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Theoretical and experimental investigation of the performance of a desiccant air-conditioning system
Renewable Energy 35 (2010) 1368–1375
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Renewable Energy journal homepage: www.elsevier.com/locate/renene
Theoretical and experimental investigation of the performance of a desiccant air-conditioning system G. Panaras a, *, E. Mathioulakis a, V. Belessiotis a, N. Kyriakis b a b
Solar & Other Energy Systems Laboratory, NCSR ‘‘DEMOKRITOS’’, 15310, Agia Paraskevi Attikis, Greece Mechanical Engineering Department, Process Equipment Design Laboratory, Aristotle University of Thessaloniki, Greece
a r t i c l e i n f o
a b s t r a c t
Article history: Received 25 May 2009 Accepted 20 September 2009 Available online 25 November 2009
Solid desiccant air-conditioning systems present an interesting alternative solution with regard to the conventional vapour compression systems, in the sense that they do not use any refrigerants and they present the opportunity to exploit thermal energy, and more specifically solar thermal energy, instead of electrical energy. In the present work a theoretical model is presented for the operation of a desiccant air-conditioning system, developed on the basis of existing approaches for the modelling of the main subsystems of such a device. The model is experimentally validated on a real scale system, through the exploitation of a significant number of measurements, which correspond to a typical range of operation conditions for these systems. The proposed model is used for the investigation of the performance of a system with a typical set-up, examining the influence of parameters such as the weather conditions, the level of the imposed cooling load, the efficiency level of the main subsystems of the set-up, the air flow rate and the regeneration temperature. The results confirm the potential of the examined technology to satisfy actual cooling loads, and at the same time they lead to specific conclusions for the operation of these systems. Ó 2009 Elsevier Ltd. All rights reserved.
Keywords: Solar air-conditioning Desiccant
1. Introduction Solid desiccant air-conditioning systems can be an interesting alternative approach with regard to the vapour compression systems, considering the environmental friendly operation, as well as the performance level of the proposed systems. In principle air-conditioning concerns not only the regulation of temperature, but the humidity as well. The conventional air-conditioning systems do not usually cope with the issue of humidity requirements, and in cases they do, they attempt to control the humidity through the deep cooling of the processed air, in temperatures below its dew point. The moisture condenses and the dehumidified air, which satisfies the humidity requirements of the conditioned space, is heated to the desired temperature, in order to satisfy the temperature requirements of the conditioned space as well. In principle, the control of humidity through temperature, described above, can be characterized as an energy consuming procedure and in some cases it cannot ensure the achievement of the temperature and humidity levels required by the user [1]. On the contrary, desiccant air-conditioning systems achieve the dehumidification of air on a direct way, through the use of the
* Corresponding author. Tel.: þ30 210 650 3810; fax: þ30 210 654 4592. E-mail address: [email protected] (G. Panaras). 0960-1481/$ – see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.renene.2009.11.011
desiccant material, providing processed air of high cooling potential due to its low humidity. The processed air can be evaporatively cooled, enabling the simultaneous treatment of sensible and latent loads [2,3], even if limitations as regards the effectiveness of this treatment might occur, especially in fairly hot and wet climates [4]. Desiccant air-conditioning systems do not use harmful to the environment refrigerants, neither do they consume electrical energy for the cooling process, but thermal energy, in temperature levels that favour the use of solar energy, and more specifically flat plate solar collectors. The latter can be considered as one of the main advantages of desiccant systems, pointing out that cooling loads and solar radiation follow similar time patterns. Thus the majority of research approaches refer to systems connected to a solar thermal energy source. The technology of solid desiccant air-conditioning systems presents positive prospects as regards dissemination in the market. In Europe, the respective bibliography refers to sixteen installed research or demonstration plants, which consider systems connected to solar thermal energy source [5]. Respectively, in the USA a significant commercial activity is reported, mainly for systems being supplied by conventional heating sources, installed as early as the beginning of 80’s. Nevertheless, this activity declined after 2000, mainly due to economic reasons [6]. Proceeding to a general assessment of the technological maturity and the emerging research issues of the desiccant air-conditioning
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
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Nomenclature
(UA)
a COP Cp C F1, F2 h SHF _ m N
Greek letters r density [kg/m3]
h hF1, hF2 NTU s t w Qb Qld Qreg
humidifier degree of operation [-] system coefficient of performance [-] air specific heat capacity [kJ/kgK] heat capacity rate [kW/K] combined potentials [-] enthalpy [kJ/kg] sensible heat factor [-] air flow rate [m3/s] total number of measurements [-] efficiency [-] efficiency of the desiccant wheel with regard to combined potentials F1, F2 [-] number of transfer units standard deviation [-] air temperature [ C] air absolute humidity rate [g /kg] cooling load of the building [kW] cooling power provided to the space by the system [kW] thermal power for the regeneration of the wheel [kW]
systems, as expressed in the relevant literature, the performance level of the involved subsystems is considered satisfactory, even though research efforts continue on this particular issue. The basic research aspect for this technology is the standardization of the proposed systems, including amongst others the solution of problems related to the successful integration of the subsystems to the total installation and the development of packaged systems, as well as the development of design methodologies [5–7]. The connection to a solar thermal energy heating source increases the complexity of the system, thus it stresses the importance of the discussed aspect of standardization. The development of a simple and reliable model for the analysis of the performance of the systems can lead to useful conclusions as regards the operation characteristics of the systems, and the effect of the various parameters. This model can be used as the kernel of a design methodology, since it allows the optimization of the system design, with regard to the specific requirements of a given application. Such a model based on existing approaches for the modelling of the main subsystems, is proposed in the present work. The operation of the subsystems is described by specific efficiency factors, and the validity of the assumption for these factors presenting constant values is experimentally investigated on an actual installation, for a range of typical conditions for the air-conditioning applications.
overall heat transfer coefficient times area [kW/K]
Subscripts cal simulated values (of temperature or absolute humidity) des desiccant wheel e exhaust stream hx heat exchanger hum humidifier i ¼ 1,2,3 index referring to various positions of the desiccant system j ¼ 1.N index referring to the number of measurement performed lat latent (load) reg regeneration (temperature) s supply stream sat saturated air sens sensible (load) wb wet bulb (temperature)
cooler, where it is cooled and humidified to state (6). This air can cool the hot and dry air, processed by the wheel, through the air-toair heat exchanger and exits the heat exchanger at state (7). It is then heated to the regeneration temperature, at state (8). It then passes through the wheel, where it is cooled and humidified to state (9), and finally is exhausted to the environment. It should be noted, that the cycle just described is referred in bibliography as the ventilation cycle, as it enables 100% renewal of the air of the conditioned space. In the relevant bibliography the recirculation cycle is also referred to. Through the operation of the recirculation cycle no renewal of the air of the conditioned space takes place. In practice, as regards actual installations, a combination of the two cycles is proposed, thus enabling the renewal of the air on a specific percentage, satisfying the hygienic requirements of the conditioned space [10]. Variations of the above described basic desiccant air-conditioning system do exist, mainly concerning the use of the evaporative cooler for the cooling of the supply air, the potential use of a conventional cooling coil aiming at the increase of the system sensible load capacity, and the selection of the regenerative heating source [10]. Other solutions include the staged dehumidification [11], the use of two desiccant wheels [12], or the staged regeneration of the wheel [2]. In any case, the analysis of the alternative configurations is rather extensive and exceeds the scope of this study which would concentrate on the analysis of the performance of a typical system, as the one presented in Fig. 1.
2. Desiccant technology in air-conditioning applications A complete desiccant air-conditioning system consists of a desiccant wheel (DW), two evaporative coolers (EC) and an air-toair heat exchanger (HX). A typical set-up is presented in Fig. 1 [7–9]. According to the systems’ operation principle, depicted in the psychrometric chart of Fig. 2, the following processes can be identified: on the supply stream, processed ambient air of state (1) is dehumidified and heated through the desiccant wheel. Air of state (2), enters the heat exchanger, where it is cooled (state 3). It then passes through the evaporative cooler, where it is cooled and humidified, entering the building at state (4). On the exhaust stream, air from the building, on the obtained thermal comfort conditions of state (5), enters the exhaust stream evaporative
DW
EC 4
1
2
3
Supply air Building Exhaust air 5
6 CE
8
7 HX
Environment
9
Heat source
Fig. 1. Typical desiccant air-conditioning system.
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G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
0.035
hhx ¼
0.03
w [g/kg]
0.025 0.02
9
0.015
6
7 5
0.01
8
1
4 0.005 3 0
0
20
2 40
60
80
t [oC] Fig. 2. Presentation of the desiccant air-conditioning cycle on the psychrometric chart.
In the relevant literature, experimental approaches appear, studying the effect of various parameters on the performance of a solar desiccant air-conditioning system [13,14], even though the studies referring to actual scale systems are rather limited [15]. Certain researchers concentrate on the study of the behaviour of the desiccant wheel [16,17] or on the investigation for innovative concepts/materials, as regards the wheel [18,19]. In the field of theoretical studies, the issue of the desiccant wheel modelling has been studied extensively [20]. As regards the system as a whole, the various approaches mainly investigate the effect of various parameters to the performance [7,8,21], and assess the potential of the examined technology to cope with the imposed cooling loads for various types of buildings and for different geographical areas as well [9,15,22–24].
According to the NTU approach, the efficiency is a function of the NTU. It should be noted, however, that within the NTU range of values, there is a considerable range for which the efficiency remains practically constant. Given the dependence of NTU on (UA) and the fact that in the relevant heat transfer analysis theory, the (UA) is assumed to be constant with regard to operation conditions, including flow rate and rotational speed (for the case of rotary heat exchangers), it can be assumed that efficiency remains constant over a specific range of operation conditions [25]. The present work experimentally investigates the validity of this assumption, for a range of operation conditions relevant to the conditions of the experiment, and would proceed to the determination of the proposed value of the efficiency. It should be noted that most of the theoretical studies on the performance of desiccant air-conditioning systems reported (e.g. [7,9]), follow the approach of constant efficiency in modelling the air-to-air heat exchanger of the system. 3.3. Evaporative cooler The most commonly used type of evaporative cooler in desiccant air-conditioning systems is the evaporative pad humidifier [2,3]. During the operation of the humidifier, the water in the surface of the pad is evaporated and is absorbed by the air stream. If there are no thermal losses to the environment, the process can be considered to be adiabatic and according to the relevant theory the wet bulb temperature of air remains steady [26]. In this case a useful quantity for expressing the performance of the humidifier is the efficiency in terms of temperature [27,28], (the indices refer to the case of humidifier installed at the exhaust stream, Fig. 1):
hhum;t ¼
hhum;w ¼
3.1. Complete system model As already described, a desiccant air-conditioning system consists of a number of subsystems. The theoretical analysis which is presented in this section aims at the selection of specific models for the subsystem operation. According to the set-up of a desiccant system, the input of the one subsystem is the output of the other. Thus the proposed subsystem models have to be appropriately integrated to the model describing the operation of the complete system. Such an approach has been adopted at present work, for a system referring to a typical set-up, as the one presented in Fig. 1. 3.2. Air-to-air heat exchanger For the description of the operation of the air-to-air heat exchanger, the efficiency-NTU approach is adopted, where NTU refers to the Number of Transfer Units, given by [25]:
UA Cmin
(2)
t5 t6 t5 twb
(3a)
Also the efficiency in terms of humidity can be expressed as [26]:
3. Theoretical approach of subsystems performance – the proposed system model
NTU ¼
t2 t3 t t7 ¼ 6 t2 t6 t6 t2
(1)
In the case of balanced flow heat exchanger, the efficiency can be calculated by (the indices refer to the positions of subsystems in a typical system as the one presented in Fig. 1) [25]:
w5 w6 w5 wsat
(3b)
For the case of the adiabatic process, both quantities of nhum,t and nhum,w are equal, as it can easily be concluded through the application of basic psychrometric analysis. According to Camargo et al. [27], an efficiency-NTU relation can be expressed for the nhum,t. This relation is of the exponential type, and it is analogous to the relation for the heat exchanger case. Thus, the assumption of constant value for the efficiency can be also valid in the humidifier case. The experimental investigation would study the validity of this assumption, for the range of operation conditions relevant to the conditions of the experiment, and the humidifier efficiency will be determined. As in the case of heat exchangers, most of the theoretical studies on the performance of desiccant air-conditioning systems use the constant efficiency model for the performance of the humidifier [7,9]. 3.6. Desiccant wheel For the desiccant wheel, the approach of Maclaine-Cross and Banks has been selected [29], which is mentioned in the bibliography as analogy theory. This approach models the dehumidification process, a combined heat and mass process, in analogy with a simple heat transfer process. The analogy is expressed through
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
the combined potentials F1, F2. The combined potentials have resulted through the analysis of heat and mass transfer equations in an elementary segment of the desiccant wheel and their use significantly simplifies the relevant equations, establishing the respective analogy. The discussion for the nature and the role of combined potentials is rather extensive and complicated. For the scope of the current analysis, it can be stated that constant F1 lines coincide with constant enthalpy lines, while constant F2 lines coincide with constant relative humidity lines on the psychrometric chart [30]. The combined potential functions are dependent on the temperature and relative humidity of the processed air, as well as on the thermophysical properties of the processed air and of the wheel context, including the desiccant material [31]. According to the analogy theory, the efficiency indices of the wheel hF1, hF2 with regard to the combined potentials F1, F2 can be calculated in analogy to the energy efficiency of a respective heat exchanger. Given the values of the efficiency indices of the wheel hF1, hF2, the temperature and humidity of the processed air at the outlet of the wheel could be calculated, provided a relation explicitly describing the F1, F2 dependence on temperature and humidity is available. Jurinak has expressed such a relation for the working pair of air-silica gel , leading to the formulation of the model presented through Eqs.(4) and (5) (i ¼ 1,2,8 according to the system presented in Fig. 1) [30]:
F1;i ¼
2865 ðti þ 273:15Þ
1:49
w 0:8624 i 1000
þ 4:344
(4a)
F2;i ¼
w 0:07969 ðti þ 273:15Þ1:49 i 1:127 6360 1000
(4b)
hF1 ¼
F1;2 F1;1 F1;9 F1;1
(5a)
F2;2 F2;1 ¼ F2;9 F2;1
(5b)
hF2
It is useful to note that the efficiency factor hF1 expresses the degree the process approximates the adiabatic one, while the hF2 expresses the degree of dehumidification, [30]. The pair [hF1, hF2] ¼ [0,1] corresponds to an ideal process, which is adiabatic, with the maximum dehumidification level for the respective geometry and flow conditions. The present work experimentally investigates the assumption of constant efficiency factors, in the range of operation conditions relevant to the conditions of the experiment and calculates these values for the wheel used in the specific installation. 4. Experimental investigation
1371
30 cm. The air-to-air heat exchanger used is of the counterflow plate type. The selection of these subsystems was based on their performance level, availability and cost. Especially for the heat exchanger, it should be noted that the air flow rate of the set-up is rather low, thus enabling the use of a plate device, instead of a rotary one which is usually used in desiccant systems. For the detailed and reliable recording of the experimental data, a properly designed measuring set-up was developed. The temperature was measured in all the positions 1–9 in Fig. 1, while absolute humidity has been measured in positions 1, 4, 5, 6, 9. It should be noted that the temperature and humidity field at the outlet of the wheel on the supply stream is not homogeneous. The same is observed at other positions of the system as well, as the non-homogenous temperature distribution is transferred through the operation of the heat exchanger. Thus, and in order to increase the accuracy of temperature measurement, four sensors have been installed at the outlet of the wheel, as well as at other specific positions of the system, the value of temperature used in calculation being the average of the four sensors. Regarding the humidity measurements, and due to the configuration of the sensors used, the installation of more than one is difficult. It was decided therefore to measure the humidity in positions where the field is homogenous. For the study of the performance of the wheel, this position is at the outlet of the supply humidifier (i ¼ 4, Fig. 1). It should be noted that during the study of the behaviour of the wheel, this humidifier was not in operation. According to the technical specifications of the sensors used, the accuracy of measurements is estimated to be 1 C for the temperature, 0.2 g/kg for absolute humidity rate and 50 m3/h for the air flow rate. The presented values are considered acceptable, with regard to the scope of this work. 4.2. Measurement conditions The parameters which can be regulated at the experimental setup, defining the operating conditions of the system, are the following: - the supply and the exhaust air flow rate. In this study, only the balanced flow system is investigated - the desiccant wheel rotational speed. In this application, the speed proposed by the manufacturer (6 rph) was used. - the regeneration temperature of the desiccant wheel - the operation or not of the exhaust and supply stream humidifier Table 1 presents the range of measurement conditions applied at the experimental investigation. The operation of the system was monitored on a daily basis, under steady state conditions, according to different scenarios.
4.1. Experimental set-up
4.3. Experimental validation of the subsystem models
The experimental set-up refers to a typical desiccant airconditioning system, as the one presented in Fig. 1. The system has been developed in the premises of Solar & Other Energy Systems Laboratory of NCSR DEMOKRITOS. The system is designed for the 450–1800 m3/h air flow rate range and it is connected to a small building, acting as a test-cell, of conditioned space surface of 48 m2 and volume of 150 m3. The desiccant wheel is 630 mm in diameter, thickness of 200 mm with silica gel. The evaporative cooler is of the evaporative pad type, with a useful surface of 60 60 cm and pad thickness of
The experimental validation investigates whether the operation of the main subsystems can be effectively described by the
Table 1 Measurement conditions for the experimental study.
MIN MAX Flow rate (m3/h)
t1( C)
w1(g/kg)
t8( C)
w8(g/kg)
24 40
4 14 600
50 80 1000
3 15 1200
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
RMSEt ¼
RMSEw ¼
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uP 2 u N t j ¼ 1 tj2 tj2;cal
(6a)
N vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uP 2 u N t j ¼ 1 wj2 wj2;cal
(6b)
N
For the experimental analysis of this work, the RMSE values in terms of temperature and absolute humidity are presented in Table 3 for all subsystems. For the desiccant wheel case, the measured and simulated values for temperature and absolute humidity are presented in Fig. 3a and b respectively. As regards humidity, the agreement is very good, while in the temperature case a deviation of the order of 2 C between the compared values is observed. The same conclusions can be drawn through the examination of the RMSE values of Table 3. Considering the uncertainties of the measurements implemented at an actual system, as well as the fact that in the respective air-conditioning applications the humidity factor is of higher importance, it can be stated that the above results demonstrate the effectiveness of the proposed approach for the wheel. Regarding the humidifier and the heat exchanger, it can be concluded from the respective RMSE values (Table 3) that the agreement of the measured and simulated values for both systems is very good, thus proving the effectiveness of the proposed approach. It should be noted that in the humidifier case, there is a significant difference between the determined values of efficiency in terms of temperature and humidity. This can be attributed to the presence of thermal losses of the actual evaporative cooling process, imposing deviation from the adiabatic one. The existence Table 2 Efficiency factors for the subsystems as determined through the experimental analysis.
Heat exchanger Humidifier
Desiccant wheel
N
hhx/hhum,t
s (hhx/hhum,t)
hhum,w
s (hhum,w)
141 79
0.85 0.65
0.02 0.03
– 0.79
– 0.05
N
hF1
s (hF1)
107
0.15
0.04
hF2 0.69
s (hF2) 0.06
Table 3 RMSE values in terms of temperature and absolute humidity for the effectiveness of the examined models for the subsystems.
Heat exchanger Humidifier Desiccant wheel
RMSEt ( C)
RMSEw (g/kg)
0.34 0.31 1.24
– 0.26 0.30
of thermal losses is connected to the metallic surfaces of the humidifier, allowing heat to be transferred to the environment, as well as to the sensible heat transfer through the basin for the recycling of the supplied water. Given that the specific phenomena can be eliminated through proper actions for the needs of the performance analysis which follows on, a common efficiency index for both temperature and humidity (nhum) can be assumed, presenting the value determined for nhum,w. 4.4. Experimental validation of the complete system model The validation of the proposed model for the complete system has been performed through the comparison of experimental and simulated values of temperature and absolute humidity of the air supplied to the building by the system (t5, w5) during steady state operation periods. The experimental and simulated values for temperature and absolute humidity are presented in Fig. 4a and b respectively. The
a
simulated temperature values [°C]
proposed models with constant values of the selected efficiency factors, not depended on the operating conditions. For the validation of the above assumption, and for each subsystem, the average value of the selected efficiency factors is calculated, for the total number of available measurements. This value together with the standard deviation is presented in Table 2 for all subsystems. The values of standard deviation are rather low, suggesting the validity of the discussed assumption. According to the examined model and the efficiency factors determined, the temperature, in the case of all subsystems, and the absolute humidity, in the case of the humidifier and the desiccant wheel, at the outlet of the examined subsystem can be calculated, under the operating conditions of each measurement. The so calculated values would be referred hereafter as ‘‘simulated values’’. Thus, N pairs of measured and simulated values of temperature and/or absolute humidity are available, where N denotes the number of measurements for each subsystem. The validation of the examined models is based on the comparison between the measured and simulated values. The Root Mean Standard Error (RMSE) is used as a quantitative index for the comparison of experimental and simulated values, [32]. RMSE is calculated by the following equations:
60 55 50 45 40 35 30 30
35
40
45
50
55
60
10
12
measured temperature values [°C]
b simulated abs.humidity values [g/kg]
1372
12 10 8 6 4 2 0 0
2
4
6
8
measured abs. humidity values [g/kg] Fig. 3. (a) Measured and simulated values of temperature at the outlet of the wheel on the supply stream. (b) Measured and simulated values of absolute humidity at the outlet of the wheel on the supply stream.
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
5. Analysis of the performance of a desiccant cooling system
simulated temperature values [°C]
a
40
The analysis investigates the performance of a system coupled to a building, on a similar configuration to the one used at the experimental set-up (Fig. 1) for steady state operation. The assumptions considered for the analysis are the following:
35 30 25
- Building with a conditioned space surface of 50 m2, presenting two levels of cooling load: - Moderate load: 2.8 kW sensible and 0.7 kW latent - Peak load: 4.2 kW sensible and 1.05 kW latent
humidifiers : OFF humidifier Exh : ON humidifiers : ON
20 15 10 10
15
20
25
30
35
40
simulated abs.humidity values [g/kg]
measured temperature values [°C]
b
1373
The calculation of the load of the building is based on the simple relations:
Qsens ¼ UAðt6 t1 Þ
(7a)
12
Qlat ¼ ð1 SHFÞQsens
(7b)
10
The value for the (UA) is set at 0.35 kW/K, while for the SHF is set at 0.75 according to [9,28], leading to the above mentioned values for the cooling load, which are considered realistic for the case studied. Even though Eq.(7a) and (7b) are a simplified approach of the cooling load, this approach is considered sufficient as regards the scope of the present study, which is not the analysis of the energy behavior of the building but that of the system.
8 6 humidifiers : OFF humidifier Exh : ON humidifiers : ON
4 2 0 0
2
4
6
8
10
12
measured abs. humidity values [g/kg] Fig. 4. (a) Measured and simulated values of temperature at the outlet of the desiccant system on the supply stream. (b) Measured and simulated values of absolute humidity at the outlet of the desiccant system on the supply stream.
marking is different, depending on the operation or not of the humidifiers, as this parameter determines whether the humidifier model would be included into the complete system model. Regarding temperature, an underestimation of the behaviour of the system is observed, which is higher in the case the exhaust humidifier is in operation, with values in the order of 2 C. Regarding absolute humidity, such an underestimation is observed in the case when no humidifier is in operation, while in the case where both humidifiers are in operation the model seems to overpredict the outlet values. The respective RMSE values are presented in Table 4. The reason for the observed deviation being higher than the deviation characterizing the submodels could be attributed to the inter-connection of the models with regard to the lay-out of the system, in other words the fact that the output data for the wheel constitute input data for the heat exchanger, etc. Regarding the effectiveness of the model, and given the fact that the proposed approach aims to present a simple and easy-toimplement tool for the analysis of the performance of the specific systems, results are satisfactory and it is concluded that the proposed model can effectively describe the operation of the complete system.
Table 4 RMSE values in terms of temperature and absolute humidity for the effectiveness of the proposed models for the complete system.
Humidifiers:OFF Humidifier exhaust:ON Humidifiers:ON All cases
N
RMSEt ( C)
RMSEw (g/kg)
55 52 11 118
0.72 1.87 0.61 1.35
0.51 0.31 0.48 0.43
- Moderate weather conditions, t1 ¼ 32 C – RH1 ¼ 30% (corresponding to the moderate cooling load) and peak weather conditions, t1 ¼ 36 C – RH1 ¼ 40% (corresponding to the peak cooling load), [33]. - The building is considered to be in the following thermal comfort conditions: t5 ¼ 24 C – RH5 ¼ 50% - The flow rate of processed air is in the range 600–1500 m3/h (0.167–0.417 m3/s), considered adequate for the specific airconditioned space. The flow rate of supply and exhaust air streams is the same. - The regeneration temperature is in the range 50–80 C, a realistic assumption for the connection of the system to a solar thermal energy source. - Two efficiency levels of the subsystems are considered: the first adopting the efficiency factors determined experimentally and the second the optimum values: - experimental : nF1 ¼ 0.15, nF2 ¼ 0.69, nhx ¼ 0.85, nhum ¼ 0.79 - optimum : nF1 ¼ 0.05, nF2 ¼ 0.95, nhx ¼ 0.95, nhum ¼ 0.95 Through the calculation procedure implemented, for the specific values of ambient and building conditions assumed, and the air flow rate and regeneration temperature selected, the degree of operation of the humidifiers is calculated. If the calculated values lie within the range (0,1), then the system can cope with the imposed cooling load. Otherwise, different values for the air flow rate and/or regeneration temperature have to be selected. For the assessment of the performance of the system, the COP of the system has been calculated using [7,34]:
Q ld Q reg
(8)
_ rðh6 h5 Þ Q ld ¼ m
(9a)
_ rCp ðt9 t8 Þ Q reg ¼ m
(9b)
COP ¼ where:
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
- The system can cover the loads imposed by the building for moderate weather conditions, with air flow rate values in the range 950–1500 m3/h. - For the same value of regeneration temperature it is feasible to implement different values of air flow rates. - The increase of air flow rate enables the decrease of regeneration temperature. - The increase of air flow rate, for the same value of regeneration temperature, has a negative impact on the COP of the system. - The increase of the regeneration temperature, for the same value of air flow rate, has a negative impact to the COP of the system. - The system cannot satisfy the imposed load by the building for all sets of regeneration temperature. This is evident for low temperatures, as in this case the system presents low capacity. On the other hand, for high temperatures the appearance of a saturation point for the regeneration temperature is observed. This phenomenon can be attributed to the fact that the capacity of the system, for the specific high temperatures, is considered overrated for the specific cooling load imposed by the building. - In peak weather conditions, a figure with the results is not presented, as the experimental set-up cannot cope with the imposed loads. More specifically, the simulation suggests air flow rate values greater than 1500 m3/h, which are considered non acceptable for the building studied. - The COP values of the experimental set-up are in agreement with values of other experimental installations, [34]. As regards the set-up with the optimum values for the efficiency factors of the subsystems: - The system satisfies the imposed load conditions, for moderate and peak weather conditions, as the presented values for air flow rate and regeneration temperature are within the acceptable limits.
experimental set-up - moderate weather conditions Treg : 50 degC 0.6
Treg : 60 degC
COP [ -]
Treg : 70 degC 0.4
0.2
0.0 800
1000
1200
1400
1600
air flow rate [m3/h] Fig. 5. Calculated COP of the desiccant system according to the system model, for the experimental set-up and moderate weather conditions.
a
optimum set-up - moderate weather conditions Treg : 50 degC
1.0
Treg : 60 degC Treg : 70 degC
COP [-]
The results for the set-up with the experimentally determined efficiency factors are presented in Fig 5. For the set-up with the optimum values for the efficiency factors and for moderate weather conditions the respective results are presented in Fig. 6a, while for peak weather conditions in Fig. 6b. From the results analysis, and focusing on the set-up with the efficiency values determined experimentally, the following conclusions can be drawn:
0.8
0.6
0.4 600
800
1000
air flow rate
1200
1400
1600
[m3/h]
optimum set-up - peak weather conditions
b
Treg : 70 degC
1.0
Treg : 80 degC
COP [-]
1374
0.8
0.6
0.4 1000
1200
1400
1600
air flow rate [m3/h] Fig. 6. (a) Calculated COP of the desiccant system according to the system model, for the optimum set-up and moderate weather conditions. (b) Calculated COP of the desiccant system according to the system model, for the optimum set-up and peak weather conditions.
- The values for air flow rate are lower, compared to the ones in the case of the set-up with the experimentally determined efficiency values. - The calculated COP for the optimum set-up is higher, compared to the COP of the set-up with the experimentally determined efficiency values, and agrees with values proposed in the relevant literature [7,8].
6. Conclusions In this work, a model for a desiccant air-conditioning system has been developed and experimentally validated on an actual installation. The model is rather simple and easy to implement and it is based on the description of the operation of the main subsystems of such an installation through specific efficiency factors. The analysis has proven the validity of the assumption for the proposed efficiency factors presenting constant values, which characterize the operation of each subsystem. Especially for the desiccant wheel case, the experimental determination of the combined potential efficiencies enables the elaboration of performance analysis on the basis of data describing the behavior of the specific desiccant wheel. The investigation of the performance of a typical desiccant airconditioning system, on the basis of the model, verifies the ability of these systems to satisfy actual cooling loads, on temperature levels that enable the use of flat plate solar collectors. For a given
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
configuration of a desiccant air-conditioning system, the air flow rate and regeneration temperature constitute decisive parameters for the performance of the system. For specific weather conditions, the system can satisfy the imposed cooling load with more than one pair of values for the discussed parameters. Analysis has indicated that there is a specific regeneration temperature limit beyond which the system cannot satisfy the imposed cooling load, as the system capacity exceeds the load requirements. On the level of system energy effectiveness, the increase of the air flow rate or of the regeneration temperature has a negative impact to the COP of the system. Thus, the operation of the system in the minimum possible regeneration temperature, enabling the exploitation of low temperature thermal energy has to be combined with the appropriate air flow rate value, in order to achieve the optimum COP level. Acknowledgments This research is supported by the Greek General Secretariat for Research and Technology (GSRT), within the framework of the Operational Programme for Competitiveness (EU – 3rd Community Support Programme, 2000–2006), Action : Renewable Energy Sources and Energy Saving. References [1] Sand JR, Fischer JC. Active desiccant integration with packaged rooftop HVAC equipment. Applied Thermal Engineering 2005;25:3138–48. [2] Pesaran AA, Peney TR, Czanderna AlW. Desiccant cooling: state-of-the-art assessment. NREL; 1992. [3] Mei VC, Chen FC, Lavan Z, Collier Jr RJ, Meckler G. An assessment of desiccant cooling and dehumidification technology. ORNL; 1992. [4] Daou K, Wang RZ, Xia ZZ. Desiccant cooling air-conditioning: a review. Renewable and Sustainable Energy Reviews 2006;10:55–77. [5] Waugaman DG, Kini A, Kettleborough CF. A review of desiccant cooling systems. Journal of Energy Resources Technology 1993;115:1–8. [6] Henning H-M. Solar assisted air conditioning of buildings – an overview. Applied Thermal Engineering 2007;27:1734–49. [7] Nelson JS, Beckmann WA, Mitchell JW, Close DJ. Simulations of the performance of open cycle desiccant air-conditioning systems using solar energy. Solar Energy 1978;21:273–8. [8] Jurinak JJ, Mitchell JW, Beckmann WA. Open-cycle desiccant air conditioning as an alternative to vapor compression cooling in residential applications. Transactions of the ASME 1984;106:252–60. [9] Davanagere BS, Sherif SA, Goswami DY. A feasibility study of a solar desiccant air-conditioning system – part I. International Journal of Energy Research 1999;23:7–21. [10] AGCC Application. Engineering manual for desiccant systems. Washington: American Gas Cooling Center; 1999. [11] Ge TS, Dai YJ, Wang RZ, Li Y. Experimental investigation on a one-rotor twostage rotary desiccant cooling system. Energy 2008;33:1807–15.
1375
[12] Jain S, Dhar PL. Evaluation of solid-desiccant-based evaporative cooling cycles for typical hot and humid climates. International Journal of Refrigeration 1995;18:287–96. [13] Joudi KA, Madhi SM. An experimental investigation into a solar-assisted desiccant-evaporative air-conditioning system. Solar Energy 1987;39(2): 97–107. [14] Subramanyam N, Maiya MP, Murthy SS. Parametric studies on a desiccant assisted air-conditioner. Applied Thermal Engineering 2004;24:2679–88. [15] Henning H, Erpenbeck T, Hindenburg C, Santamaria IS. The potential of solar energy use in desiccant cooling cycles. International Journal of Refrigeration 2001;24:220–9. [16] Ahmed MH, Kattab NM, Fouad M. Evaluation and optimization of solar desiccant wheel performance. Renewable Energy 2005;30:305–25. [17] Krishna SM, Murthy SS. Experiments on silica gel rotary dehumidifier. Heat Recovery Systems & CHP 1989;9(5):467–73. [18] Jia CX, Dai YJ, Wu JY, Wang RZ. Use of compound desiccant to develop high performance desiccant cooling system. International Journal of Refrigeration 2007;30:345–53. [19] Zhang XJ, Sumathy K, Dai YJ, Wang RZ. Dynamic hygroscopic effect of the composite material used in desiccant rotary wheel. Solar Energy 2006;80:1058–61. [20] Ge TS, Li Y, Wang RZ, Dai YJ. A review of the mathematical models for predicting rotary desiccant wheel. Renewable and Sustainable Energy Reviews 2008;12:1485–528. [21] Joudi KA, Dhaidan NS. Application of solar assisted heating and desiccant cooling systems for a domestic building. Energy Conversion and Management 2001;42:995–1022. [22] Mavroudaki P, Beggs CB, Sleigh PA, Halliday SP. The potential for solar powered single-stage desiccant cooling in southern Europe. Applied Thermal Engineering 2002;22:1129–40. [23] Halliday SP, Beggs CB, Sleigh PA. The use of solar desiccant cooling in the UK: a feasibility study. Applied Thermal Engineering 2002;22:1327–38. [24] Davanagere BS, Sherif SA, Goswami DY. A feasibility study of a solar desiccant air-conditioning system – part II. International Journal of Energy Research 1999;23:103–16. [25] Kays WM, London AL. Compact heat exchangers. 3rd ed. London: Mc-Graw Hill; 1984. [26] Kuehn TH, Ramsey JW, Threlkeld JL. Thermal environmental engineering. Prentice Hall; 2001. [27] Camargo JR, Ebinuma CD, Silveira JL. Experimental performance of a direct evaporative cooler operating during summer in a Brazilian city. International Journal of Refrigeration 2005;28:1124–32. [28] ASHRAE. Handbook: 1997 fundamentals, Inch-Pound editions. Atlanta: ASHRAE; 1997. [29] Maclaine-Cross IL, Banks PJ. Coupled heat and mass transfer in regenerators – prediction using an analogy with heat transfer. International Journal of Heat and Mass Transfer 1972;15:1225–42. [30] Jurinak J.J., Open cycle solid desiccant cooling – component models and system Ssmulations, Ph.D. Thesis, Univ. of Wisconsin-Madison; (1982) [31] Banks PJ. Prediction of Heat and Mass Regenerator performance using nonlinear analogy method: part 1 – basis. Transactions of the ASME. Journal of Heat Transfer 1985;107:222–9. [32] Press W, Teukolsky SA, Vetterling WT, Flannery BP. In: Numerical recipes. 2nd ed. Oxford: Cambridge University Press; 1996. [33] Balaras CA, Kontogiannidis S, Lykoudis S. Typical meteorological year for athens. National Observatory of Athens 2006. [34] Balaras CA, Grossman G, Henning H-M, Infante Ferreira CA, Podesser E, Wang L, et al. Solar air-conditioning in Europe – an overview. Renewable and Sustainable Energy Reviews 2007;11:299–314.
Contents lists available at ScienceDirect
Renewable Energy journal homepage: www.elsevier.com/locate/renene
Theoretical and experimental investigation of the performance of a desiccant air-conditioning system G. Panaras a, *, E. Mathioulakis a, V. Belessiotis a, N. Kyriakis b a b
Solar & Other Energy Systems Laboratory, NCSR ‘‘DEMOKRITOS’’, 15310, Agia Paraskevi Attikis, Greece Mechanical Engineering Department, Process Equipment Design Laboratory, Aristotle University of Thessaloniki, Greece
a r t i c l e i n f o
a b s t r a c t
Article history: Received 25 May 2009 Accepted 20 September 2009 Available online 25 November 2009
Solid desiccant air-conditioning systems present an interesting alternative solution with regard to the conventional vapour compression systems, in the sense that they do not use any refrigerants and they present the opportunity to exploit thermal energy, and more specifically solar thermal energy, instead of electrical energy. In the present work a theoretical model is presented for the operation of a desiccant air-conditioning system, developed on the basis of existing approaches for the modelling of the main subsystems of such a device. The model is experimentally validated on a real scale system, through the exploitation of a significant number of measurements, which correspond to a typical range of operation conditions for these systems. The proposed model is used for the investigation of the performance of a system with a typical set-up, examining the influence of parameters such as the weather conditions, the level of the imposed cooling load, the efficiency level of the main subsystems of the set-up, the air flow rate and the regeneration temperature. The results confirm the potential of the examined technology to satisfy actual cooling loads, and at the same time they lead to specific conclusions for the operation of these systems. Ó 2009 Elsevier Ltd. All rights reserved.
Keywords: Solar air-conditioning Desiccant
1. Introduction Solid desiccant air-conditioning systems can be an interesting alternative approach with regard to the vapour compression systems, considering the environmental friendly operation, as well as the performance level of the proposed systems. In principle air-conditioning concerns not only the regulation of temperature, but the humidity as well. The conventional air-conditioning systems do not usually cope with the issue of humidity requirements, and in cases they do, they attempt to control the humidity through the deep cooling of the processed air, in temperatures below its dew point. The moisture condenses and the dehumidified air, which satisfies the humidity requirements of the conditioned space, is heated to the desired temperature, in order to satisfy the temperature requirements of the conditioned space as well. In principle, the control of humidity through temperature, described above, can be characterized as an energy consuming procedure and in some cases it cannot ensure the achievement of the temperature and humidity levels required by the user [1]. On the contrary, desiccant air-conditioning systems achieve the dehumidification of air on a direct way, through the use of the
* Corresponding author. Tel.: þ30 210 650 3810; fax: þ30 210 654 4592. E-mail address: [email protected] (G. Panaras). 0960-1481/$ – see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.renene.2009.11.011
desiccant material, providing processed air of high cooling potential due to its low humidity. The processed air can be evaporatively cooled, enabling the simultaneous treatment of sensible and latent loads [2,3], even if limitations as regards the effectiveness of this treatment might occur, especially in fairly hot and wet climates [4]. Desiccant air-conditioning systems do not use harmful to the environment refrigerants, neither do they consume electrical energy for the cooling process, but thermal energy, in temperature levels that favour the use of solar energy, and more specifically flat plate solar collectors. The latter can be considered as one of the main advantages of desiccant systems, pointing out that cooling loads and solar radiation follow similar time patterns. Thus the majority of research approaches refer to systems connected to a solar thermal energy source. The technology of solid desiccant air-conditioning systems presents positive prospects as regards dissemination in the market. In Europe, the respective bibliography refers to sixteen installed research or demonstration plants, which consider systems connected to solar thermal energy source [5]. Respectively, in the USA a significant commercial activity is reported, mainly for systems being supplied by conventional heating sources, installed as early as the beginning of 80’s. Nevertheless, this activity declined after 2000, mainly due to economic reasons [6]. Proceeding to a general assessment of the technological maturity and the emerging research issues of the desiccant air-conditioning
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
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Nomenclature
(UA)
a COP Cp C F1, F2 h SHF _ m N
Greek letters r density [kg/m3]
h hF1, hF2 NTU s t w Qb Qld Qreg
humidifier degree of operation [-] system coefficient of performance [-] air specific heat capacity [kJ/kgK] heat capacity rate [kW/K] combined potentials [-] enthalpy [kJ/kg] sensible heat factor [-] air flow rate [m3/s] total number of measurements [-] efficiency [-] efficiency of the desiccant wheel with regard to combined potentials F1, F2 [-] number of transfer units standard deviation [-] air temperature [ C] air absolute humidity rate [g /kg] cooling load of the building [kW] cooling power provided to the space by the system [kW] thermal power for the regeneration of the wheel [kW]
systems, as expressed in the relevant literature, the performance level of the involved subsystems is considered satisfactory, even though research efforts continue on this particular issue. The basic research aspect for this technology is the standardization of the proposed systems, including amongst others the solution of problems related to the successful integration of the subsystems to the total installation and the development of packaged systems, as well as the development of design methodologies [5–7]. The connection to a solar thermal energy heating source increases the complexity of the system, thus it stresses the importance of the discussed aspect of standardization. The development of a simple and reliable model for the analysis of the performance of the systems can lead to useful conclusions as regards the operation characteristics of the systems, and the effect of the various parameters. This model can be used as the kernel of a design methodology, since it allows the optimization of the system design, with regard to the specific requirements of a given application. Such a model based on existing approaches for the modelling of the main subsystems, is proposed in the present work. The operation of the subsystems is described by specific efficiency factors, and the validity of the assumption for these factors presenting constant values is experimentally investigated on an actual installation, for a range of typical conditions for the air-conditioning applications.
overall heat transfer coefficient times area [kW/K]
Subscripts cal simulated values (of temperature or absolute humidity) des desiccant wheel e exhaust stream hx heat exchanger hum humidifier i ¼ 1,2,3 index referring to various positions of the desiccant system j ¼ 1.N index referring to the number of measurement performed lat latent (load) reg regeneration (temperature) s supply stream sat saturated air sens sensible (load) wb wet bulb (temperature)
cooler, where it is cooled and humidified to state (6). This air can cool the hot and dry air, processed by the wheel, through the air-toair heat exchanger and exits the heat exchanger at state (7). It is then heated to the regeneration temperature, at state (8). It then passes through the wheel, where it is cooled and humidified to state (9), and finally is exhausted to the environment. It should be noted, that the cycle just described is referred in bibliography as the ventilation cycle, as it enables 100% renewal of the air of the conditioned space. In the relevant bibliography the recirculation cycle is also referred to. Through the operation of the recirculation cycle no renewal of the air of the conditioned space takes place. In practice, as regards actual installations, a combination of the two cycles is proposed, thus enabling the renewal of the air on a specific percentage, satisfying the hygienic requirements of the conditioned space [10]. Variations of the above described basic desiccant air-conditioning system do exist, mainly concerning the use of the evaporative cooler for the cooling of the supply air, the potential use of a conventional cooling coil aiming at the increase of the system sensible load capacity, and the selection of the regenerative heating source [10]. Other solutions include the staged dehumidification [11], the use of two desiccant wheels [12], or the staged regeneration of the wheel [2]. In any case, the analysis of the alternative configurations is rather extensive and exceeds the scope of this study which would concentrate on the analysis of the performance of a typical system, as the one presented in Fig. 1.
2. Desiccant technology in air-conditioning applications A complete desiccant air-conditioning system consists of a desiccant wheel (DW), two evaporative coolers (EC) and an air-toair heat exchanger (HX). A typical set-up is presented in Fig. 1 [7–9]. According to the systems’ operation principle, depicted in the psychrometric chart of Fig. 2, the following processes can be identified: on the supply stream, processed ambient air of state (1) is dehumidified and heated through the desiccant wheel. Air of state (2), enters the heat exchanger, where it is cooled (state 3). It then passes through the evaporative cooler, where it is cooled and humidified, entering the building at state (4). On the exhaust stream, air from the building, on the obtained thermal comfort conditions of state (5), enters the exhaust stream evaporative
DW
EC 4
1
2
3
Supply air Building Exhaust air 5
6 CE
8
7 HX
Environment
9
Heat source
Fig. 1. Typical desiccant air-conditioning system.
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G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
0.035
hhx ¼
0.03
w [g/kg]
0.025 0.02
9
0.015
6
7 5
0.01
8
1
4 0.005 3 0
0
20
2 40
60
80
t [oC] Fig. 2. Presentation of the desiccant air-conditioning cycle on the psychrometric chart.
In the relevant literature, experimental approaches appear, studying the effect of various parameters on the performance of a solar desiccant air-conditioning system [13,14], even though the studies referring to actual scale systems are rather limited [15]. Certain researchers concentrate on the study of the behaviour of the desiccant wheel [16,17] or on the investigation for innovative concepts/materials, as regards the wheel [18,19]. In the field of theoretical studies, the issue of the desiccant wheel modelling has been studied extensively [20]. As regards the system as a whole, the various approaches mainly investigate the effect of various parameters to the performance [7,8,21], and assess the potential of the examined technology to cope with the imposed cooling loads for various types of buildings and for different geographical areas as well [9,15,22–24].
According to the NTU approach, the efficiency is a function of the NTU. It should be noted, however, that within the NTU range of values, there is a considerable range for which the efficiency remains practically constant. Given the dependence of NTU on (UA) and the fact that in the relevant heat transfer analysis theory, the (UA) is assumed to be constant with regard to operation conditions, including flow rate and rotational speed (for the case of rotary heat exchangers), it can be assumed that efficiency remains constant over a specific range of operation conditions [25]. The present work experimentally investigates the validity of this assumption, for a range of operation conditions relevant to the conditions of the experiment, and would proceed to the determination of the proposed value of the efficiency. It should be noted that most of the theoretical studies on the performance of desiccant air-conditioning systems reported (e.g. [7,9]), follow the approach of constant efficiency in modelling the air-to-air heat exchanger of the system. 3.3. Evaporative cooler The most commonly used type of evaporative cooler in desiccant air-conditioning systems is the evaporative pad humidifier [2,3]. During the operation of the humidifier, the water in the surface of the pad is evaporated and is absorbed by the air stream. If there are no thermal losses to the environment, the process can be considered to be adiabatic and according to the relevant theory the wet bulb temperature of air remains steady [26]. In this case a useful quantity for expressing the performance of the humidifier is the efficiency in terms of temperature [27,28], (the indices refer to the case of humidifier installed at the exhaust stream, Fig. 1):
hhum;t ¼
hhum;w ¼
3.1. Complete system model As already described, a desiccant air-conditioning system consists of a number of subsystems. The theoretical analysis which is presented in this section aims at the selection of specific models for the subsystem operation. According to the set-up of a desiccant system, the input of the one subsystem is the output of the other. Thus the proposed subsystem models have to be appropriately integrated to the model describing the operation of the complete system. Such an approach has been adopted at present work, for a system referring to a typical set-up, as the one presented in Fig. 1. 3.2. Air-to-air heat exchanger For the description of the operation of the air-to-air heat exchanger, the efficiency-NTU approach is adopted, where NTU refers to the Number of Transfer Units, given by [25]:
UA Cmin
(2)
t5 t6 t5 twb
(3a)
Also the efficiency in terms of humidity can be expressed as [26]:
3. Theoretical approach of subsystems performance – the proposed system model
NTU ¼
t2 t3 t t7 ¼ 6 t2 t6 t6 t2
(1)
In the case of balanced flow heat exchanger, the efficiency can be calculated by (the indices refer to the positions of subsystems in a typical system as the one presented in Fig. 1) [25]:
w5 w6 w5 wsat
(3b)
For the case of the adiabatic process, both quantities of nhum,t and nhum,w are equal, as it can easily be concluded through the application of basic psychrometric analysis. According to Camargo et al. [27], an efficiency-NTU relation can be expressed for the nhum,t. This relation is of the exponential type, and it is analogous to the relation for the heat exchanger case. Thus, the assumption of constant value for the efficiency can be also valid in the humidifier case. The experimental investigation would study the validity of this assumption, for the range of operation conditions relevant to the conditions of the experiment, and the humidifier efficiency will be determined. As in the case of heat exchangers, most of the theoretical studies on the performance of desiccant air-conditioning systems use the constant efficiency model for the performance of the humidifier [7,9]. 3.6. Desiccant wheel For the desiccant wheel, the approach of Maclaine-Cross and Banks has been selected [29], which is mentioned in the bibliography as analogy theory. This approach models the dehumidification process, a combined heat and mass process, in analogy with a simple heat transfer process. The analogy is expressed through
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
the combined potentials F1, F2. The combined potentials have resulted through the analysis of heat and mass transfer equations in an elementary segment of the desiccant wheel and their use significantly simplifies the relevant equations, establishing the respective analogy. The discussion for the nature and the role of combined potentials is rather extensive and complicated. For the scope of the current analysis, it can be stated that constant F1 lines coincide with constant enthalpy lines, while constant F2 lines coincide with constant relative humidity lines on the psychrometric chart [30]. The combined potential functions are dependent on the temperature and relative humidity of the processed air, as well as on the thermophysical properties of the processed air and of the wheel context, including the desiccant material [31]. According to the analogy theory, the efficiency indices of the wheel hF1, hF2 with regard to the combined potentials F1, F2 can be calculated in analogy to the energy efficiency of a respective heat exchanger. Given the values of the efficiency indices of the wheel hF1, hF2, the temperature and humidity of the processed air at the outlet of the wheel could be calculated, provided a relation explicitly describing the F1, F2 dependence on temperature and humidity is available. Jurinak has expressed such a relation for the working pair of air-silica gel , leading to the formulation of the model presented through Eqs.(4) and (5) (i ¼ 1,2,8 according to the system presented in Fig. 1) [30]:
F1;i ¼
2865 ðti þ 273:15Þ
1:49
w 0:8624 i 1000
þ 4:344
(4a)
F2;i ¼
w 0:07969 ðti þ 273:15Þ1:49 i 1:127 6360 1000
(4b)
hF1 ¼
F1;2 F1;1 F1;9 F1;1
(5a)
F2;2 F2;1 ¼ F2;9 F2;1
(5b)
hF2
It is useful to note that the efficiency factor hF1 expresses the degree the process approximates the adiabatic one, while the hF2 expresses the degree of dehumidification, [30]. The pair [hF1, hF2] ¼ [0,1] corresponds to an ideal process, which is adiabatic, with the maximum dehumidification level for the respective geometry and flow conditions. The present work experimentally investigates the assumption of constant efficiency factors, in the range of operation conditions relevant to the conditions of the experiment and calculates these values for the wheel used in the specific installation. 4. Experimental investigation
1371
30 cm. The air-to-air heat exchanger used is of the counterflow plate type. The selection of these subsystems was based on their performance level, availability and cost. Especially for the heat exchanger, it should be noted that the air flow rate of the set-up is rather low, thus enabling the use of a plate device, instead of a rotary one which is usually used in desiccant systems. For the detailed and reliable recording of the experimental data, a properly designed measuring set-up was developed. The temperature was measured in all the positions 1–9 in Fig. 1, while absolute humidity has been measured in positions 1, 4, 5, 6, 9. It should be noted that the temperature and humidity field at the outlet of the wheel on the supply stream is not homogeneous. The same is observed at other positions of the system as well, as the non-homogenous temperature distribution is transferred through the operation of the heat exchanger. Thus, and in order to increase the accuracy of temperature measurement, four sensors have been installed at the outlet of the wheel, as well as at other specific positions of the system, the value of temperature used in calculation being the average of the four sensors. Regarding the humidity measurements, and due to the configuration of the sensors used, the installation of more than one is difficult. It was decided therefore to measure the humidity in positions where the field is homogenous. For the study of the performance of the wheel, this position is at the outlet of the supply humidifier (i ¼ 4, Fig. 1). It should be noted that during the study of the behaviour of the wheel, this humidifier was not in operation. According to the technical specifications of the sensors used, the accuracy of measurements is estimated to be 1 C for the temperature, 0.2 g/kg for absolute humidity rate and 50 m3/h for the air flow rate. The presented values are considered acceptable, with regard to the scope of this work. 4.2. Measurement conditions The parameters which can be regulated at the experimental setup, defining the operating conditions of the system, are the following: - the supply and the exhaust air flow rate. In this study, only the balanced flow system is investigated - the desiccant wheel rotational speed. In this application, the speed proposed by the manufacturer (6 rph) was used. - the regeneration temperature of the desiccant wheel - the operation or not of the exhaust and supply stream humidifier Table 1 presents the range of measurement conditions applied at the experimental investigation. The operation of the system was monitored on a daily basis, under steady state conditions, according to different scenarios.
4.1. Experimental set-up
4.3. Experimental validation of the subsystem models
The experimental set-up refers to a typical desiccant airconditioning system, as the one presented in Fig. 1. The system has been developed in the premises of Solar & Other Energy Systems Laboratory of NCSR DEMOKRITOS. The system is designed for the 450–1800 m3/h air flow rate range and it is connected to a small building, acting as a test-cell, of conditioned space surface of 48 m2 and volume of 150 m3. The desiccant wheel is 630 mm in diameter, thickness of 200 mm with silica gel. The evaporative cooler is of the evaporative pad type, with a useful surface of 60 60 cm and pad thickness of
The experimental validation investigates whether the operation of the main subsystems can be effectively described by the
Table 1 Measurement conditions for the experimental study.
MIN MAX Flow rate (m3/h)
t1( C)
w1(g/kg)
t8( C)
w8(g/kg)
24 40
4 14 600
50 80 1000
3 15 1200
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
RMSEt ¼
RMSEw ¼
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uP 2 u N t j ¼ 1 tj2 tj2;cal
(6a)
N vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi uP 2 u N t j ¼ 1 wj2 wj2;cal
(6b)
N
For the experimental analysis of this work, the RMSE values in terms of temperature and absolute humidity are presented in Table 3 for all subsystems. For the desiccant wheel case, the measured and simulated values for temperature and absolute humidity are presented in Fig. 3a and b respectively. As regards humidity, the agreement is very good, while in the temperature case a deviation of the order of 2 C between the compared values is observed. The same conclusions can be drawn through the examination of the RMSE values of Table 3. Considering the uncertainties of the measurements implemented at an actual system, as well as the fact that in the respective air-conditioning applications the humidity factor is of higher importance, it can be stated that the above results demonstrate the effectiveness of the proposed approach for the wheel. Regarding the humidifier and the heat exchanger, it can be concluded from the respective RMSE values (Table 3) that the agreement of the measured and simulated values for both systems is very good, thus proving the effectiveness of the proposed approach. It should be noted that in the humidifier case, there is a significant difference between the determined values of efficiency in terms of temperature and humidity. This can be attributed to the presence of thermal losses of the actual evaporative cooling process, imposing deviation from the adiabatic one. The existence Table 2 Efficiency factors for the subsystems as determined through the experimental analysis.
Heat exchanger Humidifier
Desiccant wheel
N
hhx/hhum,t
s (hhx/hhum,t)
hhum,w
s (hhum,w)
141 79
0.85 0.65
0.02 0.03
– 0.79
– 0.05
N
hF1
s (hF1)
107
0.15
0.04
hF2 0.69
s (hF2) 0.06
Table 3 RMSE values in terms of temperature and absolute humidity for the effectiveness of the examined models for the subsystems.
Heat exchanger Humidifier Desiccant wheel
RMSEt ( C)
RMSEw (g/kg)
0.34 0.31 1.24
– 0.26 0.30
of thermal losses is connected to the metallic surfaces of the humidifier, allowing heat to be transferred to the environment, as well as to the sensible heat transfer through the basin for the recycling of the supplied water. Given that the specific phenomena can be eliminated through proper actions for the needs of the performance analysis which follows on, a common efficiency index for both temperature and humidity (nhum) can be assumed, presenting the value determined for nhum,w. 4.4. Experimental validation of the complete system model The validation of the proposed model for the complete system has been performed through the comparison of experimental and simulated values of temperature and absolute humidity of the air supplied to the building by the system (t5, w5) during steady state operation periods. The experimental and simulated values for temperature and absolute humidity are presented in Fig. 4a and b respectively. The
a
simulated temperature values [°C]
proposed models with constant values of the selected efficiency factors, not depended on the operating conditions. For the validation of the above assumption, and for each subsystem, the average value of the selected efficiency factors is calculated, for the total number of available measurements. This value together with the standard deviation is presented in Table 2 for all subsystems. The values of standard deviation are rather low, suggesting the validity of the discussed assumption. According to the examined model and the efficiency factors determined, the temperature, in the case of all subsystems, and the absolute humidity, in the case of the humidifier and the desiccant wheel, at the outlet of the examined subsystem can be calculated, under the operating conditions of each measurement. The so calculated values would be referred hereafter as ‘‘simulated values’’. Thus, N pairs of measured and simulated values of temperature and/or absolute humidity are available, where N denotes the number of measurements for each subsystem. The validation of the examined models is based on the comparison between the measured and simulated values. The Root Mean Standard Error (RMSE) is used as a quantitative index for the comparison of experimental and simulated values, [32]. RMSE is calculated by the following equations:
60 55 50 45 40 35 30 30
35
40
45
50
55
60
10
12
measured temperature values [°C]
b simulated abs.humidity values [g/kg]
1372
12 10 8 6 4 2 0 0
2
4
6
8
measured abs. humidity values [g/kg] Fig. 3. (a) Measured and simulated values of temperature at the outlet of the wheel on the supply stream. (b) Measured and simulated values of absolute humidity at the outlet of the wheel on the supply stream.
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
5. Analysis of the performance of a desiccant cooling system
simulated temperature values [°C]
a
40
The analysis investigates the performance of a system coupled to a building, on a similar configuration to the one used at the experimental set-up (Fig. 1) for steady state operation. The assumptions considered for the analysis are the following:
35 30 25
- Building with a conditioned space surface of 50 m2, presenting two levels of cooling load: - Moderate load: 2.8 kW sensible and 0.7 kW latent - Peak load: 4.2 kW sensible and 1.05 kW latent
humidifiers : OFF humidifier Exh : ON humidifiers : ON
20 15 10 10
15
20
25
30
35
40
simulated abs.humidity values [g/kg]
measured temperature values [°C]
b
1373
The calculation of the load of the building is based on the simple relations:
Qsens ¼ UAðt6 t1 Þ
(7a)
12
Qlat ¼ ð1 SHFÞQsens
(7b)
10
The value for the (UA) is set at 0.35 kW/K, while for the SHF is set at 0.75 according to [9,28], leading to the above mentioned values for the cooling load, which are considered realistic for the case studied. Even though Eq.(7a) and (7b) are a simplified approach of the cooling load, this approach is considered sufficient as regards the scope of the present study, which is not the analysis of the energy behavior of the building but that of the system.
8 6 humidifiers : OFF humidifier Exh : ON humidifiers : ON
4 2 0 0
2
4
6
8
10
12
measured abs. humidity values [g/kg] Fig. 4. (a) Measured and simulated values of temperature at the outlet of the desiccant system on the supply stream. (b) Measured and simulated values of absolute humidity at the outlet of the desiccant system on the supply stream.
marking is different, depending on the operation or not of the humidifiers, as this parameter determines whether the humidifier model would be included into the complete system model. Regarding temperature, an underestimation of the behaviour of the system is observed, which is higher in the case the exhaust humidifier is in operation, with values in the order of 2 C. Regarding absolute humidity, such an underestimation is observed in the case when no humidifier is in operation, while in the case where both humidifiers are in operation the model seems to overpredict the outlet values. The respective RMSE values are presented in Table 4. The reason for the observed deviation being higher than the deviation characterizing the submodels could be attributed to the inter-connection of the models with regard to the lay-out of the system, in other words the fact that the output data for the wheel constitute input data for the heat exchanger, etc. Regarding the effectiveness of the model, and given the fact that the proposed approach aims to present a simple and easy-toimplement tool for the analysis of the performance of the specific systems, results are satisfactory and it is concluded that the proposed model can effectively describe the operation of the complete system.
Table 4 RMSE values in terms of temperature and absolute humidity for the effectiveness of the proposed models for the complete system.
Humidifiers:OFF Humidifier exhaust:ON Humidifiers:ON All cases
N
RMSEt ( C)
RMSEw (g/kg)
55 52 11 118
0.72 1.87 0.61 1.35
0.51 0.31 0.48 0.43
- Moderate weather conditions, t1 ¼ 32 C – RH1 ¼ 30% (corresponding to the moderate cooling load) and peak weather conditions, t1 ¼ 36 C – RH1 ¼ 40% (corresponding to the peak cooling load), [33]. - The building is considered to be in the following thermal comfort conditions: t5 ¼ 24 C – RH5 ¼ 50% - The flow rate of processed air is in the range 600–1500 m3/h (0.167–0.417 m3/s), considered adequate for the specific airconditioned space. The flow rate of supply and exhaust air streams is the same. - The regeneration temperature is in the range 50–80 C, a realistic assumption for the connection of the system to a solar thermal energy source. - Two efficiency levels of the subsystems are considered: the first adopting the efficiency factors determined experimentally and the second the optimum values: - experimental : nF1 ¼ 0.15, nF2 ¼ 0.69, nhx ¼ 0.85, nhum ¼ 0.79 - optimum : nF1 ¼ 0.05, nF2 ¼ 0.95, nhx ¼ 0.95, nhum ¼ 0.95 Through the calculation procedure implemented, for the specific values of ambient and building conditions assumed, and the air flow rate and regeneration temperature selected, the degree of operation of the humidifiers is calculated. If the calculated values lie within the range (0,1), then the system can cope with the imposed cooling load. Otherwise, different values for the air flow rate and/or regeneration temperature have to be selected. For the assessment of the performance of the system, the COP of the system has been calculated using [7,34]:
Q ld Q reg
(8)
_ rðh6 h5 Þ Q ld ¼ m
(9a)
_ rCp ðt9 t8 Þ Q reg ¼ m
(9b)
COP ¼ where:
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
- The system can cover the loads imposed by the building for moderate weather conditions, with air flow rate values in the range 950–1500 m3/h. - For the same value of regeneration temperature it is feasible to implement different values of air flow rates. - The increase of air flow rate enables the decrease of regeneration temperature. - The increase of air flow rate, for the same value of regeneration temperature, has a negative impact on the COP of the system. - The increase of the regeneration temperature, for the same value of air flow rate, has a negative impact to the COP of the system. - The system cannot satisfy the imposed load by the building for all sets of regeneration temperature. This is evident for low temperatures, as in this case the system presents low capacity. On the other hand, for high temperatures the appearance of a saturation point for the regeneration temperature is observed. This phenomenon can be attributed to the fact that the capacity of the system, for the specific high temperatures, is considered overrated for the specific cooling load imposed by the building. - In peak weather conditions, a figure with the results is not presented, as the experimental set-up cannot cope with the imposed loads. More specifically, the simulation suggests air flow rate values greater than 1500 m3/h, which are considered non acceptable for the building studied. - The COP values of the experimental set-up are in agreement with values of other experimental installations, [34]. As regards the set-up with the optimum values for the efficiency factors of the subsystems: - The system satisfies the imposed load conditions, for moderate and peak weather conditions, as the presented values for air flow rate and regeneration temperature are within the acceptable limits.
experimental set-up - moderate weather conditions Treg : 50 degC 0.6
Treg : 60 degC
COP [ -]
Treg : 70 degC 0.4
0.2
0.0 800
1000
1200
1400
1600
air flow rate [m3/h] Fig. 5. Calculated COP of the desiccant system according to the system model, for the experimental set-up and moderate weather conditions.
a
optimum set-up - moderate weather conditions Treg : 50 degC
1.0
Treg : 60 degC Treg : 70 degC
COP [-]
The results for the set-up with the experimentally determined efficiency factors are presented in Fig 5. For the set-up with the optimum values for the efficiency factors and for moderate weather conditions the respective results are presented in Fig. 6a, while for peak weather conditions in Fig. 6b. From the results analysis, and focusing on the set-up with the efficiency values determined experimentally, the following conclusions can be drawn:
0.8
0.6
0.4 600
800
1000
air flow rate
1200
1400
1600
[m3/h]
optimum set-up - peak weather conditions
b
Treg : 70 degC
1.0
Treg : 80 degC
COP [-]
1374
0.8
0.6
0.4 1000
1200
1400
1600
air flow rate [m3/h] Fig. 6. (a) Calculated COP of the desiccant system according to the system model, for the optimum set-up and moderate weather conditions. (b) Calculated COP of the desiccant system according to the system model, for the optimum set-up and peak weather conditions.
- The values for air flow rate are lower, compared to the ones in the case of the set-up with the experimentally determined efficiency values. - The calculated COP for the optimum set-up is higher, compared to the COP of the set-up with the experimentally determined efficiency values, and agrees with values proposed in the relevant literature [7,8].
6. Conclusions In this work, a model for a desiccant air-conditioning system has been developed and experimentally validated on an actual installation. The model is rather simple and easy to implement and it is based on the description of the operation of the main subsystems of such an installation through specific efficiency factors. The analysis has proven the validity of the assumption for the proposed efficiency factors presenting constant values, which characterize the operation of each subsystem. Especially for the desiccant wheel case, the experimental determination of the combined potential efficiencies enables the elaboration of performance analysis on the basis of data describing the behavior of the specific desiccant wheel. The investigation of the performance of a typical desiccant airconditioning system, on the basis of the model, verifies the ability of these systems to satisfy actual cooling loads, on temperature levels that enable the use of flat plate solar collectors. For a given
G. Panaras et al. / Renewable Energy 35 (2010) 1368–1375
configuration of a desiccant air-conditioning system, the air flow rate and regeneration temperature constitute decisive parameters for the performance of the system. For specific weather conditions, the system can satisfy the imposed cooling load with more than one pair of values for the discussed parameters. Analysis has indicated that there is a specific regeneration temperature limit beyond which the system cannot satisfy the imposed cooling load, as the system capacity exceeds the load requirements. On the level of system energy effectiveness, the increase of the air flow rate or of the regeneration temperature has a negative impact to the COP of the system. Thus, the operation of the system in the minimum possible regeneration temperature, enabling the exploitation of low temperature thermal energy has to be combined with the appropriate air flow rate value, in order to achieve the optimum COP level. Acknowledgments This research is supported by the Greek General Secretariat for Research and Technology (GSRT), within the framework of the Operational Programme for Competitiveness (EU – 3rd Community Support Programme, 2000–2006), Action : Renewable Energy Sources and Energy Saving. References [1] Sand JR, Fischer JC. Active desiccant integration with packaged rooftop HVAC equipment. Applied Thermal Engineering 2005;25:3138–48. [2] Pesaran AA, Peney TR, Czanderna AlW. Desiccant cooling: state-of-the-art assessment. NREL; 1992. [3] Mei VC, Chen FC, Lavan Z, Collier Jr RJ, Meckler G. An assessment of desiccant cooling and dehumidification technology. ORNL; 1992. [4] Daou K, Wang RZ, Xia ZZ. Desiccant cooling air-conditioning: a review. Renewable and Sustainable Energy Reviews 2006;10:55–77. [5] Waugaman DG, Kini A, Kettleborough CF. A review of desiccant cooling systems. Journal of Energy Resources Technology 1993;115:1–8. [6] Henning H-M. Solar assisted air conditioning of buildings – an overview. Applied Thermal Engineering 2007;27:1734–49. [7] Nelson JS, Beckmann WA, Mitchell JW, Close DJ. Simulations of the performance of open cycle desiccant air-conditioning systems using solar energy. Solar Energy 1978;21:273–8. [8] Jurinak JJ, Mitchell JW, Beckmann WA. Open-cycle desiccant air conditioning as an alternative to vapor compression cooling in residential applications. Transactions of the ASME 1984;106:252–60. [9] Davanagere BS, Sherif SA, Goswami DY. A feasibility study of a solar desiccant air-conditioning system – part I. International Journal of Energy Research 1999;23:7–21. [10] AGCC Application. Engineering manual for desiccant systems. Washington: American Gas Cooling Center; 1999. [11] Ge TS, Dai YJ, Wang RZ, Li Y. Experimental investigation on a one-rotor twostage rotary desiccant cooling system. Energy 2008;33:1807–15.
1375
[12] Jain S, Dhar PL. Evaluation of solid-desiccant-based evaporative cooling cycles for typical hot and humid climates. International Journal of Refrigeration 1995;18:287–96. [13] Joudi KA, Madhi SM. An experimental investigation into a solar-assisted desiccant-evaporative air-conditioning system. Solar Energy 1987;39(2): 97–107. [14] Subramanyam N, Maiya MP, Murthy SS. Parametric studies on a desiccant assisted air-conditioner. Applied Thermal Engineering 2004;24:2679–88. [15] Henning H, Erpenbeck T, Hindenburg C, Santamaria IS. The potential of solar energy use in desiccant cooling cycles. International Journal of Refrigeration 2001;24:220–9. [16] Ahmed MH, Kattab NM, Fouad M. Evaluation and optimization of solar desiccant wheel performance. Renewable Energy 2005;30:305–25. [17] Krishna SM, Murthy SS. Experiments on silica gel rotary dehumidifier. Heat Recovery Systems & CHP 1989;9(5):467–73. [18] Jia CX, Dai YJ, Wu JY, Wang RZ. Use of compound desiccant to develop high performance desiccant cooling system. International Journal of Refrigeration 2007;30:345–53. [19] Zhang XJ, Sumathy K, Dai YJ, Wang RZ. Dynamic hygroscopic effect of the composite material used in desiccant rotary wheel. Solar Energy 2006;80:1058–61. [20] Ge TS, Li Y, Wang RZ, Dai YJ. A review of the mathematical models for predicting rotary desiccant wheel. Renewable and Sustainable Energy Reviews 2008;12:1485–528. [21] Joudi KA, Dhaidan NS. Application of solar assisted heating and desiccant cooling systems for a domestic building. Energy Conversion and Management 2001;42:995–1022. [22] Mavroudaki P, Beggs CB, Sleigh PA, Halliday SP. The potential for solar powered single-stage desiccant cooling in southern Europe. Applied Thermal Engineering 2002;22:1129–40. [23] Halliday SP, Beggs CB, Sleigh PA. The use of solar desiccant cooling in the UK: a feasibility study. Applied Thermal Engineering 2002;22:1327–38. [24] Davanagere BS, Sherif SA, Goswami DY. A feasibility study of a solar desiccant air-conditioning system – part II. International Journal of Energy Research 1999;23:103–16. [25] Kays WM, London AL. Compact heat exchangers. 3rd ed. London: Mc-Graw Hill; 1984. [26] Kuehn TH, Ramsey JW, Threlkeld JL. Thermal environmental engineering. Prentice Hall; 2001. [27] Camargo JR, Ebinuma CD, Silveira JL. Experimental performance of a direct evaporative cooler operating during summer in a Brazilian city. International Journal of Refrigeration 2005;28:1124–32. [28] ASHRAE. Handbook: 1997 fundamentals, Inch-Pound editions. Atlanta: ASHRAE; 1997. [29] Maclaine-Cross IL, Banks PJ. Coupled heat and mass transfer in regenerators – prediction using an analogy with heat transfer. International Journal of Heat and Mass Transfer 1972;15:1225–42. [30] Jurinak J.J., Open cycle solid desiccant cooling – component models and system Ssmulations, Ph.D. Thesis, Univ. of Wisconsin-Madison; (1982) [31] Banks PJ. Prediction of Heat and Mass Regenerator performance using nonlinear analogy method: part 1 – basis. Transactions of the ASME. Journal of Heat Transfer 1985;107:222–9. [32] Press W, Teukolsky SA, Vetterling WT, Flannery BP. In: Numerical recipes. 2nd ed. Oxford: Cambridge University Press; 1996. [33] Balaras CA, Kontogiannidis S, Lykoudis S. Typical meteorological year for athens. National Observatory of Athens 2006. [34] Balaras CA, Grossman G, Henning H-M, Infante Ferreira CA, Podesser E, Wang L, et al. Solar air-conditioning in Europe – an overview. Renewable and Sustainable Energy Reviews 2007;11:299–314.