the rated total condenser fan consumption accounts for only about one-tenth of the rated chiller consumption. Having identified the operating modes of condenser fans for both control methods, it is necessary to investigate the effect of such modes on the condensing temperatures under the varying outdoor temperature conditions. The results are shown in Fig. 8. At the peak outdoor temperature of 33 C, when using condensing temperature control, the condensing temperature can drop to 40 C at part-load ratio of
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Fig. 8. Changes of the condensing temperature with varying outdoor temperature and part-load ratio.
0.33 by cycling on all condensing fans. At below-design outdoor conditions, the condensing temperature can be as low as 26 C, which indeed is limited by the required temperature for proper lubrication of the compressors. Under the conventional head pressure control, there are also some changes in the condensing temperature due to heat and refrigerant flow balance among the evaporator, compressor and condenser. It fluctuates between 50 and 44 C throughout the entire loading range at outdoor temperature above 20 C. Comparing the two control methods, it is readily noticed that for most of the time the chiller plant under condensing temperature control operates at much lower condensing-temperatures and with fewer compressors. 4.3. Energy saving by applying condensing-temperature control For a multiple-chiller plant with equally sized chillers, the cooling load will be evenly handled by the operating chillers. To determine the number of operating chillers, it is necessary to establish the weather-load profile that correlates the chiller part load ratio to the outdoor temperatures based on the anticipated building cooling load and the local weather conditions. The profile can be used to establish the optimum combination of chillers and pumps at part load operation for maximizing the chiller efficiency and minimizing the overall plant consumption. The operating mode of the condenser fans shown in Fig. 8(b) is converted into a control algorithm to simulate the annual chiller consumption using the condensing temperature control. This is then applied to the two existing chiller plants in hotel A and B. The potential annual energy savings by the two chiller plants under the condensing-temperature control can be predicted by inputting their actual weather-load data to the simulation model. Fig. 9 shows the chiller efficiency against the part-load ratio from the recorded operating data and under the condensing temperature control for the chiller plant in hotel A. Comparing to the high and increasing kW/refrigeration ton at decreasing part-load ratio in the historical operations, the new condensing temperature control is able to maintain the chiller consumption below 2 kW/refrigeration
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Fig. 9. Efficiency of the chiller plant in Hotel A from historical data and under the new condensing temperature (C-T) control.
ton throughout the whole range of operating conditions, giving an average efficiency of 1.08 kW/refrigeration ton. Accordingly, 18.2% of the overall annual chiller consumption can be saved in hotel A. For the chiller plant in hotel B, there is a similar improvement in the chiller efficiency and the potential saving of annual chiller consumption is 29%.
5. Conclusions The simulation model of an air-cooled reciprocating chiller with a conventional head pressure control and the new condensing-temperature control for energy efficiency, is presented. The model equations for the compressor efficiency are calibrated by the actual chiller data to reflect the in-situ chiller efficiency at part load conditions. Under the condensing temperature control, a significant reduction in the compressor power can be achieved compared with the conventional head pressure control. The energy saving is due to the drop of the condensing temperature at belowdesign conditions leading to a smaller compression work input, reduced number of operating compressors and better chiller efficiency. This saving outweighs the energy imposition due to the cycling on of more condenser fans. The condensing temperature can be as low as 26 C, but any further drop is restricted by the requirement for compressor lubrication. Potential energy savings of 18.2 and 29% in the annual total chiller consumption have been identified for two existing chiller plants in hotels by applying the condensing temperature control. Based on the established operating modes of the condenser fans and compressors, the chiller consumption can be maintained at below 2 kW/refrigeration ton throughout the entire range of weather-load conditions, and the average efficiency is 1.08 kW/refrigeration ton.
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These results support the need to develop the condensing-temperature control as an alternative to the conventional head pressure control for improving the efficiency of air-cooled reciprocating chillers. Having investigated the operating mode of condenser fans in relation to the outdoor temperature and the part-load ratio, it is possible to establish the control algorithm for a multiple-chiller plant with microprocessor control as the cooling load can be determined by measuring the differential enthalpy of the supply and return chilled-water. Further experimental and simulation studies have to be carried out to realize the condensing-temperature setting based on the local weather, taking into account the chiller dynamics and the constraints of the chiller components balance together with the controllability of the operating parameters. Then the long-term energy saving by applying the condensingtemperature control, coupled with the optimum chiller and pump combinations, can be quantified.
Acknowledgements This study is supported by the Research Grant Council of the Hong Kong SAR and the central research grant of the Hong Kong Polytechnic University.
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[14] Wang SK. Handbook of air conditioning and refrigeration. New York: McGraw-Hill; 2001. [15] Tassou SA, Ge YT, Datta D, Chan KY. Modelling and control of multi-compressor food retail refrigeration systems. IMechE Seminar Publication 1998;15:53–66. [16] Dossat RJ. Principles of refrigeration. Upper Saddle River, NJ: Prentice Hall; 1996. [17] Kempiak MJ, Crawford RR. Three-zone, steady-state modeling of a mobile air-conditioning condenser. ASHRAE Transactions 1992;98(1):475–88. [18] Roper MA. Energy efficient chiller control. BSRIA Technical Note 16/2000, 2000. [19] ARI Standard 550/590-Water chilling packages using the vapor compression cycle. Arlington, Virginia: Air-Conditioning and Refrigeration Institute; 1998.
Applying condensing-temperature control in air-cooled reciprocating water chillers for energy efficiency K.T. Chan*, F.W. Yu Department of Building Services Engineering, The Hong Kong Polytechnic University, Hong Kong, China Received 1 May 2002; received in revised form 17 June 2002; accepted 22 June 2002
Abstract This paper reports on the modelling and findings of the energy performance of an aircooled reciprocating multiple-chiller plant under the conventional head pressure control and the new condensing-temperature control in a subtropical climate. The simulation model was validated using the operating data of an existing chiller plant. As noted from this existing aircooled reciprocating chiller plant, there was a substantial efficiency drop at part-load resulting from the head pressure control. If operating at variable lower condensing-temperatures based on the established operating mode of the condenser fans and compressors, it is shown that the chiller consumption can be maintained below 2 kW/refrigeration ton throughout the entire range of outdoor temperature and part-load conditions, giving an average efficiency of 1.08 kW/refrigeration ton. The energy imposition due to cycling on more condenser fans can be compensated by the reduced compressor consumption. Potential energy savings of 18.2 and 29% in the annual chiller consumption are achievable by applying the condensing-temperature control to two existing chiller plants studied. This supports the need to develop the condensing-temperature control as an improvement to the conventional head pressure control. # 2002 Elsevier Science Ltd. All rights reserved. Keywords: Air-cooled chillers; Condensing temperature control; Head pressure control; Energy efficiency
1. Introduction Air-cooled reciprocating chillers have been widely installed in local commercial buildings for providing space cooling and dehumidification. The operations of chiller plants have a significant impact on electricity consumption and maximum power * Corresponding author. Fax: +852-27746146. E-mail address: [email protected] (K.T. Chan). 0306-2619/02/$ - see front matter # 2002 Elsevier Science Ltd. All rights reserved. PII: S0306-2619(02)00053-3
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Nomenclature Symbols AU overall heat-transfer coefficient (kW K-1) CR compression ratio Cpa specific heat-capacity of air (kJ kg1 K-1) Cpw specific heat-capacity of water (kJ kg1 K-1) Ecc power input to operating compressors (kW) Ecf power input to operating condenser fans (kW) Ech overall power input to a chiller (kW) Fcc derating factor for the compressor efficiency hi specific enthalpy of refrigerant at state i kW/refr. ton chiller operating efficiency in kW per refrigeration ton mr refrigerant mass flow rate per compressor (kg s1) mw chilled water mass flow rate (kg s1) n index of reversible polytropic expansion process Ncc number of operating compressors P saturated pressure of refrigeration circuit (kPa) PLR chiller part-load ratio (given by Qcl/Qcr) Qcd total heat rejection (kW) Qcl required chiller cooling output (kW or tons of refrigeration) Qcr rated chiller cooling capacity (kW or tons of refrigeration) qrf refrigeration effect (kJ kg1) T saturated temperature of refrigeration ( C) Tcdae entering condenser air temperature or outdoor air temperature ( C) Tcdal leaving condenser air temperature ( C) Tchwr return chilled water temperature ( C) Tchws supply chilled water temperature ( C) Va airflow provided by operating condenser fans (m3 s1) Vp piston displacement of each compressor (m3 s1) vr refrigerant specific volume at compressor suction pressure (m3 kg1) win isentropic work input to the compressor (kJ kg1) " heat-exchange effectiveness isen isentropic efficiency cc compressor efficiency v volumetric efficiency of a reciprocating compressor a air density (kg m3)
Subscripts cd ev tot
condenser evaporator total
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demand. The chiller coefficient of performance (COP) of two existing chiller plants was less than 2.5 most of the time [1], which did not meet with the minimum COP of 2.7 stipulated in the local energy code for energy efficiency of air-conditioning installations [2]. This poor chiller efficiency was associated with oversized equipment [3], inappropriate matching of operating chillers with the cooling-load demand, and the chillers always worked at a high condensing-temperature based on the design outdoor condition of 33 C (often referred to as head pressure control) [1]. The local weather data [4] indicated that the outdoor temperatures are below 25 C half of the time in a year. There is a considerable scope for lowering the condensing temperature and thereby improving the chiller efficiency [1,5]. An experimental study by Smith and King [5] on the reduction in condenser pressure lift for a reciprocating chiller rated at 35 kW cooling capacity illustrated that a 10% reduction in the overall chiller consumption was achieved by driving the condenser fan harder at all outdoor temperatures below 25 C. Nevertheless, no detailed algorithms have been reported for variable condensing-temperature. The trade-off between increased condenser fan power and decreased compressor power, and the long-term energy saving by lowering the condensing temperature based on the weather and part load conditions has not been assessed. Many simulation models for chiller systems were documented [6–12]. There were a few models developed for air-cooled reciprocating chillers: among those was a reciprocating chiller model by Leverenz and Bergan [10] for hourly energy analysis using a ‘‘black box’’ concept. Such a concept directly relates the input and output parameters of the chiller by a set of empirical equations without considering the operating balance among the components of the evaporator, compressor and condenser. Moreover, an optimisation study cannot be carried out using the ‘‘black box’’ models, as both the controllable and controlled variables are absent. Other reciprocating chiller models deal with a set of energy-balance equations without considering the capacity control of the reciprocating compressors influencing the power input [11]. A mechanistic model taking into account the thermodynamics and operating balance within the components of an air-cooled reciprocating chiller is developed here for the study of the chiller performance under the conventional head pressure control and the condensing temperature control. Necessary calibration of the model is made according to the manufacturer’s specification and the operating data from two existing chillerplants. This paper presents the validated chiller model, the chiller’s operating parameters and the analysis of chiller efficiency. Energy saving potential by lowering the condensing temperature at part load conditions for the two existing chiller plants is assessed.
2. Model description 2.1. Characteristics of the modelled chiller An air-cooled reciprocating chiller, using refrigerant R-22 (chlorodifluoromethane) with rated capacity of 199 tons of refrigeration (i.e. 700 kW, 1 refrigeration ton=3.52 kW), was studied. The modelled chiller is commonly used in multiple-chiller
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Fig. 1. Refrigeration cycle of the modelled chiller.
systems for small to medium sized buildings requiring cooling capacity in the range of 300–1000 tons of refrigeration (1056–3520 kW). In general, individual chillers with different or identical sizes in a multiple-chiller plant operate in parallel to meet the varying cooling load demand. The chiller has six steps of capacity control from full load down to a minimum of 16.7% by staging the six compressors, and contains two refrigerant circuits to enhance the reliability and standby capacity. In order to meet the required heat rejection under the head pressure control, twelve steps of control on the heat rejection airflow rate are achieved by cycling the twelve condenser fans. 2.2. Refrigeration cycle and assumptions for the model Fig. 1 shows the refrigeration cycle for the modelled chiller. In the refrigeration cycle, heat loss or gain from the surroundings is ignored. Pressure drops in the refrigerant pipelines are neglected. It is assumed that the refrigerant throttling process through the expansion valve is adiabatic. The compression process is polytropic and the actual compression work is the isentropic compression work input, win divided by the isentropic efficiency, isen. The degrees of sub-cooling and superheat are maintained at 6.5 and 5 C respectively under all loading conditions. 2.3. Evaporator model The chiller modelled contains a shell and tube evaporator. Energy balance for the heat flow between the refrigerant and water at the evaporator can be expressed as: Qcl ¼ mw Cpw ðTchwr Tchws Þ
ð1Þ
¼ Ncc mr qrf
ð2Þ
¼ mw Cpw "ev ðTchwr Tev Þ
ð3Þ
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The mass flow rate of chilled water, mw and the supply chilled-water temperature, Tchws are maintained at 27.4 kg s1 and 7 C respectively based on the manufacturer’s recommendation. The return chilled-water temperature, Tchwr will vary with the cooling capacity, Qcl in the range of 7.9–12.5 C. Referring to Fig. 1, the refrigeration effect, qrf in Eq. (2) is given by: qrf ¼ h1 h4
ð4Þ
In determining the evaporating temperature, Tev, the classical heat-exchange effectiveness method is applied using Eq. (3). The heat-exchange effectiveness at the
Fig. 2. Flow chart for parameter evaluation.
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waterside, "ev is determined by Eq. (5). The overall heat-transfer coefficient, AUev is described by the empirical relation in Eq. (6) [12], where c1, c2 and c3 are constants evaluated using the chiller’s performance data. AUev "ev ¼ 1 exp ð5Þ mw Cpw AUev ¼
c1 m0:8 w
1 þ c2 Q0:745 þ c3 cl
ð6Þ
For a constant supply chilled-water temperature, Tchws, the evaporating temperature, Tev varies from 5.3 to 5.8 C as calculated by Eq. (6) based on the anticipated range of return chilled-water temperature. The evaporating pressure, Pev and hence the specific enthalpy of refrigerant at compressor suction state, h1 can be determined using the state equations of the refrigerant R-22 [13]. The variation of the refrigeration effect, qrf, is small under the head pressure control but is substantial under the condensing temperature control with varying condensing temperatures. The refrigeration effect increases with the reduction in the condensing temperature at a constant evaporating temperature. For a certain cooling capacity, the number of operating compressors, Ncc, can be determined by Eq. (2). The refrigerant mass flow rate is calculated by Eq. (8) and the refrigeration effect is evaluated from the refrigerant’s state equations. 2.4. Compressor model The compressor power input, refrigerant mass flow rate per compressor at the compression process and the volumetric efficiency [14] are given by Eqs. (7)–(9): Ecc ¼ Ncc mr
mr ¼
win isen cc
Vp v vr
v ¼ 0:94 0:015 CR
ð7Þ
ð8Þ ð9Þ
where CR is the compression ratio: CR ¼
Pcd Pev
ð10Þ
The piston displacement of a compressor, Vp, in Eq. (8) relating to the cylinder diameter, stroke of the piston and number of cylinder per compressor, is regarded as a constant for a constant-speed compressor. It can be evaluated using the manufacturer’s data at the full-load condition. The specific volume of refrigerant at the
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suction of the compressor, vr was determined from the evaporating pressure and the degree of superheat using the refrigerant state equations. The compression ratio varies with the condensing pressure lift to a small extent under the head pressure control, but it becomes a dependent variable changing with the condensing pressure under the condensing-temperature control. Based on the refrigeration cycle in Fig. 1, the isentropic work input by the compressor, win in Eq. (7), is given by: n n- 1 CR n 1 win ¼ Pev vr ð11Þ n1 The isentropic compression index, n, is taken to be 1.2 for the semi-hermetic compressor motors by refrigerant cooling. The specific enthalpy of the refrigerant at state 2 can be evaluated using Eq. (12): h2 ¼ h1 þ
win isen
ð12Þ
where the isentropic efficiency, isen, can be calculated using the empirical equation in relation to the compression ratio, CR, [15]: isen ¼ 0:73 þ 0:1 CR 0:026 CR 2 þ 0:0017 CR 3
ð13Þ
Since there is no mathematical relation governing the compressor’s efficiency, cc, in the first trial, it is approximated by the volumetric efficiency with an additional 10% to offset the power loss due to the mechanical friction in the compressor [16]: cc ¼ 1:1 v
ð14Þ
Verification of the approximation is carried out subsequently by comparing the simulation result with the operating data. 2.5. Condenser model The chiller is installed with an air-cooled condenser of finned coils integrated with sub-cooling coils. The total heat rejection, Qcd, is the sum of refrigeration load, Qcl, and compressor power input, Ecc. It can be calculated by the following energy balance equations: Qcd ¼ Qcl þ Ecc
ð15Þ
¼ Ncc mr ðh2 h3 Þ
ð16Þ
¼ Va a Cpa ðTcdal Tcdae Þ
ð17Þ
¼ Va a Cpa "cd ðTcd Tcdae Þ
ð18Þ
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The evaluation of the condenser’s effectiveness, "cd, together with AUcd is analogous to that of the evaporator. Modifications to the two variable terms with empirical coefficients c4 and c5 of AUcd are made to account for the effects of the airflow on the airside heat-transfer coefficient and the refrigerant mass flow on the condensing refrigerant heat-transfer coefficient [17]. AUcd "cd ¼ 1 exp Va a Cpa AUcd ¼
1 0:8 þ c c4 V0:5 þ c 5 mr 6 a
ð19Þ
ð20Þ
The power input to the operating condenser fans, Ecf, is calculated by: Ecf ¼
Va Ecf;tot Va;tot
ð21Þ
The required heat rejection, Qcd, for a certain cooling capacity is readily calculated upon the compressor power input is evaluated from the compressor model. The calculated Qcd can then be used to find out the required condenser airflow, Va, leaving condenser air temperature, Tcdal from Eqs. (17) and (18), the number of operating condenser fans and hence the fan power, Ecf, in Eq. (21).
3. Method of study 3.1. Parameter evaluation Having established the equations for the chiller components, Fig. 2 shows the flow chart for parameter evaluation. The cooling capacity, Qcl, the entering condenser’s air temperature, Tcdae, and the condensing temperature, Tcd, constitute the independent input parameters to the model components. The input parameters of Tchws and mw are regarded as constant for simplifying the chiller dynamics. Since the energybalance equations in the condenser model are implicit functions, the parameters Tcd, Pcd and Va have to be solved by an iterative approach. The main output parameters are the operating status of compressors, Ncc, condenser fans, Ncf, and their corresponding power consumption’s, Ecc and Ecf, for calculating the chiller operating efficiency in terms of kW/refrigeration ton, as given in the following section. 3.2. Operating efficiency Electricity consumption of an air-cooled chiller, Ech, is given by: Ech ¼ Ecc þ Ecf
ð22Þ
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The chiller operating efficiency can be expressed as: kW=refr:ton ¼
Ech Qcl
ð23Þ
It is desirable to maintain the chiller efficiency at a low value under the entire weather-load conditions. The chiller efficiency curve, defined as a plot between the kW/refrigeration ton and the chiller part load ratio at a particular outdoor temperature, can be used for realizing the model and assessing the cooling energy. 3.3. Head pressure (H-P) control Under the head pressure control, the condensing pressure is principally set at 1943 kPa corresponding to the condensing temperature of 50 C based on the design condition of 33 C outdoor temperature. However, some variation of the condensing pressure and temperature states will occur while equating the energy equations at the condenser side due to the step change in the condenser airflow and the interdependence of the condensing temperatures with the leaving and entering condenser air temperatures. Readjustment to the condensing refrigerant states is then made throughout the capacity control range in the parameter evaluation. The required condenser airflow is first obtained by predicting the number of operating condenser fans, so as to meet the required heat rejection and to maintain the condensing refrigerant settings. The condensing pressure and temperature are then rebalanced by the associated energy equations using an iterative procedure. It is expected that more condenser fans will be cycled off at decreasing outdoor temperatures during the mild weather. 3.4. Condensing temperature (C-T) control The condensing temperature and the condensing pressure are controlled parameters instead of controllable parameters as under the head pressure control. The maximum value of the input parameters Ncf and Va are determined, subject to the lowest condensing temperature limit of 27 C for proper compressor lubrication [18], for the evaluation of the output parameters Tcd and Pcd. The output parameters are solved iteratively by substituting back to the energy equations in the condenser model. Under the local weather conditions, the condensing temperatures can be varied from 28 to 50 C and the condensing pressures are controlled from 1131 to 1943 kPa accordingly.
4. Discussion of results 4.1. Model validation Model validation can be accomplished by calibrating the empirical model equations based on a comparison of simulation results with the manufacturer specification and
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the operating data of existing chiller plants. The manufacturers usually specify the chiller efficiency in terms of Energy Efficiency Ratio (EER) or COP according to the specific weather-load condition stipulated in the ARI Standard 550/590 [19]. The insitu chiller efficiency was evaluated from the operating data taken from two chiller plants in two local hotels, namely Hotel A and Hotel B. The two chiller plants consist of four chiller-pump pairs, with each chiller having the capacity of 185 and 199 tons of refrigeration (650 and 700 kW) for the Hotels A and B respectively. The size and steps of capacity control of individual chillers in both chiller plants are comparable to the modelled chiller. Fig. 3(a) shows the efficiency curves of the modelled chiller without adjustment to the modelling equations, and of the existing chiller plants at outdoor temperature of 33 C together with the manufacturer efficiency curve rated at the ARI Standard. Different characteristics of chiller efficiency were observed from the three scenarios. The manufacturer’s curve demonstrates an improved efficiency when the chiller part load decreases, which obviously cannot represent the actual chiller operation under the head pressure control. Using the manufacturer curve to determine the chiller combinations for the multiple-chiller systems would give rise to overstated chiller efficiency at part-load conditions and oversized equipment. Basically, the manufacturers state the chiller part-load efficiency for the integrated part-load value (IPLV) based on the weather-load conditions for a single chiller under the ARI Standard [19]. Such weather-load conditions may not reflect the actual load conditions by individual chillers in the multiple-chiller plant under the local weather. The efficiency curves by the simulation and the in-situ data exhibit a similar characteristic that the chiller efficiency is depressing at decreasing part loads, but the degree of depression is larger in the existing chillers. This probably associates with energy losses imposed on the compressors leading to extra compressor consumption at small part-load ratios. Apart from the deviation in the chiller efficiency at partload ratios below 0.8, the model can replicate the chiller efficiency of the in-situ chillers. This finding supports the need to readjust the model equation for the compressor efficiency to account for the extra energy loss at small part-load ratios.
Fig. 3. Chiller efficiency curves at part load.
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A derating factor, Fcc relating the chiller part-load ratio, PLR in the form of Eq. (24) is therefore introduced to the Eq. (14) to calibrate the compressor efficiency. Eq. (14) then becomes: cc ¼ 1:1v Fcc Fcc ¼ a2 PLR 2 þ a1 PLR þ a0
ð14aÞ ð24Þ
where the constants a0, a1 and a2 are determined to be 0.32, 0.14 and 0.55 respectively characterizing the energy losses by the compressor of the existing chillers. Fig. 3(b) shows that the efficiency curve of the calibrated model is adequate to represent the in-situ chillers. An analysis of the chiller efficiency and the operating parameters influencing the energy consumption within the chiller components under the condensing temperature control can then be carried out with the calibrated model. 4.2. Chiller efficiency and parametric analysis A set of chiller-efficiency curves for the head pressure and condensing temperature controls at outdoor temperatures corresponding to ARI standard rating is shown in Fig. 4. The maximum chiller efficiency occurs at full load condition irrespective of the outdoor temperature under both control methods. This suggests that the lead chillers in the multiple-chiller plants should be fully loaded before putting the lag chiller into operation for maximum efficiency. In assessing the chiller efficiency, a typical efficiency curve can be used under the head pressure control, whereas under the condensing temperature control, a family of efficiency curves at different outdoor temperatures should be employed. Under the condensing temperature control, the chiller efficiency can be improved by reducing the condensing temperatures at
Fig. 4. Chiller efficiency curves under head pressure (H-P) control and condensing temperature (C-T) controls. (Tcdae: entering condenser air temperature).
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Fig. 5. Number of operating compressors under head pressure and condensing temperature controls.
low outdoor temperatures as indicated by the downward shift of the efficiency curves, and the improvement is more significant at low-load conditions. Based on the local outdoor conditions, giving the entering condenser air-temperature from 33 to 12.8 C, the efficiency of the modelled chiller varies widely from 1.08 to 3.37 kW/refrigeration ton under the head pressure control. When using the condensing temperature control, the efficiency is significantly improved and varies in the range from 0.68 to 2.95 kW/refrigeration ton. It is possible, in a plant with multiple chillers, to obtain the chiller efficiency at above the minimum COP requirement of 2.7 (i.e., less than 1.3 kW/refrigeration ton) for most of the operating conditions. However, the minimum COP requirement cannot be complied with when the chillers operate under the head pressure control. This is also observed in the recorded data of the existing chillers in hotels A and B, where the COP is below 2.7 most of the time. In Fig. 5, the number of operating compressors is compared between the head pressure control and the condensing temperature control at various part-load ratios. There are six compressors in the modelled chiller. The number of operating compressors can be reduced by one under the condensing temperature control when producing the same cooling load as under the head pressure control. Hence, by adopting condensing-temperature control, the compressor power will be reduced. Fig. 6 shows the operating modes of the condenser fans at different weather-load conditions. The number of operating condenser fans increases with the outdoor temperature, irrespective of the part-load ratios. Under the head pressure control, all the condenser fans are running only at the design load and outdoor condition, whereas fewer fans will be energized at other conditions because smaller airflows can satisfy the heat-rejection load. For the condensing-temperature control, more condenser fans will be operated to induce larger airflow for suppressing the condensing pressure. Maximum number of fans may be switched on at both full-load and partload conditions. More fan power is consumed under this circumstance. Nevertheless, such power imposition can be compensated by the energy saving from the
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Fig. 6. Required number of condenser fans at different outdoor temperatures and part-load ratios.
Fig. 7. Energy consumption components by compressors, Ecc and condenser fans, Ecf under head pressure (H-P) and condensing-temperature (C-T) controls.
reduction in the number of operating compressors and the smaller compressor work input at lower condensing temperatures. A significant reduction in the compressor power is observed under the condensingtemperature control compared to the conventional head pressure control, as indicated by the trend lines in Fig. 7. The energy saving from the compressors outweighs the additional energy for running more condenser fans throughout the loading range. Substantial net energy saving is expected as the rated total condenser fan consumption accounts for only about one-tenth of the rated chiller consumption. Having identified the operating modes of condenser fans for both control methods, it is necessary to investigate the effect of such modes on the condensing temperatures under the varying outdoor temperature conditions. The results are shown in Fig. 8. At the peak outdoor temperature of 33 C, when using condensing temperature control, the condensing temperature can drop to 40 C at part-load ratio of
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Fig. 8. Changes of the condensing temperature with varying outdoor temperature and part-load ratio.
0.33 by cycling on all condensing fans. At below-design outdoor conditions, the condensing temperature can be as low as 26 C, which indeed is limited by the required temperature for proper lubrication of the compressors. Under the conventional head pressure control, there are also some changes in the condensing temperature due to heat and refrigerant flow balance among the evaporator, compressor and condenser. It fluctuates between 50 and 44 C throughout the entire loading range at outdoor temperature above 20 C. Comparing the two control methods, it is readily noticed that for most of the time the chiller plant under condensing temperature control operates at much lower condensing-temperatures and with fewer compressors. 4.3. Energy saving by applying condensing-temperature control For a multiple-chiller plant with equally sized chillers, the cooling load will be evenly handled by the operating chillers. To determine the number of operating chillers, it is necessary to establish the weather-load profile that correlates the chiller part load ratio to the outdoor temperatures based on the anticipated building cooling load and the local weather conditions. The profile can be used to establish the optimum combination of chillers and pumps at part load operation for maximizing the chiller efficiency and minimizing the overall plant consumption. The operating mode of the condenser fans shown in Fig. 8(b) is converted into a control algorithm to simulate the annual chiller consumption using the condensing temperature control. This is then applied to the two existing chiller plants in hotel A and B. The potential annual energy savings by the two chiller plants under the condensing-temperature control can be predicted by inputting their actual weather-load data to the simulation model. Fig. 9 shows the chiller efficiency against the part-load ratio from the recorded operating data and under the condensing temperature control for the chiller plant in hotel A. Comparing to the high and increasing kW/refrigeration ton at decreasing part-load ratio in the historical operations, the new condensing temperature control is able to maintain the chiller consumption below 2 kW/refrigeration
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Fig. 9. Efficiency of the chiller plant in Hotel A from historical data and under the new condensing temperature (C-T) control.
ton throughout the whole range of operating conditions, giving an average efficiency of 1.08 kW/refrigeration ton. Accordingly, 18.2% of the overall annual chiller consumption can be saved in hotel A. For the chiller plant in hotel B, there is a similar improvement in the chiller efficiency and the potential saving of annual chiller consumption is 29%.
5. Conclusions The simulation model of an air-cooled reciprocating chiller with a conventional head pressure control and the new condensing-temperature control for energy efficiency, is presented. The model equations for the compressor efficiency are calibrated by the actual chiller data to reflect the in-situ chiller efficiency at part load conditions. Under the condensing temperature control, a significant reduction in the compressor power can be achieved compared with the conventional head pressure control. The energy saving is due to the drop of the condensing temperature at belowdesign conditions leading to a smaller compression work input, reduced number of operating compressors and better chiller efficiency. This saving outweighs the energy imposition due to the cycling on of more condenser fans. The condensing temperature can be as low as 26 C, but any further drop is restricted by the requirement for compressor lubrication. Potential energy savings of 18.2 and 29% in the annual total chiller consumption have been identified for two existing chiller plants in hotels by applying the condensing temperature control. Based on the established operating modes of the condenser fans and compressors, the chiller consumption can be maintained at below 2 kW/refrigeration ton throughout the entire range of weather-load conditions, and the average efficiency is 1.08 kW/refrigeration ton.
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These results support the need to develop the condensing-temperature control as an alternative to the conventional head pressure control for improving the efficiency of air-cooled reciprocating chillers. Having investigated the operating mode of condenser fans in relation to the outdoor temperature and the part-load ratio, it is possible to establish the control algorithm for a multiple-chiller plant with microprocessor control as the cooling load can be determined by measuring the differential enthalpy of the supply and return chilled-water. Further experimental and simulation studies have to be carried out to realize the condensing-temperature setting based on the local weather, taking into account the chiller dynamics and the constraints of the chiller components balance together with the controllability of the operating parameters. Then the long-term energy saving by applying the condensingtemperature control, coupled with the optimum chiller and pump combinations, can be quantified.
Acknowledgements This study is supported by the Research Grant Council of the Hong Kong SAR and the central research grant of the Hong Kong Polytechnic University.
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[14] Wang SK. Handbook of air conditioning and refrigeration. New York: McGraw-Hill; 2001. [15] Tassou SA, Ge YT, Datta D, Chan KY. Modelling and control of multi-compressor food retail refrigeration systems. IMechE Seminar Publication 1998;15:53–66. [16] Dossat RJ. Principles of refrigeration. Upper Saddle River, NJ: Prentice Hall; 1996. [17] Kempiak MJ, Crawford RR. Three-zone, steady-state modeling of a mobile air-conditioning condenser. ASHRAE Transactions 1992;98(1):475–88. [18] Roper MA. Energy efficient chiller control. BSRIA Technical Note 16/2000, 2000. [19] ARI Standard 550/590-Water chilling packages using the vapor compression cycle. Arlington, Virginia: Air-Conditioning and Refrigeration Institute; 1998.