Environmental performance and economic analysis of all-variable speed chiller systems with load-based speed control

Applied Thermal Engineering 29 (2009) 1721–1729

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Environmental performance and economic analysis of all-variable speed chiller systems with load-based speed control F.W. Yu *, K.T. Chan Department of Building Services Engineering, The Hong Kong Polytechnic University, Hung Hom, Kowloon, Hong Kong, China

a r t i c l e

i n f o

Article history: Received 30 November 2007 Accepted 5 August 2008 Available online 12 August 2008 Keywords: Water-cooled chiller Simulation Variable speed Electricity and water consumption

a b s t r a c t There are increasing views on implementing all-variable speed chiller plants in place of conventional constant speed plants. Supporters of these views claim that all-variable speed chiller systems can operate much more efficiently at part load in response to changes in building cooling load. This paper introduces load-based speed control for all-variable speed plants to optimize their environmental performance. Thermodynamic-behaviour chiller system models were developed to perform environmental assessment (in terms of annual electricity and water consumption) for typical constant speed and all-variable speed chiller systems operating for the cooling load profile of a local office building. Operating cost differences between the two systems were calculated and compared in an economic analysis. Applying load-based speed control to the variable speed chiller plant can decrease the annual total electricity use by 19.7% and annual water use by 15.9% relative to the corresponding constant speed plant. The significance of this study is to provide more insights into how to make chiller systems more sustainable. Ó 2008 Elsevier Ltd. All rights reserved.

1. Introduction Central chiller plants have long been used to provide cooling energy for comfort air conditioning at the expense of considerable electricity. Where water-cooled heat rejection systems are concerned, each plant contains multiple chillers, chilled water distribution pumps, condenser water pumps and cooling towers. For many existing chiller plants, variable speed applications are limited to the tower fans and secondary-loop chilled water pumps if the plant employs a primary/secondary pumping system. It is worth widening the use of variable speed drives (VSD) for all the components in order to make chiller systems operate as efficiently as possible. According to Crowther and Furlong’s study [1], applying VSD to cooling tower fans with optimized control could make a 5–8% improvement in system energy performance, relative to the equivalent system but with on/off fan control and a fixed condenser water temperature set point. They addressed challenges of varying the flow of condenser pumps during part load conditions and indicated that the increased condenser water temperature range under the reduced flow conditions could influence the chiller work, tower heat transfer effectiveness, and the system as a whole. There are some studies focusing on applying VSD individually to the chillers, chilled water pumps and condenser water pumps. Qureshi and Tassou [2] made a review on the use of VSD for chiller systems. They confirmed that VSD has been applied successfully to perform capacity modulation for chiller compressors. The VSD * Corresponding author. Tel.: +852 3746 0416; fax: +852 2364 7375. E-mail address: [email protected] (F.W. Yu). 1359-4311/$ - see front matter Ó 2008 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2008.08.003

chillers bring about superior part load efficiency compared with chillers using multiple constant speed compressors with cycling control. Some major manufacturers have launched VSD chiller products but many modern central cooling plants tend to be designed with a hybrid of variable and constant speed chillers. The idea of using all-variable speed chillers in a multiple-chiller plant is not fully implemented by traditional system designers. Bahnfleth and Peyer [3] conducted extensive research on the application of variable primary flow for chillers. They carried out parametric modelling studies on chilled water pumping system alternatives and found that the variable flow, primary-only pumping systems could reduce the total annual plant energy use by 2–5%, first cost by 4–8%, and life cycle cost by 3–5% relative to the equivalent primary/secondary systems in which the primary loop is served by constant speed pumps and the secondary loop by variable speed pumps. They also pointed out that more solid data are required to address some issues about control complexity and stability of chilled water supply temperature when using the variable flow, primary-only pumping systems. Taylor [4] stated that using variable speed chillers and/or variable flow, primary-only pumping systems is a viable means to eliminate significant degradation in system performance at part load operation while accommodating the low ‘‘delta-T” syndrome of chilled water circuits. Hydeman et al. [5] developed a modified DOE-2 chiller model which can be used to evaluate the performance of chillers with VSD or under varying condenser water flow conditions. Such a modelling tool would form a good basis for examining the potential benefits of chiller systems using VSD. Hydeman and Zhou [6] presented a parametric analysis technique to

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optimize the control sequence of chilled water plants using VSD for the chillers, condenser water pumps and cooling tower fans. Based on simulation results of an example chiller plant, they demonstrated that the speed of the condenser pumps and tower fans should be adjusted in response to the chiller load in order to achieve optimized plant performance. Gordon et al. [7] highlighted that the condenser water flow rate could be a control variable in improving the energy performance of chiller systems. They established an analytic semi-empirical chiller model to study variations of chiller COP (coefficient of performance) at different condenser water flow rates. No analysis was made on the system level involving the interaction between the compressor power, pump power and tower fan power. The model serves well for fault detection and diagnosis purposes, but is incapable of accounting the varying operating characteristics of cooling towers. No control regime was generalized on how the condenser water flow should be varied in response to various chiller load and wet-bulb temperature conditions in order to achieve optimal energy performance of the system. Hartman is a pioneer promoting all-variable speed chiller plants where all the chillers, condenser pumps and tower fans are driven by VSD [8]. He explained criteria for designing such plants, selecting the equipment and operating the equipment in order to achieve the highest overall plant performance at part load conditions. Based on his simulation analyses, the annual energy use of all-variable speed chiller plants with optimized controls would be on average 28% lower than the corresponding conventional constant speed plants with equipment of the same nominal efficiency at design conditions. He also developed the equal marginal performance principle (EMPP) for designing efficient air-conditioning systems [9]. The EMPP involves understanding the power relationships between the system components and formulating the powerbased speed control algorithms to operate the variable speed equipment. All the past studies have brought individual contributions to promoting VSD applications in chiller systems. Yet there is a lack of simple but generic approach to controlling all-variable speed chiller plants. More track records of using all-variable speed chiller systems with advanced control are considered necessary, considering the use of variable speed technology will continue to grow to sustain highly efficient chiller systems. The aim of this paper is to investigate how load-based speed control should be applied to all-variable speed chiller plants to enhance their environmental and economic benefits. This paper first describes an all-variable speed chiller plant and a conventional constant speed plant considered in the simulation study. Optimal regions of constant and variable speed chillers will be presented to illustrate how to determine optimum chiller sequencing. Based on the chiller system models developed, the annual electricity and water consumption will be predicted for the two systems operating for the cooling load profile of an office building. Operating cost differences between the two systems will be calculated and compared in an economic analysis. Discussion will be given on how the pumps and tower fans should operate based on the simple load-speed relationship in order to achieve near optimal system control. After that, some remarks will be made on using the EMPP and power-based speed control for system optimization. The significance of this study is to provide more solid data which helps widen the use of all-variable speed chiller systems for greener buildings. 2. Methods of study 2.1. Equipment and plant arrangement Figs. 1 and 2 show two systems with the conventional and allvariable speed configurations considered in the simulation study.

CT 1

CT2

CT 3

Cooling towers (CT) Heat rejection rate: 2480kW each

Chiller 1 (2130 kW) Chiller 2 (2130 kW) Condenser water pumps

Chiller 3 (2130 kW)

Primary chilled water pumps

Decoupling bypass VSD

Air side system coils with 2-way control valves

VSD Variable speed secondary chilled water pumps

Fig. 1. Schematic of the conventional system with constant speed configuration.

Regarding the conventional system, each modelled component is based on commonly used equipment types. Three identical constant speed centrifugal chillers were used, each of which had a full load COP of 6.1 and an integrated part load value (IPLV) of 7.0. The IPLV formula, as defined in the ARI standard 550/590 [10], is given by Eq. (1) and used for the part load performance rating of a single chiller. The weighting factors of 0.01, 0.42, 0.45 and 0.12 are based on the weighted average of the most common building types and climate conditions for 29 US cities, with and without airside economizers. Regarding the constant speed chillers studied, the COP is 6.1 at 100% full load, 7.1 at 75%, 7.4 at 50% and 5.5 at 25%.

IPLV ¼ 0:01A þ 0:42B þ 0:45C þ 0:12D

ð1Þ

where A, COP at 100% full load with 29.4 °C entering condenser water (ECWT); B, COP at 75% full load with 23.9 °C entering condenser water (ECWT); C, COP at 50% full load with 18.3 °C entering condenser water (ECWT); D = COP at 25% full load with 18.3 °C entering condenser water (ECWT). Each of the chillers operated with 7 °C leaving chilled water and 29.4 °C entering condenser water for all operating conditions. The flow rate per cooling capacity was designed at 0.043 l/s per kW for chilled water and 0.053 l/s per kW for condenser water, which is related to a temperature difference of 5.5 °C for both the chilled water and condenser water at peak load conditions. Constant speed pumps were considered for the primary chilled water loop and condenser water loop while variable speed pumps were used for the secondary chilled water loop. The chilled water distribution system contained a decoupling bypass pipe linking the primary and secondary chilled water loops to balance the flow under part load conditions.

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CT 1 VSD

CT 2 VSD

VSD

CT 3 VSD

Cooling towers (CT) Heat rejection rate: 2480 kW each

Chiller 1 VSD (2130 kW)

VSD

VSD VSD Condenser water pumps

Chiller 2 (2130 kW)

VSD

Chiller 3 (2130 kW)

VSD

VSD VSD Primary chilled water pumps

Low load bypass

Air side system coils with 2-way control valves

tion of chiller loads and wet-bulb temperatures, the chiller system model would search for optimal speeds of the operating chillers, pumps and tower fans in order to minimize the overall system power. Instead of controlling directly the chilled water and condenser water temperatures, they were allowed to float within their boundary conditions to achieve the optimal speed control with maximum energy performance. The possible change of the temperatures was identified from the simulation and will be presented in the results and discussion section. For both systems, the plant models decided the sequencing of chillers based on any given hourly building cooling load. The chillers carried the equal percentage load and each of which operated in parallel with one primary chilled water pump, one condenser water pump, and one cooling tower. The equal percentage load strategy is inevitable when chillers operate with the same temperature difference of chilled water and each of them carries their nominal chilled water flow delivered by constant speed pumps. It is possible to have equal or non-equal load fractioning between the chillers operating with variable speed pumps producing varying chilled water flows. Considering that variable speed chillers operate with maximum COP at part load regions instead of at full load, it is desirable to apply the equal percentage load strategy in order to ensure that each of the chillers (of the same size and performance) carried the modest load (lying within the optimal region) as far as possible. It should be noted that no stand-by equipment capacity is considered for the two systems and the total plant capacity matches with the peak building cooling load. Table 1 summarizes other details of the two systems. 2.2. Cooling loads and climate The cooling load profile of a reference office building was considered. Detailed descriptions of the building and its load profile

Fig. 2. Schematic of all-variable speed chiller system.

Three identical cross-flow cooling towers were used, each of which catered for the heat rejection rate of each chiller at design conditions and was equipped with a fan running at 100/50% of full speed. Through the high/low fan speed control, the towers produced the leaving cooling water at 29.4 °C required for heat rejection at the condensers for various combinations of load and wetbulb temperature conditions. The total chilled water pump head required was 400 kPa, 160 kPa of this was allocated to the primary pumps and the remainder to the secondary pumps. The two secondary pumps operated one by one to meet the changing chilled water flow demand of the airside system coils. Each of the condenser water pumps was designed to operate at a head of 200 kPa while circulating the nominal flow rate to the condenser. With regard to the all-variable speed system, the number and design capacity of the components are quite similar to those of the conventional system except that all the components are controlled by VSD and the primary-only chilled water distribution system was used. The variable speed chillers had a full load COP of 6.1, the same as the constant speed chillers, but with a much higher IPLV of 9.8 because of their superior part load performance. The system pressure head of 400 kPa (same as the conventional system ones) was applied to determine the head requirement of the variable speed chilled water pumps. The minimum speed for all the pumps was set at 50% of full speed to ensure sufficient water flow rate at the evaporator and condenser of each operating chiller. The tower fan speed could be adjusted continuously from 10% to 100% of full speed. The surplus chilled water would flow through the low load bypass pipe only when the load which the lead chiller carried called for less than 50% of the rated flow. For any given combina-

Table 1 Physical data of two chiller systems System

Conventional

All-variable speed

Total plant cooling capacity (kW) For each chiller Refrigerant type Nominal cooling capacity per chiller (kW) Nominal compressor power (kW) COP at full load Integrated part load value (IPLV) Design chilled water supply/return temperature (°C) Design chilled water flow rate (l/s) Design condenser water entering/leaving temp. (°C) Design condenser water flow rate (l/s) For each cooling tower Type Heat rejection capacity (kW) Design entering/leaving temperature (°C) Water flow rate (l/s) Fan type

6390

6390

R134a 2130 350 6.1 7.0 7/12.5

R134a 2130 350 6.1 9.8 7/12.5

92.4 30/35.5

92.4 30/35.5

107.6

107.6

Cross-flow 2480 30/35.5 107.6 Axial, two speed 84 100.8 30.4 0.2 0.53

Cross-flow 2480 30/35.5 107.6 Axial, variable speed 84 100.8 30.4 0.2 0.53

26 26.1

26 62.7

54.9

–

28.8

28.8

Air volume flow rate (m3/s) Air mass flow rate (kg/s) Fan motor power (kW) Drift loss (% of nominal flow) Design water evaporation rate (% of nominal flow) Design wet-bulb outdoor temperature (°C) Rated power of each primary chilled water pump (kW) Rated power of each secondary chilled water pump (kW) Rated power of each condenser water pump (kW)

F.W. Yu, K.T. Chan / Applied Thermal Engineering 29 (2009) 1721–1729

30

No. of operating hours

800

25

700 600

Building load profile Average wet bulb temperature

500

15

400 300

10

200

5

100 0

20

0.1-.2 .2-.3 .3-.4 .4-.5 .5-.6 .6-.7 .7-.8 .8-.9

.9-1

0

9 ECWT (oC)

8

Chiller COP

900

Average wet bulb temperature at each load bin (oC)

1724

18.3 23.9 29.4

7 6

Optimal region

5 4 3 0

0.2

evaluation are given elsewhere [11]. The building had a peak load of 6389 kW with annual cooling hours of 2834 h (90.5% of the building’s annual opening hours). The hourly cooling loads of the building were modelled in the subtropical Hong Kong climate using an example weather year. The annual cooling energy was 7,423,883 kWh. Fig. 3 shows the load-frequency distribution of the building along with the average outdoor wet-bulb temperature at each load bin. The hourly building cooling load data are expressed as ratios to the peak level of 6389 kW. It is undesirable to have a single chiller system operating for the building load profile as the first bar at the load bin of 0.1–0.2 in Fig. 3 illustrates, the chiller needs to work at a part load ratio of 0.2 or below with very low COP for 28% of the total cooling hours. Given this situation, three identical chillers were considered for the two systems serving the building. It is possible to ensure that each of the three chillers carries a part load ratio (PLR) of at least 0.28 which is well above its minimum allowable capacity step. The minimum PLR of 0.28 refers to a situation where one of the three chillers rated at 2130 kW operates for the lowest building load of 596 kW. 2.3. Optimal region of chiller Given that chiller power dominates the system power, it is worth determining how to stage chillers for their most efficient operation. Running chillers at full load as far as possible is a conventional way to minimize system power, considering that most existing chillers with constant speed compressors have maximum COP at full load and that pumping energy per unit cooling capacity is lowest at full load conditions. However, prolonging the full load operation of chillers cannot achieve the least power consumption if they have superior COP at part load instead of at full load. It should be noted that in absence of storage devices, operating the chiller at part load in order not to dissipate surplus cooling energy is convenient, although the COP could result in a lower level. To meet the actual cooling load, it is essential to look for minimum power consumption of the whole plant rather than maximum chiller COP. A more generic approach to operating chillers with maximum efficiency is based on the optimal loading points or region at which the maximum COP takes place for any given condenser water entering/leaving temperature. Figs. 4 and 5 show the part load performance curves of the constant speed and variable speed chillers together with their optimal region at different entering condenser water temperatures (ECWT). With regard to the constant speed chiller, the COP drops considerably at low part load ratios and the local maxima are situated in a flat region of the curve (PLR of 0.8–1) with a slight variation of ±5% in COP, regardless of the condenser water temperatures. For the variable speed chiller, the COP increases considerably, taking into

0.6

0.8

1

Fig. 4. Part load performance curves for the constant speed centrifugal chiller.

15

Chiller COP

Fig. 3. Frequency distribution of hourly building load ratios with average wet-bulb temperature at each load bin.

0.4

Chiller part load ratio

Building load ratio

10

ECWT ( o C) 18.3 23.9 29.4

5 Optimal region

0

0

0.2

0.4

0.6

0.8

1

Chiller part load ratio Fig. 5. Part load performance curves for the variable speed centrifugal chiller.

account the condenser temperature relief under part load conditions. The maximum COP occurs at lower chiller loads when ECWT drops, resulting in the optimal region leaning towards the low load conditions. ECWT could be considered as a parameter other than COP used to adjust the optimal chiller load region. The schedules of chiller sequencing shown in Tables 2 and 3 were formulated for the two systems in order to maximize the chiller performance. Regarding the conventional system, the traditional principle of chiller sequencing is still applicable as the optimal region of the chillers is quite close to the full load for all condenser water entering temperatures. The least number of chillers operating is still used to meet the changing building cooling load. While for the all-variable speed system, all the three chillers would be staged starting from a building load ratio of 0.63, instead of 0.67, in order to allow them to run as closely as possible to the optimal region. The start-up frequencies of the three chillers in the two systems are roughly the same. 3. Simulation of electricity and water use for the chiller system The chilled water plant simulation was performed with the use of component models for the chillers, pumps and cooling towers developed under the TRNSYS 15 environment [12]. Fig. 6 shows a flow diagram of the modelling the chiller under TRNSYS 15. Please refer to Ref. [13] for details of the model and equations used. Input variables of the model comprise the chiller load (Qcl), dry-bulb temperature (Tadb), wet-bulb temperature (Tawb), chilled water supply temperature (Tchws), chilled water mass flow rate (mw), condenser water flow rate (mcdw) and condenser water entering temperature (Tcdwe). Given these input variables, the model computes over 30 operating variables, the power-related variables and chiller

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Variable speed: gcc

Table 2 Schedule of staging chillers for the conventional system Building load ratio (BLR)

0 < BLR 6 0.33 0.33 < BLR 6 0.67 0.67 < BLR 6 1

¼ ð0:0291T cdwe  1:58ÞPLR2 þ ð0:0485T cdwe þ 2:904ÞPLR

No. of total cooling hours

No. of operating chillers

Total capacity of operating chillers

Chiller part load ratio

1171 1086 577

1 2 3

2130 4260 6390

0.28 0.5 0.67

Minimum

1 1 1

Table 3 Schedule of staging chillers for the all-variable speed system Building load ratio (BLR)

No. of total cooling hours

No. of operating chillers

Total capacity of operating chillers

Chiller part load ratio Minimum

Maximum

0 < BLR 6 0.33 0.33 < BLR 6 0.63 0.63 < BLR 6 1

1171 969 694

1 2 3

2130 4260 6390

0.28 0.5 0.63

1 0.95 1

Data file of operating conditions Data reader

Constant parameters

Input variables: Qcl, Tadb, Tawb , Tchws, mw, mcdw, Tcdwe

þ 0:0095T cdwe  0:0475

ð3Þ

Maximum

Equations for evaluating refrigerant properties

The compressor and condenser have to satisfy the mass balance of refrigerant and energy balance at the evaporator. The chiller model was calibrated to imitate the part load COP curves shown in Figs. 4 and 5. These curves were constructed based on performance data from a chiller catalogue with given flow and temperature conditions of the chilled water and condenser water. The model is capable of capturing the COP with deviation of 2.8% to 9.1% at most under various operating conditions when the design and control of chiller components are changed. Regarding the conventional system, the power of the constant speed pumps varied step by step, depending on the number of chillers operating. While variable speed pumps were used, their power was evaluated by using pump affinity laws with the computed flow rate or rotating speed. With regard to the cooling tower model, the Ntu-effectiveness (Ntu—number of transfer units) approach was used [14], under which a change in the heat transfer effectiveness of the tower at various combinations of airflow and water flow is taken into account. This approach enabled the heat rejection airflow required to be computed, based on the given wet-bulb temperature and the condenser water entering temperature calculated by the condenser model. The fan power (Ect) was evaluated by using polynomial approximations of fan characteristics under two speed control or under variable speed control, based on the airflow rate (Va) calculated. Equations for the fan power calculation are given in Eqs. (4) and (5). Ectr and Var represent the values of Ect and Va at the full speed condition.

Two speed : Ect ¼0:5Ectr Va =Var

Chiller model input file

for Va 6 0:5Var ;

Ect ¼Ectr ð1:5Va =Var  0:5Þ for 0:5V ar < Va Evaporator

Compressor

Condenser

Variable speed : Ect ¼Ectr ðVa =Var Þ

Printer Operating variables in each operating condition

Control algorithm of cooling towers

Fig. 6. Flow diagram of modelling the chiller under TRNSYS 15.

COP, through solving over 36 algebraic equations in an iterative process. The structure of the chiller model is based on the thermodynamic models used extensively to investigate the energy performance of chillers. Mechanistic relations between chiller components were taken into account. The log mean temperature difference (LMTD) method was used to model the heat transfer characteristics of the evaporator and condenser under the full load and part load conditions. Given an isentropic work input (Win) to the compressor, a combined motor and transmission efficiency (gcc) curve was used to determine the actual chiller power (Ecc = Win/gcc) at various part load conditions when the capacity control was done either by modulating the inlet guide vanes at constant speed or by regulating the rotating speed via VSD. gcc was determined using regression analysis with the performance data of Ecc in various operating conditions. As Eqs. (2) and (3) illustrate, gcc is expressed as a function of chiller part load ratio (PLR) and condenser water entering temperature (Tcdwe).

Constant speed: gcc ¼ ð0:0275T cdwe þ 0:2708ÞPLR2 þ ð0:025T cdwe þ 0:5612ÞPLR þ 0:0123T cdwe  0:1059

ð2Þ

3

ð4Þ ð5Þ

The overall annual electricity consumption of the chiller plants is the sum of all hourly chiller power, pump power and tower fan power, with respect to the building cooling load profile and the schedule of staging chillers. The water consumption of a cooling tower is made up of three parts of water loss: evaporation, drift and bleed-off. The cooling tower model considered the overall mass and energy balance of cooling water and outdoor air to simulate the states of the air leaving the tower and, in turn, to estimate the evaporation loss rate. The drift rate and bleed-off rate were assumed to be 0.2% and 0.6%, respectively, of the cooling water circulation rate, with regard to the use of traditional chlorination water treatment.

4. Results and discussion 4.1. Annual electricity and water consumption The annual electricity and water consumption was estimated for the two chiller systems, as summarized in Table 4. All the values are normalized by the total air-conditioned floor area (i.e. 42840 m2) of the building for easy comparison with other buildings of the same type, of similar use and located in a similar climate zone. For the variable speed chiller system, the savings of energy use components in relation to the conventional case are given. Some important observations can be made about the simulation results. Using the all-variable speed system could bring about savings in all the energy use components and water use. A 19.7% decrease in the annual plant electricity consumption is due mainly to the electricity saving of the chilled water pumps, followed by

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Table 4 Annual plant electricity and water consumption per unit of air-conditioned floor area of the building System

Conventional (1)

All-variable speed (2)

Normalized annual electricity consumption (kWh/m2) Chiller 28.2 25.7 Chilled water pump 5.8 3.0 Condenser water pump 3.4 1.3 Cooling tower fan 1.0 0.9 Total 38.5 30.9 Average chiller COP 6.1 6.7 507.2 426.7 Normalized annual water 2 consumption (L/m )

Savings from the conventional case (1)–(2) 2.5 2.8 2.1 0.1 7.6 – 80.6

(9.0%) (48.5%) (61.8%) (10.3%) (19.7%) (15.9%)

the condenser water pumps, chillers and tower fans. This suggests that reducing the flow rate of chilled water and condenser water is a viable means to decrease greatly the system power during part load operation, though it may give a diminishing return on the reduction of chiller power resulting from a moderate increase in the COP. With regard to the variable speed system, there was little difference in the chiller COP between the nominal and reduced chilled water flow conditions. However, when the condenser water flow rate decreased with the chiller load or the wet-bulb temperature, the extent to which the chiller COP can increase would diminish in various degrees because the reduced condenser water flow tempered the condensing temperature relief to save the compressor power. Indeed, the average chiller COP—the annual cooling energy of 7,423,883 kWh over the annual chiller electricity consumption—increased moderately to 6.7 even when the variable speed chillers with a high IPLV of 9.8 were used. Given the reduced condenser water flow conditions, the water consumption of the cooling towers could drop by 15.9% in the all-variable speed case in relation to the conventional case. 4.2. Economic analysis In order to judge the cost-effectiveness of the variable speed chiller system, it is necessary to predict the capital cost investment and how much operating cost this system can save. Local tariff structures were considered to evaluate the operating costs associated with the annual electricity and water consumption of the chiller systems. The maximum demand tariff of one of the local power companies was used to calculate the annual electricity consumption. This tariff involves monthly demand charges of HK$42.1/ kVA for the first 400 kVA and HK$41.1/kVA for the next additional kVA, and monthly energy charges of HK$1.023/kWh for the first 200 units (i.e. kWh) supplied per month per kVA of maximum demand and of HK$0.962/kWh for each additional unit supplied. Regarding the water charge tariff, the make-up water cost of HK$4.58/m3 and the sewage charge of HK$1.2/m3 from bleed-off were used, based on guidelines stated in the pilot scheme for the wider use of fresh water in evaporative cooling towers for energy-efficient air-conditioning systems [15]. Assuming that the variable speed system does not incur any additional recurrent costs (such as repair and maintenance costs), its cost-effectiveness can be analysed by using a simple payback, net present value, and the internal rate of return which is a measure of return in percentage to be expected on a capital investment. The life cycle operating cost was also considered, which is the sum of the discounted value of the annual electricity and water costs over the lifespan of the system. A 15-year economic life was used for the chillers with reference to the economic analysis for mechanical equipment in ASHRAE 90.1 [16]. During this 15-year lifespan, it was assumed that no major alternation of the chiller

Table 5 Operating costs comparison for the two chiller systems System

Conventional (1)

180 Capital cost (HK$/m2) Annual operating cost (HK$/m2) Electricity 46.4 Water 2.7 Total 49.1 Simple payback (yr) – Internal rate of return – (for 15 yr) (%) Life cycle operating cost (HK$/m2) Electricity 353.0 Water 20.2 Total 373.1 Net present value 553.1

All-variable speed (2)

Savings from the conventional case (1)–(2)

198

18

39.4 2.2 41.6 2.4 (=18/7.5) 41.5

299.8 16.7 316.5 514.5

7.0 0.5 7.5

53.2 3.5 56.7 38.6

system would take place. Based on the current bank loan interest rate and the perception of the future monetary value, a discount rate of 10% was used in the calculation of the life cycle operating cost [17]. The 15-year economic life was considered in the calculation of the internal rate of return. Table 5 gives the economic indicators of the two systems. All the cost figures are presented as HK$ per unit of the total air-conditioned floor area in m2 of the building. For the two systems, the annual electricity cost accounts for around 95% of the annual total operating cost. Given this, the operating cost saving is virtually proportional to the annual electricity saving. Considering that the capital investment of the conventional chiller system is HK$180/ m2 (assuming HK$1200 per kW of plant cooling capacity – an indicative figure from informal consultation with local chiller suppliers), the simple payback could be 2.4 years based on the assumption that the variable speed drives and the associated control system of the all-variable speed system could cause an additional cost by up to HK$18/m2 or 10% of the capital cost of the conventional system. In this situation, the building owner who purchases the variable speed chiller system instead of the conventional ones can enjoy a net savings of HK$38.6/m2 when operating the chiller system throughout the 15-year lifespan. The internal rate of return was identified to be 41.5% which is much higher than the assumed discount rate of 10%. This suggests that the variable speed system is a better-paying investment. It will be more economically attractive to adopt such an all-variable speed chiller plant if there is a cost premium to balance the additional cost resulting from the use of VSD and a demand side management programme which provides a rebate for electricity savings by use of energy-efficient measures. 4.3. Load-based speed control for all-variable speed chiller system All the equipment in a variable speed chiller system should be provided with proper speed control in order to effectively operate the system with minimum power consumption. Each chiller makes use of its own microprocessor with proprietary control to adjust the compressor speed in response to the changing chiller load while controlling the chilled water supply temperature at its set point. The minimization of energy use for the system, therefore, refers to optimizing the operation of chilled water pumps, condenser water pumps and cooling tower fans for all ambient and chiller load conditions. Given the variable primary flow conditions, it is worth ascertaining how to control the speed of the chilled water pumps which deliver the flow required to meet the chiller load while allowing the chillers to run closely at their optimal region in order to minimize the total power consumption. Based on the

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simulation analysis, the variable speed drive should regulate the pump speed (RPMchp) as a function of chiller part load ratio (PLR) based on Eq. (6). RPMchp,full is the full speed of the pump.

RPMchp ¼



PLR RPMchp;full if PLR > 0:5

ð6Þ

0:5RPMchp;full if PLR 6 0:5

This direct load-speed correlation in Eq. (6) is equally applicable for the operation of the condenser water pumps to achieve system optimization. Such a load dependent control scheme for pump speed has been supported by another study on optimizing chiller plant controls [6], with which the condenser water pump control is given by: % CW pump speed  no. of pumps = 0.08  plant load (tons) + 50.42. The actual form of the load-speed relationship will be different if the chiller system has different designs with respect to the number and size of chillers and pumps to meet the peak building demand. It is envisaged that the simple load-speed relationship could facilitate the commissioning procedure for optimized system performance and the implementation of automated fault detection and diagnostics because it is executed by feed forward control in response to the changing chiller load instead of the conventional feed backward control in maintaining the chilled water and condenser water temperatures. Compared to Hartman’s power-based control that the speed of pumps and tower fans is related to the power drawn by the operating chiller, the load-speed relationship reported here is even straightforward and easy to be recognized for other similar chiller applications. It is preferable to use the load-based speed control than the power-based speed control because the chiller load is dictated purely by the building cooling load and is independent on the interdependence of the power relationships between the chillers, pumps and tower fans. Furthermore, using the load-based speed control enables the flow rates of chilled water and condenser water to vary linearly with the chiller load, given that speed is directly proportional to flow rate according to the pump laws. In this situation, the ‘‘delta-T” or temperature difference of the chilled water and condenser water can remain constant at their design conditions. This results in more stable temperature control which is essential for maintaining the dehumidifying capacity of airside system coils and helps avoid the low delta-T syndrome which leads to premature staging of chillers at light load conditions. Overall applying load-based speed control to pumps with design ‘‘deltaT” is a viable means to optimize the trade-off between the compressor power and pump power. To achieve absolute system optimization with minimum energy consumption, the speed of cooling tower fans should be controlled in response to not only the chiller load but also to the outdoor wet-

bulb temperature, given that the wet-bulb temperature is one of the key variables influencing the heat rejection effectiveness of the towers. Yet it is possible to determine the optimum tower fan speed based simply on the chiller load only to have near optimal control of the system. The data set shown in Fig. 7 refers to the optimum fan speed at which the minimum total power consumption took place when the lead chiller operated for the building cooling profile with various combinations of chiller loads (PLR of 0.14–0.95) and wet-bulb temperatures (Tawb of 7–22 °C). Based on the regression curve with the coefficient of determination of (R2) of 0.95, the near-optimum percentage fan speed (RMPfan) can be predicted by Eq. (7), where RMPctf,full is the full speed of the fan.

RMPctf ¼ RMPctf;full ð0:7039PLR þ 0:2512Þ

A generic form for the direct fan speed control would be: RPMctf = c1 PLR + c2, where c1 and c2 are constants to be determined based on the performance characteristics and power rating of cooling towers. Such kind of simplified tower fan control has been given in a study on a generic control algorithm for cooling towers [18]. Referring to traditional cooling water temperature control, the approach (condenser water entering temperature minus ambient wet bulb) may be fixed irrespective of operating conditions to enhance chiller– tower control. Yet based on the graphical analysis shown in Fig. 8, the optimal approach with minimum power consumption varied between 3.6 and 4.9 °C in response to the chiller load, but the approach-load relationship is very vague when the chiller operated at above half load. This suggests that it is not feasible to use a cooling water temperature (with varying values at a given chiller load) to control the tower fan speed in order to achieve minimum power consumption. Another chiller system simulation was carried out to identify the likely implications of using load-based speed control on the environmental performance of the variable speed system. It is found that the control brought about no appreciable difference in the annual water consumption, but only caused a slight increase of 2.8% in the annual electricity consumption of the chillers and tower fans. The load-based speed control for cooling tower fans, indeed, eliminates the need of using high quality temperature or relative humidity sensor to measure the wet-bulb temperature. This, in turn, helps minimize the noisy control signal resulting from the measurement errors, and at the same time, to provide more stable fan speed control because the wet-bulb temperature tends to vary often compared to the chiller load. It is essential to equip variable speed chiller plants with proper instruments to perform real-time measurement and verification of

5.0

Optimal approach (Tcdwe - Tawb) with minimum power consumption (o C)

% full speed of fans at minimum power consumption

100

80

60

40

20

Near-optimum % fan speed = 70.387 PLR + 25.119 2

R = 0.9509

0 0.0

ð7Þ

0.2

0.4

0.6

0.8

Chiller part load ratio Fig. 7. Load-based speed control for cooling tower fans.

1.0

4.5

4.0

3.5

3.0 0.0

0.2

0.4

0.6

0.8

1.0

Chiller part load ratio Fig. 8. Relationship between optimal approach and chiller part load ratio.

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the building cooling load and the chiller load in order to implement load-based speed control for the pumps and tower fans. The monitoring of these loads is crucial for ascertaining whether the chillers are staged according to their optimal region and for allocating the system load evenly to the operating chillers. Under the variable flow conditions, neither the flow rate nor temperature difference of chilled water can serve to indicate the actual chiller load. The load which each operating chiller carries should be computed directly from the measured flow rate and temperatures of the chilled water. Based on the chiller load calculated, the speeds of the pumps and tower fans associated with the chillers operating can be readily determined for optimized operation. 4.4. Remarks on usingequal marginal performance principle (EMPP) to achieve system optimization The results of this study indicate that the energy performance of a chiller system can be improved considerably by using all-variable speed configurations. To achieve such improvement, it is important to minimize the aggregate power of all the system components for all operating conditions. The load-based speed control has been introduced as one of the means to achieve minimum power consumption of all-variable speed chiller plants. Hartman, on the other hand, has demonstrated the EMPP as a theoretical approach to understanding the power relationships between the components for optimizing all-variable speed chiller plants [8,9]. One criterion for applying EMPP is to express the system output (cooling capacity, Q) as a function of components’ power inputs in the form: Q = f(Component 1 power, Component 2 power . . . Component n power). When the mathematical expression of this formula is identified, the optimum system COP can be determined by equating all marginal COPs of the power components, where marginal COP for component ‘‘x” is given by o Q/oComponent ‘‘x” power. Creating a cooling output expression for a multiple-chiller system is a complicated and even difficult process, which is not the scope of this study. Yet some remarks are presented below on using the EMPP to implement system optimization. There are a total of 12 power components with regard to a variable speed system with three pairs of chillers, chilled water pumps, condenser water pumps and cooling tower fans. Due to the sequencing of chillers, each component power varies in discrete steps at certain loading points over the entire range of building cooling loads. This discrete power variation can cause difficulty in evaluating the marginal COPs for each power component. To counter this, it may be necessary to create a set of system output expressions, each of which caters for a certain building cooling load range handled by a given number of operating chillers. The time and effort to create the system output expressions and to determine the marginal COPs would be highly intensified if the chiller system is on a large scale with many sequencing (or staging) patterns of the power components. While there is a generic form for expressing the system output, the coefficients and mathematical formula for representing each power component are specific, given that each chiller system should have its own design and power requirements. At present no analytic tool has been developed for using EMPP to determine the mathematical expression of cooling outputs for any given system design with respect to the number and size of chillers. Under this circumstance, the power relationships in mathematical form may be identified by using curve fitting techniques in cases where an ample set of power-related operating data is collected for a system with various cooling outputs. This suggests that the EMPP is viable for use after the system is working in the post-operation stage. It is more desirable, however, to apply the EMPP in the system design stage to facilitate optimum equipment selection and power-based control development. It remains to be seen how the evaluation of system design alternatives with the use of EMPP be-

comes a more simple and straightforward process. Compared to the EMPP, the load-based speed control reported in this study is generic and easy to be implemented for any chiller system with typical design as it considers only the relationship between the chiller load ratio and fractional speed of the equipment for optimum operation. Based on the system output expression, the EMPP is applicable for any system where each component has varying power characteristics in response to changes in the cooling output only. This means that for a given system output, there is no other factor varying the power components except their interaction with each other. Yet it is found from the simulation results that the cooling tower fan power could change in response to various wet-bulb temperatures while the system load and power components of the chillers and pumps remain unchanged. The dependence of the tower fan power with the wet-bulb temperature may call for a set of system output expressions, each of which deals with a specific wet-bulb temperature. This may complicate the evaluation of marginal COP for each power component and the power-based speed control requirements. Another criterion for applying EMPP is that each system component has to be sized and optimized in its operation using identical methodologies. This means that all the components should have identical performance at all operating conditions and when running in parallel, they need to operate at the identical fractional speed to achieve system optimization. The design and control of chiller systems under the EMPP philosophy are contrary to what have been implemented to conventional constant speed chiller plants which are preferable to be designed with unequally sized chillers operating primarily with an uneven load sharing strategy in order to prolong their full load operation for plant optimization. This study has not been dealt with some universal rules that govern the optimal design of all-variable speed chiller plants with respect to the number and capacity of chillers, the associated pumps and cooling towers. The system output formula used with EMPP may form a good basis for evaluating optimal design solutions as the marginal (or maximum) system performance can be determined numerically with a given equation representing power relationships between all the system components. It is hoped that an explicit methodology will be launched by EMPP proposers or developers to facilitate the evaluation of an output equation for a given system configuration and to examine the equation’s accuracy. Designing a chiller plant with EMPP may help ascertain if certain oversizing of some of the system components with an additional investment cost still gives an attractive payback while prolonging their operation in the optimal region with maximum COP.

5. Conclusions This study examines how all-variable speed chiller systems with load-based speed control yield economic benefits and superior environmental performance with reduced electricity and water consumption relative to the conventional constant speed systems. Given the superior performance of variable speed devices at part load operation, the basic principles of designing an all-variable speed chiller plant are using fewer and equally sized chillers and considering a single-loop chilled water circuit with variable flow. The variable speed chillers should be staged based on their optimal region with maximum COP. With regard to an all-variable speed chiller system operating for the cooling load profile of an office building, the annual electricity consumption could decrease the annual total electricity use by 19.7% and annual water use by 15.9% relative to the corresponding constant speed plant. Loadbased speed control is introduced as a simple and direct means

F.W. Yu, K.T. Chan / Applied Thermal Engineering 29 (2009) 1721–1729

to achieve optimal operation of all the variable speed equipment while complying intrinsically with temperature requirements of chilled water and condenser water. Under this control, the speed of the chilled water pumps, condenser water pumps and tower fans is regulated simply based on the chiller part load ratio. This makes the optimization technology more easy and generic and eliminates the need to identify power relationships between the system components as required by EMPP. Acknowledgement The work described in this paper was supported by a grant from the central research grant of The Hong Kong Polytechnic University, Project A/C Code: G-U272. References [1] H. Crowther, J. Furlong, Optimizing chillers and towers, ASHRAE J. 46 (7) (2004) 34–40. [2] T.Q. Qureshi, S.A. Tassou, Variable-speed control in refrigeration systems, Appl. Therm. Eng. 16 (2) (1996) 103–113. [3] W.P. Bahnfleth, E.B. Peyer, Energy use and economic comparison of chilledwater pumping system alternatives, ASHRAE Trans. 112 (2) (2006) 198–208. [4] S.T. Taylor, Degrading chilled water plant delta-T: causes and mitigation, ASHRAE Trans. 108 (1) (2002). Paper No. AC-02-6-1. [5] M. Hydeman, P. Sreedharan, N. Webb, S. Blanc, Development and testing of a reformulated regression-based electric chiller model, ASHRAE Trans. 108 (2) (2002). Paper No. HI-02-18-2.

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