Selective water sorbent for solid sorption chiller: experimental results and modelling

International Journal of Refrigeration 27 (2004) 284 293 www.elsevier.com/locate/ijrefrig Selective water sorbent for solid sorption chiller: experimental results and modelling G. Restuccia a, *, A. Freni a, S. Vasta a, Yu Aristov b a CNR- Istituto di Tecnologie Avanzate per l Energia Nicola Giordano, S. Lucia sopra Contesse, 98126 Messina, Italy b Boreskov Institute of Catalysis, Pr. Ak. Lavrentieva 5, Novosibirsk, 630090 Russia Received 10 January 2003; received in revised form 16 July 2003; accepted 4 September 2003 Abstract In this paper the experimental results of a lab-scale chilling module working with the composite sorbent SWS-1L (mesoporous silica gel impregnated with CaCl 2 ) are presented. The interesting sorption properties of this material yield a high COP=0.6 that gives a promising alternative to the common zeolite or silica gel for application in solid sorption units driven by low temperature heat (T 4100 C). The measured low specific power of the device is a result of not optimised geometry of the adsorber and of the pelletised shape of the adsorbent. Heat transfer optimisation is currently under progress to increase the specific power. The experimental results are compared with those of a mathematic model able to describe the dynamic behaviour of the system. The model is used to study the influence of the main operating parameters on the system performance. # 2003 Elsevier Ltd and IIR. All rights reserved. Keywords: Adsorption system; Silica gel; Calcium chloride; Water; Experiment; Performance; Modelling Refroidisseur à adsorption solide muni d un adsorbeur permettant l adsorption sélective d eau : re sultats expérimentaux et mode lisation Mots cle s :Système à adsorption ; Silice ; Chlorure de calcium ; Eau ; Expérimentation ; Performance ; Modélisation 1. Introduction The air conditioning systems based on the adsorption/desorption process represent an important alternative to the traditional technology of vapour compression. The gas solid systems, in fact, have a low environmental impact and use heat as an energy source, therefore they allow a diversification in the use of the * Corresponding author. Tel.: +39-090-624229 Fax: +39-090-624247. E-mail addresses: giovanni.restuccia@itae.cnr.it. primary energy sources (e.g. heat wastes or methane instead of electricity) [1,2]. Recently, a new family of composite sorbents called Selective Water Sorbents (SWSs) was presented for sorption cooling and heating [3 5]. A typical SWS is a two-component material based on a porous host matrix and an inorganic salt impregnated inside pores. A large variety of both host matrices (such as mesoporous or microporous silica gels, aluminas, porous carbons, polymers, etc.) and salts (CaCl 2, LiBr, MgCl 2, LiCl, etc.) gives a wide possibility to change the sorbent properties in a wide range to fit the demands of particular applications among which are sorption heat pumps and refrigeration machines discussed in this 0140-7007/$35.00 # 2003 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2003.09.003

G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 285 Nomenclature H Sorption enthalpy (J kg 1 ) A Heat exchange surface area (m 2 ) C Thermal capacity (J K 1 ) c p Specific heat (J kg 1 K 1 ) e L(T) Efficiency Latent heat of condensation/evaporation (J kg 1 ) m : Mass flow (kg s 1 ) m Mass (kg) P Pressure (Pa) Q Heat (J) R Universal gas constant (J kg 1 K 1 ) t Time (s) T Temperature ( C) U Global heat transfer coefficient (W m 2 K 1 ) w Uptake (kg kg 1 ) cycle Cycle time (s) Subscripts a Isosteric heating phase b Desorption phase bed Adsorbent bed c Isosteric cooling phase con Condenser d Adsorption phase eq equivalent, i.e. of the adsorbent and the uptake ev Evaporator ex Heat exchanger f Heat transfer fluid in Inlet out Outlet s Sorbent paper. Among the different SWSs, the SWS-1L (mesoporous silica gel impregnated with CaCl 2 ) showed a very high water sorption ability (up to 0.7 g of water per 1 g of dry sorbent), that results in a high heat storing capacity (up to 2000 kj kg 1 ). Furthermore, according to the SWS-1L water sorption isobars, most of the sorbed water can be removed at temperatures of 90 100 C [6,7]. The sorption equilibrium curves of the SWS-1L showed a combination of heterogeneous adsorption, chemical reaction with formation of solid crystalline hydrates of confined salt and liquid absorption. The complex nature of the sorption poses great difficulties for its theoretical description, so that in a simplified manner it is possible to say that SWS-1L shows a divariant behaviour typical for liquid solutions of CaCl 2 for a water uptake higher then 2 moles of water per mole of salt. While for lower water uptake the formation of solid hydrates was recognised. Nevertheless, the properties of SWS-1L in the typical operating conditions of a refrigeration unit are macroscopically similar to a solid sorbent and its sorption properties can be described by the isosteric curves. The thermodynamic performances of SWSs for application in air conditioning systems were evaluated by means of a specific model and compared with those obtained using other adsorbents [5, 8]. The results showed that these new materials allow us to obtain COP (0.7 0.8 for basic cooling cycle, 1.4 1.6 for basic heating cycle) respectively 30 and 60% higher than those calculated at the same regeneration temperature (about 95 C) for the 4A zeolite water or a micro-porous silica gel water pairs. Another important benefit is that the relatively low temperature required for regeneration, yields SWSs very attractive for utilization of low grade heat sources, such as solar heat, industrial waste heat, automotive exhaust gas, or cascade and tri-generative systems. The main aim of this study is to evaluate experimentally the performance of the SWS-1L for solid sorption air conditioning and to compare the experimental results with a dynamic simulation. In the following, a lab-scale single bed module working with the pair SWS-1L/water is described, and the results of its experimental testing for cooling are presented. Finally, a predictive dynamic model of the unit operation is presented and a comparison is made between experimental and calculated results. 2. Experimentation The lab-scale prototype consists of an adsorbent bed connected to evaporator and condenser by means of vacuum valves (Figs. 1 and 2). The sorbent material consists of 1.1 kg of dry SWS-1L granules (0.8 1.6 mm size). The composite sorbent was prepared, by filling pores of mesoporous silica gel with an aqueous solution of CaCl 2, according to the procedure described in [6, 7]. The calcium chloride content in the anhydrous sample was measured to be 33.0 wt.%. The heat exchanger is made of finned stainless steel tubes placed inside the sorbent and its main features are presented in Table 1 and Fig. 3. The adsorbent bed is placed in a cylindrical vacuum chamber connected, through vacuum piping, to a vacuum pump, an evaporator, and a condenser. An external circuit allows the heating/cooling of the adsorbent bed. It consists mainly of four thermocryostats that simulate the heat sources/sinks. The temperature measurements are performed through several thermocouples (T-type) properly

286 G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 placed: four inside the bed, into the evaporator, into the condenser and in the heat transfer circuit at the inlet/ outlet of the evaporator, condenser and bed heat exchanger (Fig. 1). Three vacuum gauges are connected to the main components of the prototype. Furthermore, Fig. 1. Scheme of the tested lab-scale chiller: 1 adsorbent bed, 2 condenser, 3 evaporator, 4 5 evaporator condenser thermocryostats, 6 7 heating cooling fluid thermostated baths, 8 vacuum pump, F flow-meters, T temperature sensors, P vacuum gauges. three flow-meters are used to measure the flow rate of the thermal vector fluids. The total amount of water adsorbed by the sample is obtained by directly measuring in the evaporator the amount of water evaporated. The data acquisition system consists of an analogical-todigital converter and of a personal computer; the recorded data are those coming from the above described sensors. So that, all the energies supplied to (or removed from) the three main components of the sorption system are continuously measured during the tests. The experiments were carried out at the standard conditions of a sorption air conditioner. In particular, the experimental cycle was followed by fixing the following criteria. The heating/cooling isosteric steps were considered completed when the pressure of the adsorbent reached the condenser/evaporator pressure respectively. The end of the isobaric desorption/adsorption phases was considered when the mean temperature of the bed was equal to the fixed maximum and minimum temperature. The experimental cooling COP was calculated as ratio between the useful cooling effect and the energy supplied by the external source: Table 1 Features of the adsorber Number of tubes 8 Length of tubes, mm 200 Fin spacing, mm 7 Fin thickness, mm 0.4 Exchange area, m 2 0.4 Metal/adsorbent mass ratio 3 Fig. 2. View of the tested lab-scale chiller. Fig. 3. View of the adsorber without and with the adsorbent.

P ½m : fc pf ðt in T out ÞŠ ev Dt d COP ¼ P ½m : fc pf ðt in T out ÞŠ bed Dt aþb where d is the isobaric adsorption phase of the cycle and a+b means that the summatory is extended to the isosteric heating (a) and the isobaric desorption (b) steps. The previous formula shows that the energies measured are those transferred from the heat transfer fluids to the evaporator and to the adsorber respectively. Thus it is necessary to single out that the measured values represent the effective gross COP of the prototype, that is affected by the various heat losses, the heat capacity of the inert masses and the heat exchangers efficiency. So that, this COP is a practical value that depends on the design of the experimental device. The specific cooling power P S was calculated as: P ½m : fc pf ðt in T out ÞŠ ev Dt d P S ¼ ð2þ m s cycle where cycle is the total cycle time. This parameter strongly depends on the heat and mass transfer efficiency, so that it can be considered as the indicator of the dynamic efficiency of the device, that is not optimised for the unit studied. G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 287 ð1þ the heating fluid. The water desorbed from the bed flows to the condenser. During the cooling run [phases (c) and (d)] the heat is transferred from the bed to the cooling fluid, as well as the water from the evaporator to the adsorber. Further details on the operating principles and on the thermodynamic of the sorption cycle can be easily found in literature [9]. Being the sorption prototype a single bed system with pelletised sorbent material, we developed a simplified model that includes the thermal capacities of the various components and considers the adsorbent as a black box. The latter assumption leads to consider the heat transfer by means of a global heat transfer coefficient which accounts the thermal conductivities and the convective coefficients inside the bed. Furthermore, being the heat transfer the limiting factor for pelletised bed [10], the mass transfer resistance has been neglected in the model. In order to simplify the simulation, each component (adsorbent bed, condenser, evaporator) is considered to be homogeneous. Furthermore, the properties of the metal, the water vapour and the thermal vector fluids are assumed to be constant. The thermal losses from the components are neglected. Based on the previous assumptions, the governing equations for each phase of the sorption cycle can be written as follow: 3. Simulation model The main parameters considered by the model for each component are presented in Fig. 4. The basic sorption cooling cycle, is composed of four steps, namely: (a) isosteric heating; (b) isobaric desorption; (c) isosteric cooling and (d) isobaric sorption. During the phases (a) and (b) heat is supplied to the sorbent bed by 3.1. Isosteric heating/cooling dt s m s ð1 þ wþ c peq þ m ex c pex dt ¼ m : fc pf " T f bed T s bed m : fc pf " T f bed T s bed ¼ m: fc pf ðt in T out Þ bed ð3þ ð4þ where e is the heat exchanger efficiency calculated according to Eq. (9), T in and T out are respectively the inlet and outlet temperature of the thermal vector fluid in the adsorber heat exchanger, T f-bed =(T in +T out )/2 is the average temperature of the thermal vector fluid inside the heat exchanger, w is the uptake calculated according to Eq. (10). 3.2. Isobaric desorption/adsorption dt s m s ð1 þ wþ c peq þ m ex c pex dt m s ¼ m : fc pf " T f bed T s bed dw dt DH ð5þ Fig. 4. Schematisation of the studied system and of the parameters considered for each component. m : fc pf " T f bed T s bed ¼ m: fc pf ðt in T out Þ bed ð6þ

288 G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 3.3. Energy balance for the condenser/evaporator dw m s LT ð Þþc pv ðt s TÞ dt con=ev ¼ m : dt fc pf " T T f con=ev þ C dt con=ev ð7þ Q ev P s ¼ ð15þ cycle m s The differential equations were solved with the implicit finite difference method of Crank Nicholson, described in Ref. [12]. m : fc pf " T f con=ev T ¼ m : fc pf ðt in T out Þ con=ev con=ev where T is the temperature inside the condenser or evaporator and T f-con/ev =(T in +T out )/2 is the average temperature of the thermal vector fluid inside the heat exchanger of the condenser/evaporator. The efficiency e of the condenser, evaporator and bed heat exchanger is calculated as follows: " ¼ 1 e : U A m f cp f ð9þ The adsorbent/adsorbate equilibrium is represented by the following equation [5,6]: lnp ¼ Aw ð Þþ Bw ð Þ ð10þ T while the sorption/desorption enthalpy H depends on the uptake and is obtained from the relationship reported in [5,6]. The equivalent specific heat of the adsorbent c peq is expressed as a function of water uptake and temperature as reported in [5,11]; it represents the specific heat of the adsorbent and of the water adsorbed as a whole. The energy required for the heating-cooling of the bed during steps (a) and (c) is calculated as: ð Q a=c ¼ m : fc pf " T f bed T s ð11þ a=c bed dt The energy required for the isobaric phases is calculated as: ð Q b=d ¼ m : fc pf " T f bed T s bed dt b=d ð þ m s DH dw ð12þ b=d The heat of condensation-evaporation is calculated as: ð Q con=ev ¼ m : fc pf " T f con=ev T ð13þ b=d con=ev dt Finally, the cooling COP and the specific cooling power of the machine are calculated as: Q ev COP ¼ Q a þ Q b ð8þ ð14þ 4. Experimental results and comparison with the model Most of the experiments have been performed to investigate several cycles: the maximum value of the average temperature of the bed (T max ) changed in the range 80 100 C while the minimum temperature of the bed (T min ) changed in the range 35 40 C. Furthermore, the inlet of the external fluids of the evaporator/condenser was settled at T ev-in =10 C and T con-in =35 40 C. These temperatures were constant during experiments, being fixed by means of thermocryostats, while the internal temperatures T ev and T con changed during the evaporation/condensation phase. In order to compare the experimental and calculated data in a proper way, it was necessary to determine the values of the parameters appearing in the above described equations and reported in Table 2. The overall heat transfer coefficient of the adsorber was calculated treating the measured bed temperatures, the flow rate and temperature of the heat transfer fluid. It was measured as U bed =8.5 W m 2 K 1, that is a typical value for a grain shaped sorbent bed [13]. In Figs. 5 7 the experimental results (T ev-in =10 C, T con-in =40 C, T max =95 C, T min =40 C) are compared to calculations. In Fig. 5, the calculated and experimental cycles are presented in the Clapeyron diagram. It can be seen that during the isobaric phases the pressure is indeed constant. It means that the size and Table 2 Input parameters for the model Description Symbol Unit Value Thermal capacity of bed HEX C ex JK 1 1564 Thermal capacity of condenser C con JK 1 2300 Thermal capacity of evaporator C ev JK 1 1380 Mass flow, adsorber inlet m : f kg s 1 0.3 Mass flow, condenser inlet m : f kg s 1 0.06 Mass flow, evaporator inlet m : f kg s 1 0.06 Adsorber total heat transfer U bed Wm 2 K 1 8.5 coefficient Condenser total heat transfer U con Wm 2 K 1 250 coefficient Evaporator total heat transfer U ev Wm 2 K 1 250 coefficient Adsorber exchanging surface A bed m 2 0.4 Condenser exchanging surface A con m 2 0.11 Evaporator exchanging surface A ev m 2 0.08

G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 289 Fig. 5. Comparison between the experimental and calculated cycles. Fig. 6. Behaviour of temperature and pressure vs. time during a cycle operated at T con-in =40 C and T ev-in =10 C. the heat transfer rate in both the condenser and the evaporator are quite sufficient to follow the cycle. This also confirms that the mass transfer resistance can be neglected and the limiting dynamic factor for this configuration of the bed is the heat transfer. The isosteric heating also follows the proper isosteric curve. The deviation of the experimental cycle from the calculated one during the isosteric cooling phase probably is due to the thermal inertia of the system. This effect is less evident during the isosteric heating, because the thermal power of the external heating device is higher than that used for the cooling phase. Fig. 6 shows the dynamic behaviour of the temperature and pressure during one complete cycle, here the experimental bed temperature refers to the average temperature of the bed. A good qualitative and quantitative agreement between the model and the experiment is observed. The total cycle time of the specific experiment was about 160 min. The isosteric phases were fast, and the slowest phase is the isobaric heating that tooks about 75 min, probably, due to the poor heat transfer inside the sorbent bed. As a consequence the thermal power exchanged during this phase was rather low (about 80 W), as shown in Fig. 7 in which the heat transfer rate as a function of time is presented for the four phases of the cycle. Also, in this case there is a good agreement between the model and the experiment. The curve corresponding to the heating/cooling power is characterised by a strong initial peak and by a continuous decreasing since the difference of temperature adsorbent/external fluid diminishes. In the same graph the curve of the power supplied to the evaporator is also presented; its behaviour is qualitatively in agreement with the expected results.

290 G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 Fig. 7. Experimental and calculated thermal power vs. time at T con-in =40 C, T ev-in =10 C and T max =95 C. It is interesting to mention that during isosteric phases the power supplied to (released from) the bed was much larger than during the isobaric phases and could reach 200 250 W. One of the possible reason of this enhancement is that in the closed volume adsorber in addition to the common heat conduction process through the adsorbent bed, the so called heat pipe effect can also contribute to the heat transfer [14,15]. This effect is caused by the migration of the vapour from the hot zones of the bed (where the heat is extracted) to the colder ones (where the heat is released). This successive desorption and adsorption can dramatically increase the heat transport inside the sorbent layer. The coefficient of performance COP was calculated from the experimental data by means of Eq. (1) for two different temperature of condensation and then compared with that calculated using Eq. (14). The values presented in Fig. 8 as a function of the desorption temperature, are found to be sensitive to both parameters and the behaviour of the curves is in agreement with the theoretical one; the typical experimental error in the various tests was about 5%. The COP gradually increases with the rise of the desorption temperature until it reaches 90 95 C, so that the material can be used for adsorption cooling driven by low temperature heat (heat wastes, solar energy, automotive exhausted gases). At T cond =35 C the COP can be as large as 0.6, that is a very promising starting point for the future development, especially considering that: (a) this is a single bed system without any heat recovery; (b) it is a gross COP of the unit that is affected by heat losses, the heat capacity of the inert masses and the heat exchangers efficiency which are not optimised yet in the unit tested. Nevertheless, the experimental COP is high enough mainly because of the prominent properties of the sorbent SWS-1L and, first of all, due to the large amount of water exchanged between the evaporator and the condenser during the cycle (w 14 wt.% at T cond =35 C). Evaporation of this amount of water gives a large cooling energy per cycle of Q ev =280 kj kg 1. At higher T con the value of w is lower that results in the COP decrease (Fig. 8). Regarding the useful instantaneous cooling power, the values measured in the evaporator during the isobaric adsorption stage ranged between 35 and 60 W; the corresponding mean specific cooling power calculated for the whole cycle is about 20 W kg 1 that is typical for chillers with a pelletised adsorbent [15,16]. The specific power can be increased changing the adsorbent shape Fig. 8. Experimental and numerical COP vs. desorption temperature at T con-in =35 C and 40 C, T ev-in =10 C.

G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 291 and the heat exchanger configuration (e.g. coated HEX or consolidated sorbent layers). The differences between the numerical simulation and the experimental data presented in the previous Figs. 6 8, are probably due to the assumption that some parameters are temperature and pressure independent (such as density and viscosity of the heating/cooling fluid, convective coefficients, etc.) and to the use of a simplified heat exchange description. Regarding the former consideration, it has been experimentally demonstrated that the thermal conductivity of the bed is not constant along the cycle but increases with the uptake rise [17]. Also, as already shown by Pons et al. [18], a proper account of the real properties of the heat transfer fluids could yield more reliable prediction of the results. Furthermore, as predicted by a previous detailed model [19] and experimentally found in this work, there are axial and radial thermal gradients inside the bed. Nevertheless, being the realisation of a more detailed model out of the aim of this work, the difference between experimental results and simulations can be considered acceptable. Finally, a parametric analysis was performed in order to estimate the possible increase of the specific power obtainable designing a more efficient system. In Fig. 9 the specific cooling power P s and the cooling COP are reported as a function of the bed heat exchanger efficiency e bed at various values of the evaporator efficiency e ev. The operating conditions chosen for the simulation are the following: T max =95 C, T min =35 C, T evin=10 C and T con-in =35 C According to the data presented in Table 2 and to Eq. (9), the global efficiency of the bed was calculated as 0.0045, while the evaporator efficiency e ev is 0.07; the corresponding COP and P s are respectively 0.58 and 20 W kg 1. It is evident that maintaining the same efficiency of the evaporator and increasing the efficiency of the bed, the COP decreases sharply while the increase in specific power P s is almost negligible. From the behaviour of the curve of the experimental device used (e ev =0.07) it is evident that this configuration is dynamically efficient for a maximum value e bed of 0.008. From the analysis of the figure it is clear that increasing both the efficiency of the bed and of the evaporator, an increase in the specific cooling power of one order of magnitude could be reached. In fact, if the sorbent bed heat transfer efficiency increases, the evaporating flow rate should increase as well, so that the evaporator efficiency must be enhanced. According to Eq. (9), the increase of the bed heat transfer efficiency e bed can be achieved by acting on the flow rate m : f of the external fluid used for transferring heat to the sorbent bed, on the heat exchange surface A bed or on the global heat transfer coefficient of the bed U bed. This result confirms that further efforts of the research should be addressed to the development of consolidated or coated beds with good heat and mass transfer properties. Another very important fact that was revealed in these experiments is the stability of the adsorbent material. In fact, after 60 complete cycles, like that presented in Fig. 5, the SWS-1L properties remained unchanged: no sorbent destruction and dust formation was found, and no traces of calcium chloride was detected in the adsorbed/desorbed water. Furthermore, X-Ray analysis showed no change in the structure of the host matrix, Fig. 9. Specific power P s (continuous lines) and COP (dot lines) as a function of the bed efficiency e bed, for different evaporator efficiencies e ev. The operating conditions are: T max =95 C, T min =35 C, T ev-in =10 C and T con-in =35 C.

292 G. Restuccia et al. / International Journal of Refrigeration 27 (2004) 284 293 while the measurement of some equilibrium points by the thermal balance system confirmed the adsorption properties measured on the unused material. This does not completely prove the hydrothermal stability of the SWS-1L but gives good indication about it and confirms the SWS usability for practical application in adsorption air-conditioning technologies. This conclusion is valid at least for cycles which minimum uptake is higher than that corresponding to solid crystalline hydrate formation (11% for SWS-1L) while the maximum uptake (29% in our experiments) is far lower than the pore overloading threshold (that is about 50% for this SWS-1L). 5. Conclusions The experimental results on the lab-scale chilling module working with the composite sorbent SWS-1L (the mesoporous silica gel KSK impregnated with CaCl 2 ) were presented together with a theoretical simulation. It was shown that the use of this composite material allows to reach the cooling COP up to 0.6 at the low desorption temperature of 90 95 C. This is a gross COP that is affected by heat losses, the heat capacity of the inert masses and the heat exchangers efficiency which are not optimised yet in the unit tested. Of course, it can be further improved in the multi-bed system with internal heat recovery. So large COP shows the advantage of the new composite sorbent with respect to common commercial materials (zeolite or silica gel) at the same regeneration temperature. It is also shown that the low specific power of the tested device is due to the not optimised design of the adsorber heat exchanger as well as to the grain shape of the adsorbent. So that it can be considerably increased changing the adsorbent shape and the heat exchanger configuration. Finally, the predictive dynamic model of the unit operation was presented; despite the simplicity of the mathematical description of the cycle, good agreement between the model and the experimental measurements were found. The model was also used to study the influence of the main operating parameters on the system performance and to make recommendations on how to improve the chiller design and process parameters. Acknowledgements This work is supported in part by the NATO (Collaborative Linkage Grant PST.CLG.979051) and the Program of Scientific Co-operation between the CNR and the RAS. References [1] Ziegler F. State of the art in sorption heat pumping and cooling technologies. Int J Refrigeration 2002;25(4):450 9. [2] Meunier F. Sorption contribution to climate change mitigation. In: Proceedings of international sorption heat pump conference, 24 27 September 2002, Shanghai, P.R. China. p. 1 9. [3] Cacciola G, Restuccia G, Aristov YuI, Tokarev MM. 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