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Performance evaluation of an indirect evaporative cooler under controlled environmental conditions
Energy and Buildings 62 (2013) 278–285
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Performance evaluation of an indirect evaporative cooler under controlled environmental conditions Aftab Ahmad ∗ , Shafiqur Rehman, Luai M. Al-Hadhrami Center for Engineering Research, Research Institute, King Fahd University of Petroleum and Minerals, Dhahran-31261, Saudi Arabia
a r t i c l e
i n f o
Article history: Received 8 August 2012 Received in revised form 10 March 2013 Accepted 11 March 2013 Keywords: Indirect evaporative cooler Dew-point effectiveness Wet-bulb depression Intake air energy efficiency ratio Water evaporation rate Wet-bulb temperature
a b s t r a c t The study investigated the performance of a 5-ton capacity indirect evaporative cooler under controlled environmental conditions (43.9 ◦ C dry-bulb temperature and 19.9% relative humidity) but for different air flow rates (631 to 2388 m3 /h). The experimental results showed that the intake air energy efficiency ratio of the cooler varied from 7.1 to 55.1 depending on test conditions and air flow rate. The power consumption of indirect evaporative cooler was found to vary from 68.3 to 746 watts. Water consumption was found to vary between 0.0160 and 0.0598 m3 /h. At full fan speed, an average of 58.7% of the total water consumed by indirect evaporative cooler was evaporated. The results indicated that intake air energy efficiency ratio was directly proportional to the wet-bulb depression. The study also showed that the indirect evaporative cooler is suitable for hot and dry climatic conditions. © 2013 Elsevier B.V. All rights reserved.
1. Introduction Kingdom of Saudi Arabia is going through rapid economical growth. As a result of which new and modern infrastructure, which is energy intensive, is taking physical shape. In 2009, Saudi Electricity Company (SEC) delivered a total of 193,472GWh of energy, an increase of 6.8% over the previous year [1]. The number of customers increased over the same period by 5.2% to 5701,516. As shown in Fig. 1, the residential sector consumes more than half of the total electricity sold by SEC and industrial sector consumes around 18%. In Saudi Arabia the major portion of the electricity is consumed to cool buildings, offices, residential houses, and refrigeration storage systems. There are mainly three basic technologies which are used in air cooling process. These include mechanical vapor compression, absorption systems, and evaporative indirect cooling systems. Evaporative cooling is an efficient and sustainable method for cooling. Theoretically, the ultimate goal of evaporative cooling system is to achieve the ambient air wet bulb temperature, which is a bit difficult to be reached practically. In literature, a coefficient of performance (COP) of 8 to 20 has been reported for indirect evaporative cooling systems [2,3] while 2 to 3 for mechanical vapor system and 0.4 to 1.2 for absorption system [4,5]. Besides all goods, indirect evaporative cooling systems have lower cooling capacity resulting in limited temperature reduction [6]. In these systems, the cooling is limited by the
∗ Corresponding author. E-mail address: [email protected] (A. Ahmad). 0378-7788/$ – see front matter © 2013 Elsevier B.V. All rights reserved. http://dx.doi.org/10.1016/j.enbuild.2013.03.013
wet-bulb temperature of the intake air. In real time applications, the temperature of the supply air is usually higher than the inlet wetbulb air temperature [7]. The indirect evaporative cooling systems can be used effectively in in temperate climate and regions situated at high latitudes, as reported by Costelloe and Finn [8,9]. The benefits of this cooling measure are most effective when applied to buildings requiring a relatively low mechanical air flow rate and comparatively high supply air temperature. As mentioned by Holmes and Hacker [10] such criteria are mostly applicable in building designs utilizing cooling methods such as chilled beams and displacement ventilation. Smith et al. [11] reported that evaporative cooling should continue to function as an effective low-energy cooling technique in future, warming climates. Riangvilaikul and Kumar [12] presented theoretical performance of a novel dew point evaporative cooling system operating under dry, moderate and humid climate conditions and the influence of major operating parameters such as velocity, system dimension and the ratio of working air to intake air. In the recent times the evaporative cooling systems have been extensively studied and used for industrial and residential applications due to low energy consumption [13]. Hasan [14] presented a method to achieve sub-wet bulb temperature by indirect evaporative cooling of air without using a mechanical vapor compression machine. The study showed that the proposed model for indirect evaporative coolers can be based on the e-NTU method for sensible heat exchangers with some adjustments made by redefining the potential gradients, transfer coefficient, heat capacity rate parameters and by assuming a
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Nomenclature Cp DBT DPT IA-EER IEC HMX ˙ sup m Qexh Tdbt , Tdbt , sup Tdpt , Uexh Uin WBT
moist air specific heat, J kg−1 K−1 dry-bulb temperature, ◦ C dew-point temperature, ◦ C intake air energy efficiency ratio, W W−1 indirect evaporative cooler heat and mass exchanger supply air mass flow rate, kg s−1 exhaust air flow rate, m3 /h in intake air dry-bulb temperature, ◦ C supply air dry-bulb temperature, ◦ C in intake air dew-point temperature, ◦ C exhaust air humidity ratio, kg moisture (kg dry air)−1 intake air humidity ratio, kg moisture (kg dry air)−1 wet-bulb temperature, ◦ C
Subscripts dbt dry-bulb temperature wbt wet-bulb temperature dew-point temperature dpt in intake exh exhaust sup supply Greek Symbols air density, kg m−3
linear saturation temperature-enthalpy relation of air. The model results showed good agreement with results from experimental measurements and a numerical model. Riangvilaikul and Kumar [15] carried out experimental measurements on a dew point evaporative cooling and found a wet-bulb effectiveness of 92–114%. Experimental results of a two-stage combined indirect-direct evaporative air cooler system showed a wet-bulb effectiveness of 90–120% (El-Dessouky et al. [16]). Farahani et al. [17] investigated the performance of a two-stage cooling system consisting of a nocturnal radiative unit, a cooling coil, and an indirect evaporative cooler for weather conditions of Tehran, Iran. The results demonstrated that the first stage of the system increases the effectiveness of the indirect evaporative cooler. Hsu et al. [18] studied theoretically and experimentally, counter flow, cross flow and regenerative indirect evaporative coolers to achieve dew point temperature of the ambient air. Crum et al. [19] also demonstrated that dew point temperature can be attained by using multistage indirect evaporative cooling and by combining a cooling tower and a heat exchanger. For an appropriate mass flow rate and chosen geometry of the cooler, Maclaine-cross and Banks [20] showed that regenerative evaporative cooling process can attain the dew point temperature. Earlier, Pescod [21] reported that by splitting a portion of the air produced by an indirect evaporative cooler and using it in the wet passage, the target wet-bulb temperature of the cooling process could be lowered. Facao [22] compared simplified models [23,24] with detailed models such as numerical models [25] and noticed that simplified models based on an overall approach provide as good or even better results as those based on finite differences. Al-Hadhrami et al. [26] designed a hybrid solar air-conditioning system using indirect evaporative cooler. The hybrid system uses solar energy and liquid desiccant system to dry or condition the high humidity ambient air before it enters into the indirect evaporative cooler. Therefore, the designed hybrid solar air-conditioning system can be used in even regions having high humidity. The
Fig. 1. Energy consumption by sector for 2009, Saudi Arabia [1].
designed system can work in all climatic conditions. In addition to the air-conditioning, a distilled water recovery system is designed to recover water from the desiccant and from the indirect evaporatively cooled air in the form of distilled water. The present study provides the finding of experimental results of an Indirect Evaporative Cooler (IEC) tested under controlled environmental conditions in the IEC testing chamber. The testing chamber consisted of environmental section (include solid desiccant wheel system, gas-based air heating system, and an air cooler), IEC section, and air conditioned section. The IEC unit was instrumented to measure its performance by using different types of sensors. The data loggers were used to log the data automatically for analysis to determine the performance of the IEC unit. 2. Thermal comfort Thermal comfort is an important concept for climate control systems and beyond. To have “thermal comfort” means that a person wearing a normal amount of clothing feels neither too cold nor too warm. Thermal comfort is important both for one’s wellbeing and for productivity. It can be achieved only when the air temperature, humidity and air movement are within the specified range often referred to as the “comfort zone”. In general thermal comfort is defined as that condition of mind which expresses satisfaction with the thermal environment. According to ASHRAE standard [27], the thermal comfort is defined as that condition of mind which expresses satisfaction with the thermal environment and is assessed by subjective evaluation. According to British standard [28], the thermal comfort is defined as that condition of mind which expresses satisfaction with the thermal environment. The factors which affect the thermal comfort of human beings are classified as environmental and personal. The environmental factors include the air temperature, humidity, air speed and radiant temperature while personal factors include the clothing and personal activity and conditions. ASHRAE [27] has developed an industry consensus standard to describe comfort requirements in buildings. The purpose of this standard is to specify the combinations of indoor thermal environmental factors and personal factors that will produce thermal environmental conditions acceptable to a majority of the occupants within the space. One of the most recognizable features of ASHRAE standard 55-2010 [27] is the ASHRAE comfort zone as portrayed on a psychrometric chart, Fig. 2. The comfort chart shows the acceptable range of operative temperature and humidity for spaces that meet the criteria specified in Section 5.2.1.1 of the standard. The ASHRAE standard allows the comfort charts to be applied to spaces where the occupants have activity levels that result in metabolic rates between 1.0 met and 1.3 met and where clothing is worn
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Fig. 2. ASHRAE 2010 comfort chart [27].
that provides between 0.5 clo and 1.0 clo of thermal insulation. However, the comfort zone is shifted to the right or left in the psychrometric chart shown in Fig. 2 depending on the clothing, metabolic rate and radiant temperature.
additional cooler and desiccant wheel dehumidifier were used. The air flow rate to the climatic room was controlled by using the manual and automated damper through the control panel. 3.1. Working principle of the IEC
3. Experimental set-up The performance of a 5-ton capacity Indirect Evaporative Cooler (IEC) for different climatic conditions was measured in the instrumented IEC testing chamber. The Cooler was placed in an environmentally-controlled room in the chamber. The schematic diagram of testing chamber with measurement locations and other equipment is shown in Fig. 3. The testing chamber is equipped with devices for simulation of desired climatic conditions in the environmental section of the testing chamber. The gas fired burner and evaporative cooler were used for heating and humidification of outside air. For dehumidification and cooling of outside air, the
The cross-sectional view of indirect evaporative cooler (IEC) is shown in Fig. 4. Outside fresh air (1) is drawn into the cooler by a variable speed centrifugal fan (2) which forces air through the stack of 2 inch thick filters (3). The filtered air passes through the heat and mass exchange (HMX) modules (4) which split the air into two streams i.e. exhaust (5) and supply (6). The water supplied to the HMX modules is controlled by solenoid valve located on the water supply line. The supply air is sensibly cooled (without addition of humidity) while passing through the HMX modules. About half of the intake air is saturated with water inside the HMX modules and then released to the atmosphere as exhaust air.
Fig. 3. Schematic diagram of the IEC testing chamber.
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over the sampling period. The experiments were carried out by simulating the desired test conditions in the climatic room of the testing chamber and the test data of IEC unit data were automatically logged. The test data were downloaded from the data logger memory for further processing. 4. Test procedure
Fig. 4. Cross-sectional view of IEC unit.
3.2. Measurements and instrumentation The list of parameters measured and instruments used for the performance evaluation of the IEC system are listed in Table 1. All the temperature sensors were calibrated against the Traceable full thermometer from Control Company, USA. The supply and exhaust air flow rates were calibrated against the supply fan RPM by using a Fluke® pitot tube 922 air flow meter. The water flow sensor was calibrated manually by using graduated flask. The voltage and current sensors were calibrated against the Watt-Hour meter. The transducers for the differential and static pressure were calibrated using a Fluke® Pitot tube 922 air flow meter. All the sensors listed in Table 1 were connected to the ACRTM Systems data acquisition system. The data acquisition system consisted of different types of ACRTM data loggers which include 02 smart reader plus 8 eight-channel temperature data loggers, 01 smart reader plus 7 eight-channel process signal data logger, 01 smart reader plus 9 two-channel pulse data logger, and 01 smart reader plus 3 current, voltage and temperature data logger. The above data loggers were programmed for recording experimental data on minute-basis except for the pulse datalogger which was programmed for 2-minute data logging. The data loggers sampled data at every 8 seconds and stored the averaged data in its memory
The cooler was tested for different climatic conditions to evaluate its performance. The ASHRAE standard 143-2000 [29] for indirect evaporative cooler does not give specific test conditions for the evaporative coolers. However, it specifies the arrangement of the equipment, the type of measurements to be taken and accuracy of the equipment. In general, the evaporative coolers are rated mainly in terms of air flow. The test conditions for cooler are listed in Table 2. Prior to conducting test on the cooler, the evaporative media of the cooler was wetted by the soap by turning on/off pump. The purpose of soaping was to breakdown the water repelling films or oils on the evaporative media of the Heat and Mass Exchanger (HMX). The cooling unit was run for more than one hour to wet the HMX completely. The cooler test was conducted as follows: 1. The data acquisition system was started and all display readings were checked for proper functioning of the sensors. 2. The desired test conditions were set in the climatic room by the control panel by operating the air-conditioning equipment. 3. The Cooler was turned ON and allowed to run for at least one to two hours to get the experimental data for the desired test conditions. 4. After completion of the test, the logged data from data acquisition system was downloaded for further processing. 5. To get the next desired test conditions, the climatic room airconditioning was then adjusted by the control panel. The recorded data was averaged over the sampling period as set in the data acquisition system. The averaged recorded values were
Fig. 5. Plot of test cases on the psychrometric chart.
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Table 1 List of parameters measured and instruments used for Cooler testing.
1. 2. 3. 4. 5.
Parameter/Datalogger
Instrument
Range
Accuracy
Intake air dry-bulb temperature Intake air wet-bulb temperature Supply air dry-bulb temperature Supply air wet-bulb temperature Supply air static pressure
ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 Omega® Pressure transmitter PX 653, 1–5 VDC output, 0.5” WC Fluke pitot tube 922 air flow meter ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 Fluke® pitot tube 922 air flow meter ACR Systems Thermistor probes, ET–016 Ohio Semitronics voltage and current transducers 3AVT Monarch Instruments Optical Tachometer SPSR-115/230 Dwyer® Instruments multi-jet water meter, pulse type Weathronics Aneroid barometer ACR Systems current, voltage and temperature data logger, 128 kB ACR Systems Reader plus 8 eight-channel temperature data logger ACR Systems Smart Reader Plus 7 eight-channel process signal logger ACR Systems Smart Reader Plus 9 two-channel pulse data logger
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
0–0.5” W.C.
0.25% of full scale
1–80 m/s −35 to 95 ◦ C
±2.5% of reading at 10 m/s ±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
1–80 m/s −35 to 95 ◦ C
±2.5% of reading at 10 m/s ±0.2 ◦ C from 0 to 70 ◦ C
0–300 V
±1.0% of the reading
11.
Supply air flow rate Exhaust air dry-bulb temperature Exhaust air wet-bulb temperature Exhaust air flow rate Outside air dry-bulb temperature Supply fan voltage and current
12.
Fan speed (in RPM)
6. 7. 8. 9. 10.
13.
Water consumption rate
14. 15.
Barometric pressure Datalogger for current and voltage Dataloggers for wet-bulb and dry-bulb air temperatures Datalogger for differential pressure Datalogger for water flow and supply fan RPM
16. 17. 18.
1–250,000 rpm
–
0.28–5 m /h
±2% of full scale
910–1050 mbar 0–5 VDC
0.1 mbar ±0.5% of full scale
−40 to 70 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
0–5 VDC
±0.5% of full scale
4095 pulses per sample period
±1 pulse per sample period
3
Table 2 Test conditions for the indirect evaporative cooler (IEC). Test Case (TC) # 1. 2. 3. 4. 5. 6. 7. 8.
Test Conditions ◦
Fan Speed (rpm) ◦
DBT 43.9 C, RH 19.80% (WBT 24.4 C) DBT 43.9 ◦ C, RH 19.80% (WBT 24.4 ◦ C) DBT 43.9 ◦ C, RH 19.80% (WBT 24.4 ◦ C) DBT 43.9 ◦ C, RH 19.80% (WBT 24.4 ◦ C DBT 40.0 ◦ C, RH 27.72% (WBT 24.4 ◦ C) DBT 32.8 ◦ C, RH 56.58% (WBT 25.6 ◦ C) DBT 45.0 ◦ C, RH 17.96% (WBT 24.4 ◦ C) DBT 38.9 ◦ C, RH 45.38% (WBT 28.3 ◦ C)
used for performance evaluation of the Cooler. The results from all the tests were analyzed and tabulated.
Full Speed (1127) Medium Speed (1034) Low Speed (655) Very Low Speed (355) Full Speed (1127) Full Speed (1127) Full Speed (1127) Full Speed (1127)
5.2. Intake air energy efficiency ratio (IA-EER) The intake air energy efficiency ratio was determined by using the following equation:
5. Methodology used for calculation of IEC performance parameters
Intake Air EER (IA − EER), W/W =
The indirect evaporative cooler (IEC) performance parameters were calculated using thermodynamic equations and the data collected from the cooler tests conducted for different environmental conditions mentioned in the Table 2. The following section describes the various parameters used in determining the performance of the IEC.
5.1. Cooling capacity The cooling capacity of the cooler was calculated using the following equation: ˙ sup × Cp (Tdbt,in − Tdbt,sup ) Cooling Capacity (W ) = m
(1)
Cooling Capacity, W Power Input, W
(2)
5.3. Wet-bulb effectiveness In the literature, it is reported that the performance of direct evaporative coolers is compared with their wet-bulb effectiveness. The process of air cooling by direct evaporative cooler is a constant wet-bulb process. The limiting value of supply air dry-bulb temperature by the evaporative cooling process is the wet-bulb temperature of the intake air. This parameter gives an idea how close the dry-bulb temperature of the supply air approaches the wet-bulb temperature of the intake air. Therefore, the wet-bulb effectiveness is defined as: Wet-bulb Effectiveness, % =
Tdbt,in − Tdbt,sup Tdbt,in − Twbt,in
× 100
(3)
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Table 3 Test results of an indirect evaporative cooler (IEC). Test Case #/Parameters Nominal Test Conditions
1 Full Fan Speed
2 3 Medium to Low Fan Speed
4
5 Full Fan Speed
6
7
8
Inlet Dry-Bulb Temperature, ◦ C Relative Humidity, % Inlet Wet-Bulb Temperature, ◦ C Barometric Pressure, kPa Supply fan Speed, rpm Inlet Air properties Dry-Bulb Temperature, ◦ C Wet-Bulb Temperature, ◦ C Dew-point Temperature, ◦ C Humidity Ratio, g H2 O/kg air Relative Humidity, % Wet-Bulb Depression, ◦ C Supply Air Properties Dry-Bulb Temperature, ◦ C Wet-Bulb Temperature, ◦ C Dew-Point Temperature, ◦ C Humidity Ratio, g H2 O/kg air Relative Humidity, % Exhaust Air properties Dry-Bulb Temperature, ◦ C Wet-Bulb Temperature, ◦ C Dew-Point Temperature, ◦ C Humidity Ratio, g H2 O/kg air Relative Humidity, % Power Consumption Voltage, V Current, A Power Factor Power, W
43.90 19.80 24.40 101.11 1127.63
43.90 19.80 24.40 101.08 1033.70
43.90 19.80 24.40 101.06 655.10
43.90 19.80 24.40 101.06 354.67
40.00 27.72 24.40 101.06 1122.25
32.80 56.58 25.60 101.31 1128.00
45.00 17.96 24.40 101.11 1127.79
38.90 45.38 28.30 101.76 1127.88
43.92 24.47 15.95 11.40 19.95 19.45
44.00 24.35 15.64 11.18 19.49 19.65
44.00 24.28 15.49 11.07 19.30 19.72
43.91 24.15 15.27 10.91 19.12 19.76
39.84 24.45 18.05 13.07 28.27 15.39
32.95 25.72 23.17 18.01 56.53 7.23
45.0 24.4 15.2 10.84 17.96 20.58
38.90 28.30 24.90 19.99 45.33 10.60
27.55 19.83 15.95 11.40 49.20
26.95 19.46 15.63 11.17 49.91
25.71 18.93 15.38 10.99 52.87
25.95 18.91 15.21 10.87 51.54
25.08 20.22 17.93 12.97 64.54
26.37 24.05 23.19 18.04 82.71
27.2 19.4 15.4 10.88 48.62
29.80 26.20 25.00 20.09 75.31
35.68 33.46 32.96 32.67 85.97
35.55 33.42 32.94 32.64 86.48
34.35 32.31 31.82 30.55 86.78
37.55 33.84 33.02 32.79 77.85
32.72 30.59 30.04 27.46 85.85
30.97 29.62 29.25 26.13 90.61
34.6 32.9 32.5 31.80 88.96
36.00 32.40 31.54 29.83 78.14
235.02 3.24 0.98 746.40
228.48 2.63 0.98 588.75
240.15 0.87 0.88 182.09
239.37 0.49 0.58 68.25
226.29 3.21 0.98 712.45
226.22 3.32 0.98 735.82
232.54 3.21 0.98 730.58
225.76 3.36 0.98 744.25
5.4. Dew-point effectiveness It is a known fact that in direct evaporative cooling, the maximum depression in the intake air dry-bulb can be achieved is by reaching to its wet-bulb temperature, i.e. 100% wet-bulb effectiveness. However, in indirect evaporative cooling, the intake air is split into two streams as explained in the section of Working Principle of the IEC. The supply air is sensibly cooled (without addition of humidity) while passing through the heat exchanger modules. The limiting value of the supply air is the dew-point of the intake air. So, it is better to compare the performance of the indirect evaporative coolers on the basis of dew-point effectiveness. The dew-point effectiveness can be defined as given in equation below: Dew-point Effectiveness, % =
Tdbt,in − Tdbt,sup Tdbt,in − Tdpt,in
× 100
(4)
5.5. Water evaporation rate In an indirect evaporative cooling process, the supply air is sensibly cooled while the exhaust air is saturated with the moisture. The water evaporation rate can be obtained by using the thermodynamic properties of the intake and exhaust air. The water evaporation rate was calculated using the following relation: Water evaporation rate, kg/hour = exh × Qexh (Uexh − Uin )
(5)
6. Results and discussion The IEC was tested for different environmental conditions. The tests for hot and dry design conditions were carried out at four different fan speeds whereas the other tests were conducted at full fan speed. The tests and performance results of the IEC are presented in Tables 3 and 4, respectively. The state of air-water vapor mixture is represented by three variables, one of which is intrinsically the barometric pressure and the other two are dry bulb and wet bulb temperatures. In Table 3, the other variables (relative humidity,
humidity ratio, dew point temperature) were obtained by using the psychrometric calculator PsyCalc [x] from the above mentioned three measured variables. The IEC performance tests for Cases 1 to 4 were conducted for same test conditions (43.9 ◦ C DBT, 24.4 ◦ C WBT) but for different fan speeds with supply air flow rates varying from 631 to 2388 m3 /h. The other IEC performance tests (Cases 5 to 8) were carried out at full fan speed but for different test conditions. From the test results shown in Table 3, it is important point to note that for the ambient relative humidity in the range of 18 to 28 percent, the IEC provides the supply air with relative humidity in the range of 48.6 to 64.5 percent. The test cases (TC1, TC5, TC6, TC7, and TC8) are plotted on the 1 atmosphere psychrometric chart to understand the thermodynamic process that takes place inside the IEC, Fig. 5. Since the measured values are very close to 1 atmosphere, the process path is projected on the 1 atmosphere psychrometric chart without sacrificing the accuracy. The identical symbols on the psychrometric chart show the inlet air and supply air conditions for the particular test. As expected, both the inlet air and supply air conditions lie on a horizontal line for each of the reported tests. It is because of fact that the air supplied to the conditioned space by IEC is sensibly cooled. It can be seen from Fig. 5 that when the inlet air relative humidity was above 45%, the IEC delivers the supply air to the conditioned space with relative humidity of more than 75%. This shows that the IEC is more effective to provide comfort conditions for hot and dry climate. It can be inferred that from the results that the intake air energy efficiency ratio (IA-EER) of the cooler varies from 7.1 (test case 6) to 55.1 (test case 4) depending on test conditions and fan speed. The power consumption of IEC has been found to vary from 68.3 to 746 watts due to change in fan speed from low to full for the reported tests. It can be noticed from the results that at very low fan speed (supply air equal to 631 m3 /h), the measured power factor of the motor was as low as 0.58 which contributes to low electric power consumption. This explains the reason of IEC having the highest IA–EER at the lowest fan speed, Fig. 6.
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Table 4 Performance results of an indirect evaporative cooler (IEC). Test Case #/Parameters
1
2
3
4
5
6
7
8
Supply Airflow Rate, m3 /h Exhaust Airflow Rate, m3 /h Wet-bulb Effectiveness, % Dew-Point Effectiveness, % Cooling Capacity, kW Intake Air EER, W/W Total Water Consumption, m3 /h Water Evaporation Rate, m3 /h
2388 1757 84.2 58.5 12.8 17.2 0.0598 0.0406
2104 1445 86.8 60.1 11.7 19.9 0.0404 0.0337
1478 899 92.8 63.9 9.0 49.2 0.0285 0.0190
631 618 90.1 62.6 3.8 55.1 0.0160 0.0145
2368 1736 95.9 67.4 11.6 16.3 0.0477 0.0275
2390 1759 91.0 67.4 5.3 7.1 0.0332 0.0159
2389 1758 86.8 60.3 14.0 19.1 0.0598 0.0400
2389 1758 85.2 65.2 7.1 9.6 0.0355 0.0190
Fig. 6. Effect of fan speed on IA-EER for fixed test condition.
Fig. 8. Variation of water evaporation rate with wet-bulb depression at full fan speed.
The effect of wet-bulb depression on the IA-EER is presented in Fig. 7. It can be seen from the figure that IA-EER varies linearly with wet-bulb depression (difference between inlet air DBT and WBT). This means that the greater the difference between the outside air DBT and WBT the higher will be the efficiency of cooler. The water consumption has been found to vary between 0.0160 m3 /h (test case 4) and 0.0598 m3 /h (test case 1). The test results indicate that there is large variation in water evaporation rate depending on the test conditions and fan speed. At full fan speed, an average of 58.7% of the total water consumed by IEC is evaporated and is released to the atmosphere through the exhaust air. The remaining 41.3% of the water which is being drained can be re-used for the operation of IEC. The water evaporation rate is directly dependent on the inlet wet-bulb depression as evident from Fig. 8. The relationship between the water evaporation rate and wet-bulb depression is almost linear. The effect of fan speed Fig. 9. Effect of fan speed on water evaporation rate for fixed test condition.
on the water evaporation was also studied for fixed test conditions and the results are presented graphically in Fig. 9. The water evaporation rate increases with increasing fan speed with water evaporation as low as 0.0145 m3 /h. 7. Conclusions
Fig. 7. Variation of IA-EER with wet-bulb depression at full fan speed.
In the present study, the performance of the indirect evaporative cooler was evaluated for different test conditions at variable fan speed. It can be concluded from the test results that the indirect evaporative cooler provides low cost evaporative cooling without adding moisture to the air supplied to the conditioned space. The indirect evaporative cooling system can achieve wet bulb effectiveness more than 100% because theoretically the limiting value of supply air temperature to the conditioned space is the dew point temperature of the intake air. The test results indicate that intake
A. Ahmad et al. / Energy and Buildings 62 (2013) 278–285
air energy efficiency ratio of indirect evaporative cooler is directly proportional to the wet-bulb depression. The study shows that the indirect evaporative cooler is mostly suitable for hot and dry climatic conditions. In order to improve the existing system, following recommendations are made and being considered in our future work as well: The existing system requires initial soaping of the heat exchanger unit which could be avoided by using different and suitable heat exchanger material. Moreover, the existing system is suitable only for hot and dry climatic conditions. It could also be used in humid conditions as well if the inlet air can be conditioned by suitable air conditioning mechanism such as liquid desiccant system or desiccant wheel. Acknowledgement The authors wish to acknowledge the support of the Research Institute of King Fahd University of Petroleum and Minerals, Dhahran, Saudi Arabia. References [1] Annual Report - Activities and achievement of the authority 2009, Electricity and Cogeneration Regulatory Authority (ECRA), Riyadh, Saudi Arabia, http://www.ecra.gov.sa/reports.aspx (visited the website in Feb 2012). [2] C. Qun, Y. Kangding, W. Moran, P. Ning, G. Zeng-Yuan, A new approach to analysis and optimization of evaporative cooling system I: Theory, Energy 35 (6) (2010) 2448–2454. [3] C. Qun, P. Ning, G. Zeng-Yuan, A new approach to analysis and optimization of evaporative cooling system II: applications, Energy 36 (5) (2011) 2890–2898. [4] ASHRAE Handbook: Refrigeration – SI Edition, American Society of Heating, Refrigeration and Air-Conditioning Engineers, Inc., Atlanta, GA 30329, US, 2006. [5] CIBSE Knowledge Series – Sustainable Low Energy Cooling: An Overview, Latimer Trend & Co. Ltd., Plymouth PL6 7PY, UK, 2005, pp. 15–19. [6] C. Zhan, Z. Duan, X. Zhao, S. Smith, H. Jin, S. Riffat, Comparative study of the performance of the M-cycle counter-flow and cross-flow heat exchangers for indirect evaporative cooling: paving the path toward sustainable cooling of buildings, Energy 36 (2011) 6790–6805. [7] X. Zhao, S. Yang, Z. Duan, S.B. Riffat, Feasibility study of a new dew point air conditioning system for China building application, Building and Environment 44 (2009) 1990–1999. [8] B. Costelloe, D. Finn, Indirect evaporative cooling potential in air–water systems in temperate climates, Energy and Buildings 35 (2003) 573–591. [9] B. Costelloe, D. Finn, Thermal effectiveness characteristics of low approach indirect evaporative cooling systems in buildings, Energy and Buildings 39 (2007) 1235–1243.
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[10] M.J. Holmes, J.N. Hacker, Climate change, thermal comfort and energy: meeting the design challenges of the 21st century, Energy and Buildings 39 (2007) 802–814. [11] S.Th. Smith, V.I. Hanby, C. Harpham, A probabilistic analysis of the future potential of evaporative cooling systems in a temperate climate, Energy and Buildings 43 (2011) 507–516. [12] B. Riangvilaikul, S. Kumar, Numerical study of a novel dew point evaporative cooling system, Energy and Buildings 42 (2010) 2241–2250. [13] B. Costelloea, D. Finn, Indirect evaporative cooling potential in air–water systems in temperate climates, Energy and Buildings 35 (2003) 573–591. [14] A. Hasan, Going below the wet-bulb temperature by indirect evaporative cooling: Analysis using a modified (-NTU method, Applied Energy 89 (1) (2012) 237–245. [15] B. Riangvilaikul, S. Kumar, An experimental study of a novel dew point evaporative cooling system, Energy Buildings 42 (2010) 637–644. [16] H. El-Dessouky, H. Ettouney, A. Al-Zeefari, Performance analysis of two-stage evaporative coolers, Chemical Engineering Journal 102 (3) (2004) 255–266. [17] M.F. Farahani, G. Heidarinejad, S. Delfani, A two-stage system of nocturnal radiative and indirect evaporative cooling for conditions in Tehran, Energy and Buildings 42 (2010) 2131–2138. [18] S.T. Hsu, Z. Lavan, W.M. Worek, Optimization of wet-surface heat exchangers, Energy 14 (11) (1989) 757–770. [19] D.R. Crum, J.W. Mitchell, W.A. Beckman, Indirect evaporative cooler performance, ASHRAE Transactions 93 (1) (1987) 1261–1275. [20] I.L. Maclaine-Cross, P.J. Banks, A general theory of wet surface heat exchangers and its application to regenerative evaporative cooling, Journal of Heat Transfer – Transactions ASME 103 (3) (1981) 579–585. [21] D. Pescod, A heat exchanger for energy saving in an air-conditioning plant, ASHRAE Transactions 85 (2) (1979) 238–251. [22] J. Facao, A.C. Oliveira, Thermal behavior of closed wet cooling-towers for use with chilled ceilings, Applied Thermal Engineering 20 (13) (1999) 1225–1236. [23] J.L. Peterson, An effectiveness model for indirect evaporative coolers, ASHRAE Transactions 99 (2) (1993) 392–399. [24] R.I.T. Mizushina, H. Miyashita, Experimental study of an evaporative cooler, International Journal of Chemical Engineering 7 (4) (1967) 727–732. [25] R.I.T. Mizushina, H. Miyashita, Characteristics and methods of thermal design of evaporative coolers, International Journal of Chemical Engineering 8 (3) (1968) 532–538. [26] L. M.Al-Hadhrani, A., Ahmad, S. Rehman, Hybrid solar air-conditioning system, United States Patent No. US 8,141,379 B2, Issued by the United Sates Patent and Trademark Office (USPTO), 27 March 2012. [27] ANSI/ASHRAE standard 55-2010: Thermal Environmental Conditions for Human Occupancy, American Society for Heating, Refrigerating and AirConditioning Engineers, Inc., 1791 Tullie Circle, NE, Atlanta, GA 30329. [28] British standard BS EN ISO 7730:1995: Modern thermal environmentsDetermination of the PMV and PPD indices and specification of the conditions for thermal comfort. [29] ANSI/ASHRAE Standard 143-2000, Method of test for rating indirect evaporative coolers, American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc., Atlanta, USA, 2000.
Contents lists available at SciVerse ScienceDirect
Energy and Buildings journal homepage: www.elsevier.com/locate/enbuild
Performance evaluation of an indirect evaporative cooler under controlled environmental conditions Aftab Ahmad ∗ , Shafiqur Rehman, Luai M. Al-Hadhrami Center for Engineering Research, Research Institute, King Fahd University of Petroleum and Minerals, Dhahran-31261, Saudi Arabia
a r t i c l e
i n f o
Article history: Received 8 August 2012 Received in revised form 10 March 2013 Accepted 11 March 2013 Keywords: Indirect evaporative cooler Dew-point effectiveness Wet-bulb depression Intake air energy efficiency ratio Water evaporation rate Wet-bulb temperature
a b s t r a c t The study investigated the performance of a 5-ton capacity indirect evaporative cooler under controlled environmental conditions (43.9 ◦ C dry-bulb temperature and 19.9% relative humidity) but for different air flow rates (631 to 2388 m3 /h). The experimental results showed that the intake air energy efficiency ratio of the cooler varied from 7.1 to 55.1 depending on test conditions and air flow rate. The power consumption of indirect evaporative cooler was found to vary from 68.3 to 746 watts. Water consumption was found to vary between 0.0160 and 0.0598 m3 /h. At full fan speed, an average of 58.7% of the total water consumed by indirect evaporative cooler was evaporated. The results indicated that intake air energy efficiency ratio was directly proportional to the wet-bulb depression. The study also showed that the indirect evaporative cooler is suitable for hot and dry climatic conditions. © 2013 Elsevier B.V. All rights reserved.
1. Introduction Kingdom of Saudi Arabia is going through rapid economical growth. As a result of which new and modern infrastructure, which is energy intensive, is taking physical shape. In 2009, Saudi Electricity Company (SEC) delivered a total of 193,472GWh of energy, an increase of 6.8% over the previous year [1]. The number of customers increased over the same period by 5.2% to 5701,516. As shown in Fig. 1, the residential sector consumes more than half of the total electricity sold by SEC and industrial sector consumes around 18%. In Saudi Arabia the major portion of the electricity is consumed to cool buildings, offices, residential houses, and refrigeration storage systems. There are mainly three basic technologies which are used in air cooling process. These include mechanical vapor compression, absorption systems, and evaporative indirect cooling systems. Evaporative cooling is an efficient and sustainable method for cooling. Theoretically, the ultimate goal of evaporative cooling system is to achieve the ambient air wet bulb temperature, which is a bit difficult to be reached practically. In literature, a coefficient of performance (COP) of 8 to 20 has been reported for indirect evaporative cooling systems [2,3] while 2 to 3 for mechanical vapor system and 0.4 to 1.2 for absorption system [4,5]. Besides all goods, indirect evaporative cooling systems have lower cooling capacity resulting in limited temperature reduction [6]. In these systems, the cooling is limited by the
∗ Corresponding author. E-mail address: [email protected] (A. Ahmad). 0378-7788/$ – see front matter © 2013 Elsevier B.V. All rights reserved. http://dx.doi.org/10.1016/j.enbuild.2013.03.013
wet-bulb temperature of the intake air. In real time applications, the temperature of the supply air is usually higher than the inlet wetbulb air temperature [7]. The indirect evaporative cooling systems can be used effectively in in temperate climate and regions situated at high latitudes, as reported by Costelloe and Finn [8,9]. The benefits of this cooling measure are most effective when applied to buildings requiring a relatively low mechanical air flow rate and comparatively high supply air temperature. As mentioned by Holmes and Hacker [10] such criteria are mostly applicable in building designs utilizing cooling methods such as chilled beams and displacement ventilation. Smith et al. [11] reported that evaporative cooling should continue to function as an effective low-energy cooling technique in future, warming climates. Riangvilaikul and Kumar [12] presented theoretical performance of a novel dew point evaporative cooling system operating under dry, moderate and humid climate conditions and the influence of major operating parameters such as velocity, system dimension and the ratio of working air to intake air. In the recent times the evaporative cooling systems have been extensively studied and used for industrial and residential applications due to low energy consumption [13]. Hasan [14] presented a method to achieve sub-wet bulb temperature by indirect evaporative cooling of air without using a mechanical vapor compression machine. The study showed that the proposed model for indirect evaporative coolers can be based on the e-NTU method for sensible heat exchangers with some adjustments made by redefining the potential gradients, transfer coefficient, heat capacity rate parameters and by assuming a
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Nomenclature Cp DBT DPT IA-EER IEC HMX ˙ sup m Qexh Tdbt , Tdbt , sup Tdpt , Uexh Uin WBT
moist air specific heat, J kg−1 K−1 dry-bulb temperature, ◦ C dew-point temperature, ◦ C intake air energy efficiency ratio, W W−1 indirect evaporative cooler heat and mass exchanger supply air mass flow rate, kg s−1 exhaust air flow rate, m3 /h in intake air dry-bulb temperature, ◦ C supply air dry-bulb temperature, ◦ C in intake air dew-point temperature, ◦ C exhaust air humidity ratio, kg moisture (kg dry air)−1 intake air humidity ratio, kg moisture (kg dry air)−1 wet-bulb temperature, ◦ C
Subscripts dbt dry-bulb temperature wbt wet-bulb temperature dew-point temperature dpt in intake exh exhaust sup supply Greek Symbols air density, kg m−3
linear saturation temperature-enthalpy relation of air. The model results showed good agreement with results from experimental measurements and a numerical model. Riangvilaikul and Kumar [15] carried out experimental measurements on a dew point evaporative cooling and found a wet-bulb effectiveness of 92–114%. Experimental results of a two-stage combined indirect-direct evaporative air cooler system showed a wet-bulb effectiveness of 90–120% (El-Dessouky et al. [16]). Farahani et al. [17] investigated the performance of a two-stage cooling system consisting of a nocturnal radiative unit, a cooling coil, and an indirect evaporative cooler for weather conditions of Tehran, Iran. The results demonstrated that the first stage of the system increases the effectiveness of the indirect evaporative cooler. Hsu et al. [18] studied theoretically and experimentally, counter flow, cross flow and regenerative indirect evaporative coolers to achieve dew point temperature of the ambient air. Crum et al. [19] also demonstrated that dew point temperature can be attained by using multistage indirect evaporative cooling and by combining a cooling tower and a heat exchanger. For an appropriate mass flow rate and chosen geometry of the cooler, Maclaine-cross and Banks [20] showed that regenerative evaporative cooling process can attain the dew point temperature. Earlier, Pescod [21] reported that by splitting a portion of the air produced by an indirect evaporative cooler and using it in the wet passage, the target wet-bulb temperature of the cooling process could be lowered. Facao [22] compared simplified models [23,24] with detailed models such as numerical models [25] and noticed that simplified models based on an overall approach provide as good or even better results as those based on finite differences. Al-Hadhrami et al. [26] designed a hybrid solar air-conditioning system using indirect evaporative cooler. The hybrid system uses solar energy and liquid desiccant system to dry or condition the high humidity ambient air before it enters into the indirect evaporative cooler. Therefore, the designed hybrid solar air-conditioning system can be used in even regions having high humidity. The
Fig. 1. Energy consumption by sector for 2009, Saudi Arabia [1].
designed system can work in all climatic conditions. In addition to the air-conditioning, a distilled water recovery system is designed to recover water from the desiccant and from the indirect evaporatively cooled air in the form of distilled water. The present study provides the finding of experimental results of an Indirect Evaporative Cooler (IEC) tested under controlled environmental conditions in the IEC testing chamber. The testing chamber consisted of environmental section (include solid desiccant wheel system, gas-based air heating system, and an air cooler), IEC section, and air conditioned section. The IEC unit was instrumented to measure its performance by using different types of sensors. The data loggers were used to log the data automatically for analysis to determine the performance of the IEC unit. 2. Thermal comfort Thermal comfort is an important concept for climate control systems and beyond. To have “thermal comfort” means that a person wearing a normal amount of clothing feels neither too cold nor too warm. Thermal comfort is important both for one’s wellbeing and for productivity. It can be achieved only when the air temperature, humidity and air movement are within the specified range often referred to as the “comfort zone”. In general thermal comfort is defined as that condition of mind which expresses satisfaction with the thermal environment. According to ASHRAE standard [27], the thermal comfort is defined as that condition of mind which expresses satisfaction with the thermal environment and is assessed by subjective evaluation. According to British standard [28], the thermal comfort is defined as that condition of mind which expresses satisfaction with the thermal environment. The factors which affect the thermal comfort of human beings are classified as environmental and personal. The environmental factors include the air temperature, humidity, air speed and radiant temperature while personal factors include the clothing and personal activity and conditions. ASHRAE [27] has developed an industry consensus standard to describe comfort requirements in buildings. The purpose of this standard is to specify the combinations of indoor thermal environmental factors and personal factors that will produce thermal environmental conditions acceptable to a majority of the occupants within the space. One of the most recognizable features of ASHRAE standard 55-2010 [27] is the ASHRAE comfort zone as portrayed on a psychrometric chart, Fig. 2. The comfort chart shows the acceptable range of operative temperature and humidity for spaces that meet the criteria specified in Section 5.2.1.1 of the standard. The ASHRAE standard allows the comfort charts to be applied to spaces where the occupants have activity levels that result in metabolic rates between 1.0 met and 1.3 met and where clothing is worn
280
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Fig. 2. ASHRAE 2010 comfort chart [27].
that provides between 0.5 clo and 1.0 clo of thermal insulation. However, the comfort zone is shifted to the right or left in the psychrometric chart shown in Fig. 2 depending on the clothing, metabolic rate and radiant temperature.
additional cooler and desiccant wheel dehumidifier were used. The air flow rate to the climatic room was controlled by using the manual and automated damper through the control panel. 3.1. Working principle of the IEC
3. Experimental set-up The performance of a 5-ton capacity Indirect Evaporative Cooler (IEC) for different climatic conditions was measured in the instrumented IEC testing chamber. The Cooler was placed in an environmentally-controlled room in the chamber. The schematic diagram of testing chamber with measurement locations and other equipment is shown in Fig. 3. The testing chamber is equipped with devices for simulation of desired climatic conditions in the environmental section of the testing chamber. The gas fired burner and evaporative cooler were used for heating and humidification of outside air. For dehumidification and cooling of outside air, the
The cross-sectional view of indirect evaporative cooler (IEC) is shown in Fig. 4. Outside fresh air (1) is drawn into the cooler by a variable speed centrifugal fan (2) which forces air through the stack of 2 inch thick filters (3). The filtered air passes through the heat and mass exchange (HMX) modules (4) which split the air into two streams i.e. exhaust (5) and supply (6). The water supplied to the HMX modules is controlled by solenoid valve located on the water supply line. The supply air is sensibly cooled (without addition of humidity) while passing through the HMX modules. About half of the intake air is saturated with water inside the HMX modules and then released to the atmosphere as exhaust air.
Fig. 3. Schematic diagram of the IEC testing chamber.
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over the sampling period. The experiments were carried out by simulating the desired test conditions in the climatic room of the testing chamber and the test data of IEC unit data were automatically logged. The test data were downloaded from the data logger memory for further processing. 4. Test procedure
Fig. 4. Cross-sectional view of IEC unit.
3.2. Measurements and instrumentation The list of parameters measured and instruments used for the performance evaluation of the IEC system are listed in Table 1. All the temperature sensors were calibrated against the Traceable full thermometer from Control Company, USA. The supply and exhaust air flow rates were calibrated against the supply fan RPM by using a Fluke® pitot tube 922 air flow meter. The water flow sensor was calibrated manually by using graduated flask. The voltage and current sensors were calibrated against the Watt-Hour meter. The transducers for the differential and static pressure were calibrated using a Fluke® Pitot tube 922 air flow meter. All the sensors listed in Table 1 were connected to the ACRTM Systems data acquisition system. The data acquisition system consisted of different types of ACRTM data loggers which include 02 smart reader plus 8 eight-channel temperature data loggers, 01 smart reader plus 7 eight-channel process signal data logger, 01 smart reader plus 9 two-channel pulse data logger, and 01 smart reader plus 3 current, voltage and temperature data logger. The above data loggers were programmed for recording experimental data on minute-basis except for the pulse datalogger which was programmed for 2-minute data logging. The data loggers sampled data at every 8 seconds and stored the averaged data in its memory
The cooler was tested for different climatic conditions to evaluate its performance. The ASHRAE standard 143-2000 [29] for indirect evaporative cooler does not give specific test conditions for the evaporative coolers. However, it specifies the arrangement of the equipment, the type of measurements to be taken and accuracy of the equipment. In general, the evaporative coolers are rated mainly in terms of air flow. The test conditions for cooler are listed in Table 2. Prior to conducting test on the cooler, the evaporative media of the cooler was wetted by the soap by turning on/off pump. The purpose of soaping was to breakdown the water repelling films or oils on the evaporative media of the Heat and Mass Exchanger (HMX). The cooling unit was run for more than one hour to wet the HMX completely. The cooler test was conducted as follows: 1. The data acquisition system was started and all display readings were checked for proper functioning of the sensors. 2. The desired test conditions were set in the climatic room by the control panel by operating the air-conditioning equipment. 3. The Cooler was turned ON and allowed to run for at least one to two hours to get the experimental data for the desired test conditions. 4. After completion of the test, the logged data from data acquisition system was downloaded for further processing. 5. To get the next desired test conditions, the climatic room airconditioning was then adjusted by the control panel. The recorded data was averaged over the sampling period as set in the data acquisition system. The averaged recorded values were
Fig. 5. Plot of test cases on the psychrometric chart.
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Table 1 List of parameters measured and instruments used for Cooler testing.
1. 2. 3. 4. 5.
Parameter/Datalogger
Instrument
Range
Accuracy
Intake air dry-bulb temperature Intake air wet-bulb temperature Supply air dry-bulb temperature Supply air wet-bulb temperature Supply air static pressure
ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 Omega® Pressure transmitter PX 653, 1–5 VDC output, 0.5” WC Fluke pitot tube 922 air flow meter ACR Systems Thermistor probes, ET–016 ACR Systems Thermistor probes, ET–016 Fluke® pitot tube 922 air flow meter ACR Systems Thermistor probes, ET–016 Ohio Semitronics voltage and current transducers 3AVT Monarch Instruments Optical Tachometer SPSR-115/230 Dwyer® Instruments multi-jet water meter, pulse type Weathronics Aneroid barometer ACR Systems current, voltage and temperature data logger, 128 kB ACR Systems Reader plus 8 eight-channel temperature data logger ACR Systems Smart Reader Plus 7 eight-channel process signal logger ACR Systems Smart Reader Plus 9 two-channel pulse data logger
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
0–0.5” W.C.
0.25% of full scale
1–80 m/s −35 to 95 ◦ C
±2.5% of reading at 10 m/s ±0.2 ◦ C from 0 to 70 ◦ C
−35 to 95 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
1–80 m/s −35 to 95 ◦ C
±2.5% of reading at 10 m/s ±0.2 ◦ C from 0 to 70 ◦ C
0–300 V
±1.0% of the reading
11.
Supply air flow rate Exhaust air dry-bulb temperature Exhaust air wet-bulb temperature Exhaust air flow rate Outside air dry-bulb temperature Supply fan voltage and current
12.
Fan speed (in RPM)
6. 7. 8. 9. 10.
13.
Water consumption rate
14. 15.
Barometric pressure Datalogger for current and voltage Dataloggers for wet-bulb and dry-bulb air temperatures Datalogger for differential pressure Datalogger for water flow and supply fan RPM
16. 17. 18.
1–250,000 rpm
–
0.28–5 m /h
±2% of full scale
910–1050 mbar 0–5 VDC
0.1 mbar ±0.5% of full scale
−40 to 70 ◦ C
±0.2 ◦ C from 0 to 70 ◦ C
0–5 VDC
±0.5% of full scale
4095 pulses per sample period
±1 pulse per sample period
3
Table 2 Test conditions for the indirect evaporative cooler (IEC). Test Case (TC) # 1. 2. 3. 4. 5. 6. 7. 8.
Test Conditions ◦
Fan Speed (rpm) ◦
DBT 43.9 C, RH 19.80% (WBT 24.4 C) DBT 43.9 ◦ C, RH 19.80% (WBT 24.4 ◦ C) DBT 43.9 ◦ C, RH 19.80% (WBT 24.4 ◦ C) DBT 43.9 ◦ C, RH 19.80% (WBT 24.4 ◦ C DBT 40.0 ◦ C, RH 27.72% (WBT 24.4 ◦ C) DBT 32.8 ◦ C, RH 56.58% (WBT 25.6 ◦ C) DBT 45.0 ◦ C, RH 17.96% (WBT 24.4 ◦ C) DBT 38.9 ◦ C, RH 45.38% (WBT 28.3 ◦ C)
used for performance evaluation of the Cooler. The results from all the tests were analyzed and tabulated.
Full Speed (1127) Medium Speed (1034) Low Speed (655) Very Low Speed (355) Full Speed (1127) Full Speed (1127) Full Speed (1127) Full Speed (1127)
5.2. Intake air energy efficiency ratio (IA-EER) The intake air energy efficiency ratio was determined by using the following equation:
5. Methodology used for calculation of IEC performance parameters
Intake Air EER (IA − EER), W/W =
The indirect evaporative cooler (IEC) performance parameters were calculated using thermodynamic equations and the data collected from the cooler tests conducted for different environmental conditions mentioned in the Table 2. The following section describes the various parameters used in determining the performance of the IEC.
5.1. Cooling capacity The cooling capacity of the cooler was calculated using the following equation: ˙ sup × Cp (Tdbt,in − Tdbt,sup ) Cooling Capacity (W ) = m
(1)
Cooling Capacity, W Power Input, W
(2)
5.3. Wet-bulb effectiveness In the literature, it is reported that the performance of direct evaporative coolers is compared with their wet-bulb effectiveness. The process of air cooling by direct evaporative cooler is a constant wet-bulb process. The limiting value of supply air dry-bulb temperature by the evaporative cooling process is the wet-bulb temperature of the intake air. This parameter gives an idea how close the dry-bulb temperature of the supply air approaches the wet-bulb temperature of the intake air. Therefore, the wet-bulb effectiveness is defined as: Wet-bulb Effectiveness, % =
Tdbt,in − Tdbt,sup Tdbt,in − Twbt,in
× 100
(3)
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Table 3 Test results of an indirect evaporative cooler (IEC). Test Case #/Parameters Nominal Test Conditions
1 Full Fan Speed
2 3 Medium to Low Fan Speed
4
5 Full Fan Speed
6
7
8
Inlet Dry-Bulb Temperature, ◦ C Relative Humidity, % Inlet Wet-Bulb Temperature, ◦ C Barometric Pressure, kPa Supply fan Speed, rpm Inlet Air properties Dry-Bulb Temperature, ◦ C Wet-Bulb Temperature, ◦ C Dew-point Temperature, ◦ C Humidity Ratio, g H2 O/kg air Relative Humidity, % Wet-Bulb Depression, ◦ C Supply Air Properties Dry-Bulb Temperature, ◦ C Wet-Bulb Temperature, ◦ C Dew-Point Temperature, ◦ C Humidity Ratio, g H2 O/kg air Relative Humidity, % Exhaust Air properties Dry-Bulb Temperature, ◦ C Wet-Bulb Temperature, ◦ C Dew-Point Temperature, ◦ C Humidity Ratio, g H2 O/kg air Relative Humidity, % Power Consumption Voltage, V Current, A Power Factor Power, W
43.90 19.80 24.40 101.11 1127.63
43.90 19.80 24.40 101.08 1033.70
43.90 19.80 24.40 101.06 655.10
43.90 19.80 24.40 101.06 354.67
40.00 27.72 24.40 101.06 1122.25
32.80 56.58 25.60 101.31 1128.00
45.00 17.96 24.40 101.11 1127.79
38.90 45.38 28.30 101.76 1127.88
43.92 24.47 15.95 11.40 19.95 19.45
44.00 24.35 15.64 11.18 19.49 19.65
44.00 24.28 15.49 11.07 19.30 19.72
43.91 24.15 15.27 10.91 19.12 19.76
39.84 24.45 18.05 13.07 28.27 15.39
32.95 25.72 23.17 18.01 56.53 7.23
45.0 24.4 15.2 10.84 17.96 20.58
38.90 28.30 24.90 19.99 45.33 10.60
27.55 19.83 15.95 11.40 49.20
26.95 19.46 15.63 11.17 49.91
25.71 18.93 15.38 10.99 52.87
25.95 18.91 15.21 10.87 51.54
25.08 20.22 17.93 12.97 64.54
26.37 24.05 23.19 18.04 82.71
27.2 19.4 15.4 10.88 48.62
29.80 26.20 25.00 20.09 75.31
35.68 33.46 32.96 32.67 85.97
35.55 33.42 32.94 32.64 86.48
34.35 32.31 31.82 30.55 86.78
37.55 33.84 33.02 32.79 77.85
32.72 30.59 30.04 27.46 85.85
30.97 29.62 29.25 26.13 90.61
34.6 32.9 32.5 31.80 88.96
36.00 32.40 31.54 29.83 78.14
235.02 3.24 0.98 746.40
228.48 2.63 0.98 588.75
240.15 0.87 0.88 182.09
239.37 0.49 0.58 68.25
226.29 3.21 0.98 712.45
226.22 3.32 0.98 735.82
232.54 3.21 0.98 730.58
225.76 3.36 0.98 744.25
5.4. Dew-point effectiveness It is a known fact that in direct evaporative cooling, the maximum depression in the intake air dry-bulb can be achieved is by reaching to its wet-bulb temperature, i.e. 100% wet-bulb effectiveness. However, in indirect evaporative cooling, the intake air is split into two streams as explained in the section of Working Principle of the IEC. The supply air is sensibly cooled (without addition of humidity) while passing through the heat exchanger modules. The limiting value of the supply air is the dew-point of the intake air. So, it is better to compare the performance of the indirect evaporative coolers on the basis of dew-point effectiveness. The dew-point effectiveness can be defined as given in equation below: Dew-point Effectiveness, % =
Tdbt,in − Tdbt,sup Tdbt,in − Tdpt,in
× 100
(4)
5.5. Water evaporation rate In an indirect evaporative cooling process, the supply air is sensibly cooled while the exhaust air is saturated with the moisture. The water evaporation rate can be obtained by using the thermodynamic properties of the intake and exhaust air. The water evaporation rate was calculated using the following relation: Water evaporation rate, kg/hour = exh × Qexh (Uexh − Uin )
(5)
6. Results and discussion The IEC was tested for different environmental conditions. The tests for hot and dry design conditions were carried out at four different fan speeds whereas the other tests were conducted at full fan speed. The tests and performance results of the IEC are presented in Tables 3 and 4, respectively. The state of air-water vapor mixture is represented by three variables, one of which is intrinsically the barometric pressure and the other two are dry bulb and wet bulb temperatures. In Table 3, the other variables (relative humidity,
humidity ratio, dew point temperature) were obtained by using the psychrometric calculator PsyCalc [x] from the above mentioned three measured variables. The IEC performance tests for Cases 1 to 4 were conducted for same test conditions (43.9 ◦ C DBT, 24.4 ◦ C WBT) but for different fan speeds with supply air flow rates varying from 631 to 2388 m3 /h. The other IEC performance tests (Cases 5 to 8) were carried out at full fan speed but for different test conditions. From the test results shown in Table 3, it is important point to note that for the ambient relative humidity in the range of 18 to 28 percent, the IEC provides the supply air with relative humidity in the range of 48.6 to 64.5 percent. The test cases (TC1, TC5, TC6, TC7, and TC8) are plotted on the 1 atmosphere psychrometric chart to understand the thermodynamic process that takes place inside the IEC, Fig. 5. Since the measured values are very close to 1 atmosphere, the process path is projected on the 1 atmosphere psychrometric chart without sacrificing the accuracy. The identical symbols on the psychrometric chart show the inlet air and supply air conditions for the particular test. As expected, both the inlet air and supply air conditions lie on a horizontal line for each of the reported tests. It is because of fact that the air supplied to the conditioned space by IEC is sensibly cooled. It can be seen from Fig. 5 that when the inlet air relative humidity was above 45%, the IEC delivers the supply air to the conditioned space with relative humidity of more than 75%. This shows that the IEC is more effective to provide comfort conditions for hot and dry climate. It can be inferred that from the results that the intake air energy efficiency ratio (IA-EER) of the cooler varies from 7.1 (test case 6) to 55.1 (test case 4) depending on test conditions and fan speed. The power consumption of IEC has been found to vary from 68.3 to 746 watts due to change in fan speed from low to full for the reported tests. It can be noticed from the results that at very low fan speed (supply air equal to 631 m3 /h), the measured power factor of the motor was as low as 0.58 which contributes to low electric power consumption. This explains the reason of IEC having the highest IA–EER at the lowest fan speed, Fig. 6.
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Table 4 Performance results of an indirect evaporative cooler (IEC). Test Case #/Parameters
1
2
3
4
5
6
7
8
Supply Airflow Rate, m3 /h Exhaust Airflow Rate, m3 /h Wet-bulb Effectiveness, % Dew-Point Effectiveness, % Cooling Capacity, kW Intake Air EER, W/W Total Water Consumption, m3 /h Water Evaporation Rate, m3 /h
2388 1757 84.2 58.5 12.8 17.2 0.0598 0.0406
2104 1445 86.8 60.1 11.7 19.9 0.0404 0.0337
1478 899 92.8 63.9 9.0 49.2 0.0285 0.0190
631 618 90.1 62.6 3.8 55.1 0.0160 0.0145
2368 1736 95.9 67.4 11.6 16.3 0.0477 0.0275
2390 1759 91.0 67.4 5.3 7.1 0.0332 0.0159
2389 1758 86.8 60.3 14.0 19.1 0.0598 0.0400
2389 1758 85.2 65.2 7.1 9.6 0.0355 0.0190
Fig. 6. Effect of fan speed on IA-EER for fixed test condition.
Fig. 8. Variation of water evaporation rate with wet-bulb depression at full fan speed.
The effect of wet-bulb depression on the IA-EER is presented in Fig. 7. It can be seen from the figure that IA-EER varies linearly with wet-bulb depression (difference between inlet air DBT and WBT). This means that the greater the difference between the outside air DBT and WBT the higher will be the efficiency of cooler. The water consumption has been found to vary between 0.0160 m3 /h (test case 4) and 0.0598 m3 /h (test case 1). The test results indicate that there is large variation in water evaporation rate depending on the test conditions and fan speed. At full fan speed, an average of 58.7% of the total water consumed by IEC is evaporated and is released to the atmosphere through the exhaust air. The remaining 41.3% of the water which is being drained can be re-used for the operation of IEC. The water evaporation rate is directly dependent on the inlet wet-bulb depression as evident from Fig. 8. The relationship between the water evaporation rate and wet-bulb depression is almost linear. The effect of fan speed Fig. 9. Effect of fan speed on water evaporation rate for fixed test condition.
on the water evaporation was also studied for fixed test conditions and the results are presented graphically in Fig. 9. The water evaporation rate increases with increasing fan speed with water evaporation as low as 0.0145 m3 /h. 7. Conclusions
Fig. 7. Variation of IA-EER with wet-bulb depression at full fan speed.
In the present study, the performance of the indirect evaporative cooler was evaluated for different test conditions at variable fan speed. It can be concluded from the test results that the indirect evaporative cooler provides low cost evaporative cooling without adding moisture to the air supplied to the conditioned space. The indirect evaporative cooling system can achieve wet bulb effectiveness more than 100% because theoretically the limiting value of supply air temperature to the conditioned space is the dew point temperature of the intake air. The test results indicate that intake
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air energy efficiency ratio of indirect evaporative cooler is directly proportional to the wet-bulb depression. The study shows that the indirect evaporative cooler is mostly suitable for hot and dry climatic conditions. In order to improve the existing system, following recommendations are made and being considered in our future work as well: The existing system requires initial soaping of the heat exchanger unit which could be avoided by using different and suitable heat exchanger material. Moreover, the existing system is suitable only for hot and dry climatic conditions. It could also be used in humid conditions as well if the inlet air can be conditioned by suitable air conditioning mechanism such as liquid desiccant system or desiccant wheel. Acknowledgement The authors wish to acknowledge the support of the Research Institute of King Fahd University of Petroleum and Minerals, Dhahran, Saudi Arabia. References [1] Annual Report - Activities and achievement of the authority 2009, Electricity and Cogeneration Regulatory Authority (ECRA), Riyadh, Saudi Arabia, http://www.ecra.gov.sa/reports.aspx (visited the website in Feb 2012). [2] C. Qun, Y. Kangding, W. Moran, P. Ning, G. Zeng-Yuan, A new approach to analysis and optimization of evaporative cooling system I: Theory, Energy 35 (6) (2010) 2448–2454. [3] C. Qun, P. Ning, G. Zeng-Yuan, A new approach to analysis and optimization of evaporative cooling system II: applications, Energy 36 (5) (2011) 2890–2898. [4] ASHRAE Handbook: Refrigeration – SI Edition, American Society of Heating, Refrigeration and Air-Conditioning Engineers, Inc., Atlanta, GA 30329, US, 2006. [5] CIBSE Knowledge Series – Sustainable Low Energy Cooling: An Overview, Latimer Trend & Co. Ltd., Plymouth PL6 7PY, UK, 2005, pp. 15–19. [6] C. Zhan, Z. Duan, X. Zhao, S. Smith, H. Jin, S. Riffat, Comparative study of the performance of the M-cycle counter-flow and cross-flow heat exchangers for indirect evaporative cooling: paving the path toward sustainable cooling of buildings, Energy 36 (2011) 6790–6805. [7] X. Zhao, S. Yang, Z. Duan, S.B. Riffat, Feasibility study of a new dew point air conditioning system for China building application, Building and Environment 44 (2009) 1990–1999. [8] B. Costelloe, D. Finn, Indirect evaporative cooling potential in air–water systems in temperate climates, Energy and Buildings 35 (2003) 573–591. [9] B. Costelloe, D. Finn, Thermal effectiveness characteristics of low approach indirect evaporative cooling systems in buildings, Energy and Buildings 39 (2007) 1235–1243.
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