Potencial Of Solar Heat Pipe Vacuum Collector In The Desiccant Cooling Process.pdf

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Solar Energy 82 (2008) 1209–1219 www.elsevier.com/locate/solener

Potential of solar heat pipe vacuum collectors in the desiccant cooling process: Modelling and experimental results P. Bourdoukan a,*, E. Wurtz b, P. Joubert a, M. Spe´randio a a

LEPTAB Laboratoire d’Etude des Phe´nome`nes de Transfert Applique´ au Baˆtiment, Universite´ La Rochelle, Avenue Marillac, 17042 La Rochelle, France b LOCIE Laboratoire Optimisation de la Conception et Inge´nierie de l’Environnement, Campus Scientifique Universite´ de Savoie, 73376 Le Bourget du Lac, France Received 10 August 2007; received in revised form 6 June 2008; accepted 11 June 2008 Available online 11 July 2008 Communicated by: Associate Editor P. Gandhidasan

Abstract Desiccant cooling is an alternative technique to vapour compression systems. When thermally driven at moderate temperatures, it can be coupled to solar collectors. The use of flat-plate collectors and air collectors has demonstrated low efficiency in the coupling process and so a low potential of solar energy use in desiccant cooling. In this paper the use of heat pipe vacuum tube (HPVT) collectors in a solar desiccant cooling set up is investigated. First, a model for the collectors is proposed and experimentally validated under various operating conditions. A model of the storage tank taking into account thermal stratification is also validated. The experimentally evaluated efficiency of the HPVT collectors for one operating day varies between 0.6 and 0.7. Finally, simulation of the solar desiccant plant cooling a building is performed for different climates over a summer season. The solar fraction and the overall efficiency of the solar plant are calculated for this period and the potential of the vacuum tube collectors is evaluated for application to the desiccant cooling process. Ó 2008 Elsevier Ltd. All rights reserved. Keywords: Vacuum solar collectors; Desiccant cooling; Simulation; Experimental set up; Overall efficiency

1. Introduction Solar energy is used for space heating, hot water production and thermally driven air-conditioning systems including desiccant cooling systems. This air-conditioning technique would be an alternative to vapour compression systems (Jurinak et al., 1984); it has a low environmental impact since water is used as refrigerant and electrical consumption is limited to the auxiliaries. A solar desiccant cooling plant is shown in Fig. 1. It is a thermally driven open cooling cycle based on evaporative cooling and adsorption. With reference to Fig. 1, the cycle operates as follows: first, outside air (1) is dehumidified in a desiccant wheel (2); it is then cooled in the sensible rotary heat exchanger (3) by the return, cooled, air before undergoing *

Corresponding author. Tel.: +33 5 46458622; fax: +33 5 46458241. E-mail address: [email protected] (P. Bourdoukan).

0038-092X/$ - see front matter Ó 2008 Elsevier Ltd. All rights reserved. doi:10.1016/j.solener.2008.06.003

another cooling stage by an evaporative process (4). Finally, it is introduced in the room (5). The operating sequence for the return air (6) is as follows: it is first cooled to its saturation temperature by evaporative cooling (7) and then it cools the fresh air in the rotary heat exchanger (8). It is then heated in the heat exchanger fed by the solar energy tank (9) and if needed by a backup (10) and finally regenerates the desiccant wheel by removing the humidity before exiting the installation. Depending on the outside conditions, the air handling unit operates under two main modes: indirect evaporative cooling (IEV) for small cooling loads (return air humidifier and rotary heat exchanger) or desiccant mode (complete operating system, including the regeneration heat exchanger) for higher loads. For moderate summer conditions, in the morning when the outside temperature is low, the indirect evaporative cooling mode is able to keep the room in a comfortable range; there is thus no need for

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Nomenclature Symbols AHU air handling unit C heat capacity (J kg1 K1) E energy (J) e thickness (m) G solar global radiation (W m2) OE overall efficiency (dimensionless) H heat transfer coefficient (W K1 m2) IEV indirect evaporative cooling K heat loss constant (W K1 m2) L latent heat of vaporization (J kg1) m mass flow rate (kg s1) M mass (kg) p power (W) RH relative humidity (%) S area (m2) SF solar fraction (dimensionless) T temperature (K) u fluid velocity (m s1) UA global heat exchange coefficient (W K1) V volume (m3) W humidity ratio (gwater vapor kg1dry air)

regeneration and the solar heated water can be stored in the tank (Stabat, 2003; Maalouf, 2006). During the day with the rising outside temperature and increasing solar gains the indirect evaporative cooling cannot provide the cooling and so the desiccant mode is then required, while solar energy is needed for the regeneration of the

z e k g sa r q

coordinate in the fluid direction (m) emissivity (dimensionless) conductivity (W K1 m1) efficiency (dimensionless) transmission-absorptance coefficient Stefan Boltzmann constant (W K4 m2) density (kg m3)

Subscripts a ambiant b buffer c condenser f storage fluid in the manifold g glass H heat pipe i inlet p plate absorber o outlet r return from regeneration reg regeneration sat saturation

desiccant wheel. The minimum temperature required for regeneration depends on the nature of the desiccant material. It varies from 50 °C for lithium chloride to 60 °C for silica gel. We chose silica gel for its higher dehumidification performances in spite of the higher temperature needed for its regeneration.

Fig. 1. Solar desiccant cooling installation.

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The desiccant cooling process is well-suited to the requirements of non-residential buildings with high occupancy needing high air exchange rates, e.g. seminar rooms and banks (Henning et al., 2001). In these buildings the rooms are usually occupied during the day, so air-conditioning loads match solar energy availability, so coupling desiccant cooling with solar collectors would seem a very interesting option but it is important to investigate the efficiency of this coupling by studying the potential of solar energy in desiccant cooling. Numerical models are an effective way of predicting the thermal behaviour of solar collectors and thus the potential of solar energy in this process. Steady state models have certain advantages: they are simple, require little computing time and are easily coupled to an installation model. On the other hand, these simple models generally have the drawback of overestimating the solar potential, since the collectors do not usually operate in a steady state regime due to the variability of the driving factors (Schnieders, 1997; Isakson and Eriksson, 1991). In these cases, dynamic models are better at describing the collectors and at investigating the potential, as well as the control strategy, of the process using solar energy (De Ron, 1979; Perers, 1993; Henning, 1995). In recent decades collector modelling and the investigation of their potential have aroused increasing interest. Kamminga (1985) proposed a dynamic model by considering each part of a plate collector separately, i.e. the absorber, the glass cover and the fluid. Based on this model Schnieders (1997) proposed a model for direct flow vacuum plate collectors, and compared the energy-saving performance of five models. Praene et al. (2005) applied Schnieders’ model to direct flow vacuum tube collectors, more accurately taking into account the heat transfer by radiation inside the tube. Henning et al. (2001) investigated the potential of flatplate collectors and air collectors for the summer season in desiccant cooling using simulation and experimental results. Wurtz and Maalouf (2006) compared the operation of autonomous and solar-assisted desiccant cooling systems using a model based on the Hotel–Whillier–Bliss equation. Few works concerning the modelling and the potential of an evacuated heat pipe vacuum tube collector using an evaporating-condensing device have been published. And yet these collectors are very well-suited for high temperature processes such as thermally driven air-conditioning systems. Due to vacuum, their losses are limited and the tube can operate at a high efficiency even for elevated inlet water temperatures. In this paper a dynamic model of these collectors and also of a stratified tank is proposed and validated experimentally under various operating conditions. The experimental efficiency of the HPVT collectors in desiccant cooling operations is calculated and the potential of these collectors for desiccant cooling applications is evaluated for three different locations (varying from hot and humid to moderately humid) over a summer season and for typical summer days.

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2. Modelling of the solar installation In order to study the potential of HPVT collectors in the solar desiccant cooling process, a model of the above described solar installation was developed. 2.1. Heat pipe vacuum tube In an evacuated heat pipe collector (Fig. 2), a sealed copper pipe containing a vaporizable fluid is bonded to a copper fin plate absorber located inside a glass tube. A small copper condenser is attached from one side to the top of the heat pipe and from the other side to the storage working fluid. The heat pipe is an evaporating-condensing device. As the sun shines on the absorber, the pipe is heated and some of the liquid inside evaporates. The vapour rises toward the condenser at the top of the heat pipe and condenses on being cooled by the storage water circulating in the manifold. The liquid then returns to the heat pipe. The vacuum tube ensures minimization of the heat losses of the collector. A model for the HPVT is proposed below, separately considering each component of the tube, i.e. the glass cover, the absorber, the heat pipe fluid, the condenser and the storage fluid. The following assumptions are made regarding the model:  The properties of the materials are independent of the temperature.  The temperature gradient along the absorber and the condenser is negligible.  Due to the sufficient quantity of fluid in the heat pipe (more than 90% of the heat pipe volume), the vapour is not superheated (Gidas, 1971; Bricard and Chaudourne, 1997).  The liquid returning from the condenser to the heat pipe is saturated.

Fig. 2. Heat pipe vacuum tube.

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 The glass cover is clean and completely transparent to solar radiation.  The part of radiation reflected by the absorber leaves the tube.  Due to the vacuum, convection does not occur inside the tube.  Conduction in the storage fluid direction (z) is considered to be negligible. The glass cover exchanges heat by convection with the outside air and by radiation with the sky and the absorber. The absorber receives solar radiation and heats the fluid inside the tube. The vapour rising from the heat pipe enters the condenser and releases energy to the circulating storage fluid, then exits the condenser as a saturated liquid. The heat pipe liquid is heated to saturation before the evaporation process begins. The governing equations for each component (characterized by heat capacity C and temperature T) then read Glass cover: M gCg

dT g ¼ eg rS g ðT 4sky  T 4g Þ þ ep rS p ðT 4p  T 4g Þ dt þ S g hg ðT a  T g Þ

ð1Þ

Absorber: M pCp

dT p ¼ ep rS p ðT 4g  T 4p Þ þ GsaS p þ S H hH ðT sat  T p Þ dt ð2Þ

Vapour flow rate: mv L ¼ S H hH ðT p  T sat Þ Condenser:   dT c M c Cc ¼ mv L þ S c hc ðT f  T c Þ dt Storage fluid:   oT f oT f þu ¼ S c hc ðT c  T f Þ m1 C f ot oz

ð3Þ

ð4Þ

ð5Þ

2.2. Storage tank The investigated solar installation is shown in Fig. 1. In order to take into account thermal stratification in the storage tank, it is divided into n nodes (Klein, 1976). The following equations apply to the first node, the last node (n = 12) and the ith current node: q1 V 1 C f

dT 1 kS i ¼ m1 C f ðT i1  T 1 Þ þ ðT 2  T 1 Þ dt e12 þ m2 C f ðT 2  T 1 Þ þ S 1 K 1 ðT a  T 1 Þ

ð6Þ

qn V n C f

dT n kS i ¼ m1 C f ðT n1  T n Þ þ ðT n1  T n Þ dt en;n þ m2 C f ðT i2  T n Þ þ S n K n ðT a  T n Þ

qi V i C f

dT i kS i ¼ m1 C f ðT i1  T i Þ þ ðT i1  T i Þ dt ei þ m2 C f ðT iþ1  T i Þ þ S i K i ðT a  T i Þ kS i ðT i  T iþ1 Þ  ei

ð7Þ

ð8Þ

For the heat exchanger model, a counterflow configuration and an NTU efficiency model were considered (Kays and London,1984; Incropera and Dewitt, 1996). The transfer coefficient (UA) for the heat exchangers was evaluated from the experimental installation. The whole set of equations describing the solar installation (Eqs. (1)–(8)) is introduced into SPARK (SPARK, 2003), a general simulation environment that supports the definition of simulation models and permits the solution of these models via a robust and efficient differential/algebraic equation solver (Sowell and Haves, 2001). In the SPARK environment, the modeller describes the set of equations defining a model as equation-based objects. At the lowest level, an atomic object characterizes an equation and its variables. Then macroscopic objects can be created as an assembly of various atomic or macroscopic objects. The whole model is built by connecting the different elementary objects. Any class of objects can be reused as many times as necessary, without any additional effort. At this stage, we should note that a model implemented in the SPARK environment is input/output-free. So the particular problem to be solved is described by imposing adequate input data (boundary and initial conditions) and by specifying the variables that are to be solved. In fact, it is not necessary to order the equations or to express them through assignment statements (algorithms), unlike in conventional modular environments (Wurtz et al., 2006). SPARK uses a mathematical graph of the model in order to break it down into it in strong components to be solved independently. Within each component, SPARK finds the appropriate calling sequence for obtaining the solution. If a direct sequence is not possible, as indicated by a cyclic problem graph, a small ‘‘cut set” is determined to minimize the number of variables involved in the Newton-like iterative numerical solution process. As a result, this decreases the size of the Jacobian matrix involved in the Newton iteration within the component. Consequently, the way SPARK handles the solution of coupled nonlinear equations makes it a fast and robust solver for the simulation of energy problems in buildings. After the ‘‘cut sets” of variables have been identified, the problem specification file is automatically converted into a C++ program which is then compiled, linked and executed for solving the problem for a given boundary and initial conditions set.

P. Bourdoukan et al. / Solar Energy 82 (2008) 1209–1219

3. Validation of the model 3.1. Experimental setup An experimental solar installation (Fig. 3) similar to that shown in Fig. 1 was set up and used to validate the solar installation model. The installation consists of 40 m2 of heat pipe vacuum collectors (200 tubes), a 2500 l storage tank with reinforced insulation and a plate heat exchanger. Temperatures and flow rates are measured at the inlet and outlet of each component (solar collectors, heat exchanger and storage tank). The overall solar radiation is measured in the plane of the collectors. 3.2. Validation of the collector model Three different recorded days (the data (T, m and G) is that measured on the experimental installation in La Rochelle Tsky is given by Meteo France for these recorded days) were used to validate the collector model. The first day is representative of typical summer conditions, the second of atypical solar radiation and the third of desiccant cooling load conditions. Once the solar radiation in the collectors’ plane, the outside temperature, and the inlet temperature were recorded, the computed and measured collector outlet temperatures were compared. Fig. 5 plots the calculated and measured collector outlet temperature for the three studied days. Solar radiation is also plotted on the figure. Comparison between the computed and the measured temperatures for the typical summer day conditions (day 1) shows the model’s high performance in predicting collector outlet temperature: the absolute error of

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the model varies between 0 °C and 0.7 °C, with a corresponding maximum relative error of less than 1%. This accuracy is due to the fact that each component is taken into consideration by the model and the calculations are performed in each vacuum tube simultaneously (see Fig. 4). While collector outlet temperature can thus be predicted accurately in normal radiation conditions, it is very important to study the performance of the model for atypical conditions. By comparing the calculated and the measured outlet temperatures for the atypical conditions day (day 2), when the overall solar radiation fluctuated constantly, it can be seen that the model very accurately predicts the outlet temperature for the whole day, whatever the amplitude of the solar radiation fluctuations. The maximum absolute error in these conditions is 1.8 °C, i.e. a relative error of 2%. In these first two cases, the model correctly predicts the collector outlet temperature under different radiation conditions. The next step in the validation procedure is to study the model’s response to a desiccant cooling load, i.e. storage in the morning, with regeneration when the desiccant mode is enabled. This is obtained as follows, (with reference to Fig. 1): in the morning, while the outside temperature is below 30 °C, the desiccant mode is not required, so the heated water is stored in the tank (m1 > 0 and m2 = 0). When the outside temperature exceeds 30 °C, the desiccant mode is needed and the outlet valve of the buffer is opened (m2 > 0). Flow rate m2 is regulated to obtain the desired regeneration power. Under typical desiccant cooling load (day 3) there is remarkable agreement between the predicted and the measured outlet temperatures, for both storage and regenera-

Fig. 3. The experimental solar installation at La Rochelle.

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tion periods. The absolute error is less than 0.5 °C except very early in the day, when it is 1.4 °C.

3.3. Validation of the tank model The temperature and the flow rates at the two inlets and outlets of the tanks are measured, so the calculated temperatures at the top and the bottom of the buffer (node 1 and node n) can be compared to the measured ones. Two different scenarios were used to validate the tank model. The first concerns only storage (m1 > 0, m2 = 0) and the second considers typical desiccant conditions, e.g. storage in the morning and both storage and regeneration in the afternoon (m1 > 0, m2 > 0). Fig. 5 shows a comparison between the calculated and the measured temperatures at the top and the bottom of the buffer for these two operating modes. From this figure, it can be seen that the tank model is able to take into account the thermal stratification occurring in the buffer. When operating under storage only, the stratification remains fairly constant, but increases when storage and regeneration are combined. This may, at first, appear incoherent, as more mixing occurs in the latter case. The fact that stratification is greater with the regeneration load is due to the return temperature from regeneration, Tr, which

is significantly below the temperature of the bottom of the buffer. In both cases, the maximum error between the calculated and the measured temperatures of the tank is 2 °C except during the transition period between storage only and regeneration, where the error can reach 3 °C. The potential of the HPVT collectors in solar desiccant cooling is evaluated in the following section. 4. Potential of HPVT in desiccant cooling 4.1. Experimental efficiency of HPVT The efficiency of the HPVT is defined as the ratio between the energy delivered by the collectors and the solar energy they receive. With reference to Fig. 1, this efficiency reads g¼

Edelivered m1 C f ðT o  T i Þ ¼ Ereceived G:S

ð9Þ

The experimental efficiency of the HPVT collectors for the day of desiccant cooling operation is shown in Fig. 7. We see that the efficiency of the tubes varies between 0.6 and 0.7. The higher value is obtained for the minimum temperature of the collectors. It can be seen that even at high operating temperatures, the efficiency never falls below

P. Bourdoukan et al. / Solar Energy 82 (2008) 1209–1219 Desiccant load

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0.6. This represents the major advantage of vacuum collectors compared to traditional flat-plate collectors, whose efficiency significantly decreases with increasing operating temperature.

La Rochelle (Germany).

4.2. Overall efficiency and solar fraction

4.3.1. Simulation conditions In order to study the potential of HPVT collectors for solar desiccant cooling applications, simulations are performed for different locations, of the solar desiccant cooling installation similar to that of the Fig. 1 that cools an office building of a total gross area of 1800 m2 and a conditioned area of 1350 m2. It is occupied from 8 am to 6 pm, 5 days a week. The structure is thermally insulated (U = 0.52 W/m2) and the windows are double glazed. The internal gains are lightings (15 W/m2), .equipments (14 W/m2) and occupants (5 W/m2). The building is modelled by zonal method (Wurtz et al., 2006). For the air handling unit, the desiccant wheel model is that proposed by Maclaine-cross and Banks (1972) and improved by Stabat and Marchio (2008), for the sensible regenerator Kays and London (1984) model is considered while the humidifiers are modelled by constant wet bulb temperature. The locations chosen are La Rochelle in France (hot and humid with a mean humidity ratio of 13.5 g/kg) Bolzano in Italy (moderately humid with a mean humidity ratio of 11.7 g/ kg) and Berlin in Germany (relatively dry in comparison with La Rochelle and Bolzano with a mean humidity ratio of 8.7 g/kg). Locations with humidity over 14.5 g/kg are not considered since it becomes inefficient to control the humidity inside when the desiccant system is coupled to solar collectors since the needed regeneration temperature tends to increase significantly (see Fig. 6). The control strategy is following (Bourdoukan et al., 2008): In the morning the system starts with the indirect evaporative cooling (IEV) mode (return humidifiers and the sensible regenerator) and when the temperature of the room exceeds 26 °C or the humidity ratio exceeds the

An important property of a solar installation is its overall efficiency (OE), which is the ratio of the solar energy consumption of the system for regeneration to the solar energy received by the collectors. By integrating the regeneration power and the solar power received over a particular period we obtain the energy consumed for regeneration (E reg,solar) and the energy received (Ereceived) and so the overall efficiency can be calculated Z P reg;solar  dt ð10Þ Ereg;solar ¼ regeneration

P reg;solar ¼ m2 C f ðT 1  T i2 Þ Z Greceived  dt Ereceived ¼

ð11Þ ð12Þ

day

OE ¼

Ereg;solar Ereceived

ð13Þ

Another important property of a solar installation coupled to a thermal process is the solar fraction of the installation, defined as the ratio of the energy provided by the collectors for regeneration to the total energy needed for regeneration, including the auxiliary. The solar fraction is expressed as SF ¼

Ereg;solar Ereg;solar þ Ereg;backup

ð14Þ

In the following section, the potential of HPVT collectors in solar desiccant cooling is evaluated by determining the solar fraction and the overall efficiency of the collectors over a summer season and for typical summer days for

(France), Bolzano (Italy)

and Berlin

4.3. Potential for desiccant cooling applications

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10.8 g/kg the system switches to the desiccant (DEC) mode (all the components are operational). The temperature and the humidity of the conditioned space are controlled by the mean of the regeneration temperature. By increasing the regeneration temperature the dehumidification capacity is increased and thus more sensible cooling is possible by increasing the efficiency of the supply humidifier. In reversal to increase the latent capacity of the system, by increasing the regeneration temperature and by maintaining the same efficiency of the supply humidifier the sensible cooling capacity is maintained and the latent capacity is increased due the increase of the dehumidification capacity. If the temperature of the water at the outlet of the buffer (T1 in Fig. 1) is high and there is no need for high regeneration temperature a part of the water flow rate is bypassed to keep the regeneration temperature close to 60 °C. The backup is used if the solar regeneration temperature (position 9 in Fig. 1) falls below 60 °C which is the required regeneration temperature for the silica gel, or if there is a need to increase the regeneration temperature with the increasing load. 4.3.2. Simulation results The first step in the potential estimation is the evaluation of the overall efficiency and the solar fraction of HPVT installation coupled to a desiccant unit. Simulations of the solar desiccant plant, cooling the described building powered by HPVT collectors for the three mentioned locations are performed. Since the cooling load differs for a location to another the collector area for each location depends on this load so for La Rochelle the collector area is 300 m2, for Bolzano it is 245 m2 and for Berlin 205 m2. Fig. 7 shows the simulation results of the hottest days of the season for the three locations. For each case is plotted in the left series of Fig. 7 the outside temperature (T out-

side), the supply temperature (T5), the inside temperature (T6), the inside relative humidity (RH6) and the operating mode of the AHU (20 for IEV and 40 for the DEC). For the right series of Fig. 7 are plotted the temperature at the top of the buffer (T buffer), the temperature of the regeneration air before the backup (T9), the temperature of the regeneration air after the backup (T10), the position of the 3 way valve (valve position) and the fraction of the backup (backup fraction). Fig. 7 shows that for all the cases the temperature and the humidity are both controlled and maintained at an acceptable level (T < 26.5 and w < 11.8 g/kg), and the collectors can provide the major part of the needed regeneration energy. For La Rochelle conditions (humidity ratio for the hottest day is 13.4 g/kg) the systems operates only one hour in the early morning under the IEV mode since the sensible load is high with elevated outside temperature that limits the use of the IEV beside the high outside humidity ratio that reduces the potential of the IEV when elevating the inside humidity by introducing outside air at high humidity ratio. The system switches to the DEC mode and is able to keep the inside temperature near to 25 °C for the whole day and the relative humidity close to 55% (humidity ratio is then 11.7 g/kg). Even that the outside temperature reached 34 °C during the day the system was able to provide the cooling load for the whole day. The supply 3 way valve for the regeneration heat exchanger was completely open since during the DEC mode with constant need for regeneration and since the DEC mode started early in the morning it limited the buffer temperature. The HPVT collectors were able to provide most of the regeneration energy and the backup was operational for limited periods during the day, first in the morning when the collectors were not yet operational and during the peak load in the early after noon. When the backup was operational the fraction of the backup power was always below 28%. For this day in La Rochelle the calculated OE (Eqs. (10)–(13)) is 63% and the SF (Eq. (14)) is 93%, while for the whole summer season the calculated OE is 54% and the SF is 96%. The difference in the OE on the day and season basis is that for the season basis takes into account the received energy during weekend where the HPVT are not efficiently used (only for storage) which lowers the OE beside that during this particular day the DEC mode was operational during 9 h which increases the efficiency with the increase of the time of use (Bourdoukan et al., 2007). The increase in the SF from the day to the season basis is related to the fact that during the hottest days the backup is needed more than the other days of the season. For Bolzano (outside humidity ratio is 11.4 g/kg) in the early morning the outside temperature was low and the solar global radiation was not typical of a summer period which gave a small cooling load in the beginning of the day and the IEV was very efficient which delayed the DEC mode to the afternoon. The inside temperature was

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Time (hours)

Fig. 7. Performance of HPVT collectors coupled to a desiccant system for three different locations.

kept below 26 °C for the whole day and the inside relative humidity was fluctuating depending on the inside temperature (but the inside humidity ratio was always below 11.8 g/kg). Since the DEC mode was delayed more storage was possible in the buffer which increased its temperature. In this case the supply 3 way valve of the regeneration heat exchanger was regulating the flow rate depending on the load demand. For Bolzano and for this particular day the backup was needed only at the end of the day. When

it was operational the fraction of the backup power was below 20% of the total needed regeneration power. For this day the OE is 50% and the SF is 87% while for the season the OE is 51% and the SF is 97%. For this day the OE is relatively low due to the use over a short period of solar energy due to the delay of the DEC mode, which reduces the OE (Bourdoukan et al., 2007). For Berlin (outside humidity ratio is 9.5 g/kg) the scenario of the hottest day is similar to Bolzano. The DEC

was delayed due to the low cooling load in the morning and the low outside humidity ratio. In this case the inside temperature was below 25 °C for almost the whole day and the humidity ratio below 11 g/kg. The 3 way valve was completely opened only on peak load at noon and at the very end of the day with the decrease of the buffer temperature which enabled the backup with a fraction below of 10%. In this case the OE is 56% and the SF is 96.7%, while for the season the OE is 51%. For the studied cases the control strategy was able to keep the temperature and the humidity below the comfort constraint. For all the cases if the DEC mode is enabled when the inside temperature reaches a value below 26 °C (for example 24 °C instead of 26 °C), this will produce more cooling during the day and the inside temperature will be a bit lower but in reversal this strategy will consume early the energy and so the backup will be used very early in the afternoon which will decrease the SF significantly with a negligible increase in the OE and in the thermal comfort. Same, if the flow rate (at the outlet of the buffer) for the regeneration heat exchanger is not controlled by the 3 way valve the regeneration temperature will be higher and thus more cooling will be produced but the storage will be consumed rapidly and the backup will be used early and will decrease the SF. For the three cases on a season basis the SF is greater than 96%, and the OE is greater than 51%. Coupling the HPVT collectors with a desiccant cooling plant thus appears as a very interesting option. 4.4. Comparison with flat-plate collectors To compare the performance of HPVT and flat-plate collectors, simulations were carried on the same system that cools the same building, this time powered by flatplate collectors. For the same day in La Rochelle considered in the previous paragraph, Fig. 8 shows the energetic performance of the flat-plate collectors installation coupled to the desiccant system. When using these collectors the same inside conditions can be achieved but in this case the backup is used for the whole day and the backup power reaches 50% of the total regeneration power in the early morning. The OE of flat-plate collectors for this day is 44% and the SF is 72%. For the season the OE is 36% and SF is 75%. For Berlin and Bolzano the flat-plate collectors operate with an OE of 33% and a SF 77%. In order to attain the same performance as for the HPVT collectors we must increase the flat collector’s area by 20–25% which will cause a space constraint, especially for multi-storey building. Beside the collectors which are used for cooling during summer will be used for heating during winter, or HPVT are more efficient in handling the diffuse radiation and operates almost independently of outside temperature due to vacuum while the losses of flat-plate collectors increase significantly with the decrease of outside temperature. For the investment issue the HPVT are still more expensive than the flat-plate collectors but

La Rochelle (France)- flat plate collectors 80 100

70

90 60

80

50

T1 buffer T9 reg solar T10 reg backup backup fraction valve position

40 30

70 60 50 40 30

20

20 10

10

0

0 8

9

10 11 12 13 14 15 16 17 18 Time (hours)

Backup power fraction or valve position(%)

P. Bourdoukan et al. / Solar Energy 82 (2008) 1209–1219

Temperature (°C)

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Fig. 8. Performance of flat-plate collectors in desiccant cooling application for La Rochelle (France) weather conditions.

since it is a new technology with the coming years the price will decrease with the increasing demand. This clearly indicates the importance and the advantage of HPVT collectors in solar desiccant cooling. 5. Conclusion In this paper, a model of heat pipe vacuum tubes using an evaporating-condensing device was proposed and implemented in the SPARK simulation environment. The model has been validated for different conditions by comparison with an experimental solar installation and has accurately predicted collector outlet temperature with a maximum error of 2% in irregular solar conditions and an error of less than 1% in normal conditions. A tank model was also developed and validated for predicting the temperature at the top and bottom of the buffer. The experimentally observed efficiency of the collectors for desiccant cooling conditions varies between 60% and 70%. Finally, the potential of heat pipe vacuum tube collectors for a solar desiccant plant that cools an office building was investigated. A simulation over a summer season showed the capacity of a desiccant plant to maintain the building temperature and humidity at a comfort level for three different locations going from relatively dry to hot and moderately humid. These simulations render a seasonal overall efficiency greater than 51% and a seasonal solar fraction of 96% while the daily OE varies from 50% to 64% and the daily solar fraction varies from 87% to 97%. Acknowledgements This work was supported by ADEME (French Agency for Environment and Energy Management) and the Regional Council of Poitou-Charentes. The authors would like

P. Bourdoukan et al. / Solar Energy 82 (2008) 1209–1219

to thank Mr Michel Burlot for his valuable technical support on the experimental installation. Appendix A. A1-Values of the physical properties for each vacuum tube Symbol

Value

Mg (kg) Cg (J kg1 K1) eg (–) ep (–) Sg (m2) hg (W K1 m2) Cp (J kg1 K1) Mp (kg) Sp (m2) SH (m2) hH (W K1 m2) MC (kg) CC (J kg1 K1) hC (W.K1 m2) SC (m2) sa (–)

2.94 815 0.9 0.115 0.793 4 390 1.6465 1.85 0.0953 41.8 0.15 390 21.4 0.015 0.8835

A2-. List of components of the La Rochelle solar installation 40 m2 of heat pipe vacuum tube collectors. Plate heat exchanger. Storage tank (2500 l) with reinforced insulation. Three water flow meters to measure m1, m1 and m2 (Fig. 1).  7 Pt100 temperature sensors to measure the temperature TO, Ti, T1, Ti1, Ti2, Tn, and Toutside (Fig. 1).  Two radiation meters to measure the solar global radiation G.

   

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