desorption times: Experimental operation of a full-scale 3 beds adsorption chiller

Applied Energy 205 (2017) 1081–1090

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Applied Energy journal homepage: www.elsevier.com/locate/apenergy

A new management strategy based on the reallocation of ads-/desorption times: Experimental operation of a full-scale 3 beds adsorption chiller

MARK

⁎

Alessio Sapienzaa, , Valeria Palombaa,b, Giuseppe Gullìb, Andrea Frazzicaa, Salvatore Vastaa a b

CNR-ITAE - Institute of Advanced Energy Technologies “Nicola Giordano”, Salita S. Lucia sopra Contesse 5, 98126 Messina, Italy Department of Engineering, University of Messina, Contrada di Dio, 98166, S. Agata, Messina, Italy

H I G H L I G H T S innovative 3-beds adsorption chiller is presented. • An proposed are the hybrid coated/granular adsorbers and the management strategy. • Innovations management is based on the reallocation of the ads-/desorption durations. • The wide experimental characterization was carried out. • AVolumetric cooling power up to 275 kW/m was measured. • 3 adsorber

A R T I C L E I N F O

A B S T R A C T

Keywords: Adsorptive chiller Hybrid adsorbers Reallocation of the ads-/desorption durations

In this paper, a wide experimental characterization campaign aimed at measuring the performance of an innovative 3-beds adsorptive chiller prototype is reported. The prototype was designed to employ a new management strategy, based on different durations of adsorption and desorption isobaric steps of a basic temperature driven adsorptive cooling cycle and aimed at achieving high cooling power density. The performance were measured, in terms of Average Cooling Power (ACP), cooling COP and Volumetric Cooling Power (VCP), using a test bench suitable for the characterization of thermally driven cooling/heating machines, under typical boundary conditions of air conditioning applications. A sensitive analysis was carried out to assess the effect of several parameters (e.g. the temperature lift TM − TL, driving temperature TH and cycle time). on the operation of the chiller. The results of the experimental activity depict a complete performance map for the tested chiller and demonstrate the potential of the new cycle management strategy. At nominal boundary conditions (i.e. TH ∼ 90 °C, TL ∼ 18 °C and TM ∼ 25 °C), the cooling machine was able to deliver an ACP of 4.4 kW and an overall VCP of 9.4 kW/m3, with a COP of 0.35 while the VCP referred to the volume of the only adsorber was ∼275 kW/m3.

1. Introduction Utilization of energy from renewable sources and the reutilization of waste thermal energy have gained major interest during the last years, with the goal of reducing both the share of traditional fossil energy sources consumption and the energy-related environment pollution. Indeed, growing concerns towards the environment are testified by international interest in containing the high GWP/ODP emissions and the consequent climate changes [1]. The influence of HVAC systems on the atmospheric pollution has already been proved to be relevant [2,3]. Possible strategies for the reduction of the emission of polluting gases due to heating and cooling ⁎

Corresponding author. E-mail address: [email protected] (A. Sapienza).

http://dx.doi.org/10.1016/j.apenergy.2017.08.036 Received 26 May 2017; Received in revised form 21 July 2017; Accepted 8 August 2017 0306-2619/ © 2017 Published by Elsevier Ltd.

sectors are the substitution of the currently employed refrigerants, characterized by high GWP and ODP [4] as well as the application of alternative technologies, such as HVAC systems driven by thermal energy (e.g. solar heat or waste heat), instead of traditional vapour compression ones driven by electricity. Among the innovative environmentally friendly cooling technologies, adsorption machines for air conditioning and refrigeration represent a viable and promising technology, since they can be driven by low-grade heat sources and employ clean refrigerants (e.g. water, ethanol) [5,6]. Nonetheless, still opened critical issues, mainly represented by high capital cost and low COP and power density, limit the large scale

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Nomenclature

T u V̇

Symbols ACP AdHEX COP cp GHG h H HEX HTF HVAC m ṁ λ ODP Q R s S τ

temperature (°C) uncertainty (–) volumetric flow rate of heat transfer fluid (LPM) volume (m3) volumetric cooling power (W/m3)

V VCP

average cooling power (W) adsorbent heat exchanger coefficient of performance (–) specific heat [kJ/(kg K)] greenhouse gas convective heat transfer coefficient [W/m2 K] enthalpy [J/kg] heat exchanger heat transfer fluid heating, ventilation and air conditioning mass, sorbent dry mass [kg] flow rate [kg/s] thermal conductivity [W/m K] ozone depletion potential thermal energy (J) adsorption/desorption duration ratio (s/s) thickness [m] heat transfer surface of HEx [m2] time (s, min)

Subscripts AdHEx Al Amb des ev in ins H L M Opt Out Ref Sens Ss

adsorbent heat exchanger aluminum ambient desorption evaporation inlet insulating material high low medium optimum outlet refrigerant sensible stainless steel

condenser) at different levels:

diffusion of such systems. In Table 1, some of the most recent adsorption chiller prototypes/commercial machines and developed adsorber concepts are listed and their main performance indicators are reported. Where possible, a distinction has been made between the volumetric cooling power calculated with respect to the overall volume of the chiller (VCPchiller) and the one calculated considering only the volume of the adsorber (VCPAdHex). All the prototypes and commercial systems shown in Table 1 employ the classical double-bed architecture with the adsorbers operating in counter-phase by a standard management strategy based on equal duration of adsorption and desorption steps. The prototypes presented in [7,9] have low VCPchiller (up to 3 kW/m3), while the commercial systems [10,11] and the prototype developed for automotive applications presented in [8] have VCPchiller exceeding 12 kW/m3. This is due to the optimised design of all the components of these systems to fit specific dimension requirements. Taking into account only the volume of the adsorber, the VCP is below 100 kW/m3 for all the listed full scale machines while for the lab scale adsorber concepts [13–15], only the direct synthesis of the adsorbent material on the HEx assures high value of power density (up to 320 kW/m3). Numerical studies of more complex machine configurations were discussed in [16,17] demonstrating the potentiality of such kind of machine layouts. To improve the dynamic performance of the adsorptive technology and consequently the cooling power density, the research efforts are mainly addressed towards an increase of the efficiency of the main components (i.e. adsorber and evaporator/

(i) at the adsorbent material level, the research focuses on the development of new materials or adsorbent configurations [13–20] as well as the proper selection/tailoring of the optimum adsorbent for the specific application and/or boundary conditions, (ii) at components level, the development of efficient evaporator/ condenser is the key issue: especially when water is used as refrigerant, the low heat transfer coefficients and the still partial knowledge of the phase change phenomena limit the proper design of these components [21,22], (iii) finally, optimization of the coupled system (adsorbent material plus heat exchanger) is mandatory to get high VCP [23–25]. A different approach to improve the dynamic performance of an adsorption machine has been recently proposed in [26] and further investigated in [27] by testing small scale adsorbers through a dedicated test bench. This approach aims at getting high power density by a proper reallocation of the ratio between the adsorption and desorption steps duration according to the nature of the working pair and the operating conditions. Results reported in [26,27] showed that, on the basis of the boundary conditions, adsorption and desorption can have different sorption kinetics and that a proper management strategy optimization allows a remarkable increasing in overall performance. On the basis of the previously mentioned results, the authors developed a full scale adsorption chiller prototype based on an innovative “3 hybrid -adsorbers architecture” able to operate with an adsorption-

Table 1 Analysis of different adsorption air-conditioning prototypes and commercial machines. Reference [7] [8] [9] [10] [11] [12] [13] [14] [15]

Working pair Silica gel/Water AQSOA FAM Z02 (loose grains)/Water LiCl-Silica gel/methanol Silica gel/water Silica gel/water Silica gel/water SAPO 34 (coating by direct synthesis)/Water SAPO 34 (coating by dip coating)/Water SAPO 34 (coating by direct synthesis)/Water

No. adsorbers 2 2 2 2 2 2 1 1 1

Cooling capacity 5.7 kW 2.3 kW 4.9 kW 14 kW 16 kW 15 kW – 1 kW 5.6 kW

1082

COP 0.4 0.3 0.4 0.6 0.6 0.5 0.6 0.24 0.4

VCPchiller

VCPAdHex 3

3.06 kW/m 13.50 kW/m3 2.65 kW/m3 12.3 kW/m3 15.4 kW/m3 – – –

92 kW/m3

20.6 kW/m3 150 kW/m3 93 kW/m3 320 kW/m3

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to-desorption duration ratio of 2 maintaining a uniform cooling production. In [28] the machine design, manufacturing and the results of a preliminary test campaign were presented. The chiller presented is the first example, in literature, of such a configuration, since up to now only theoretical analysis or proof-of concepts in small scale were taken into consideration. This paper represents a prosecution of the activity resumed in [28] and deals with the complete experimental characterization of the prototype, carried out by using a test bench for the characterization of thermally driven systems installed at CNR-ITAE. A performance map was outlined in terms of COP, ACP and VCP (both related to the whole machine and only to the adsorbers’ volume) while a sensitive analysis was carried out, taking into account several boundary conditions and/ or operation modes (i.e. influence of the temperature lift TM − TL, influence of driving temperature TH and cycle time). The testing conditions chosen correspond to the real operation of the system. Moreover, all the investigated parameters were varied in a very wide range, thus covering different possible fields of applications, both for mobile or stationary use. Furthermore, a design analysis and a study of the performance improvement options were performed, to estimate the effect of the inert masses and the feasibility of a heat recovery mode.

Table 2 Main features of the adsorption chiller.

2. Adsorptive chiller description

3.2. Testing procedure and performance evaluation

The full scale prototype, shown in Fig. 1, was designed and realized at CNR-ITAE with the main aim of achieving high cooling power density [28]. To get this objective, two main innovative features were implemented in the design phase:

The experimental activity was realised in accordance to the test protocol under validation for the assessment of adsorption chillers performance, developed within International Energy Agency Task 48 activity [29]. The parameters employed for characterization of the chiller were calculated as follows: Cooling COP:

Dimensions [mm] Volume [m3]

860 × 790 × 690 0.47

Weight [kg]

270

Nominal COP

0.35

Adsorbent material N°1: coating Adsorbent material N°2: grains Nominal Average Cooling Power [kW] Nominal Volumetric Cooling Power [kW/m3]

Mitsubishi AQSOA FAM Z02 Microporous Silica Gel 4.4 9.36

3. Experimental activity 3.1. Test bench for thermally driven chillers The experimental activity was carried out at CNR-ITAE by a test bench specifically developed to measure the performance of thermally driven heat pumps/chillers. It allows interfacing the adsorption prototype with the external heat sources/sinks and simulating a cooling load, thus reproducing the real operating conditions of thermally driven chillers (e.g. when installed in an air conditioning system). A detailed description of the testing rig is given in [28].

(i) a new machine architecture, employing 3 adsorbers connected to a single evaporator and condenser, that allows performing an advanced machine’s management strategy by means of unbalanced durations of the isobaric ads-/desorption steps, (ii) hybrid adsorbers, realized embedding microporous Silica Gel loose grains into aluminum finned flat tube heat exchangers, previously coated with the Mitsubishi AQSOA FAM Z02 sorbent.

τ

COP =

∫0 cycle ṁ ev ·cp·(Tin,ev−Tout,ev )·dτ ∫0

τcycle

ṁ des-adhex ·cp (Tin,des-bed−Tout ,des-bed )·dτ

(1)

Average Cooling Power (ACP): τ

ACP = Indeed, according to results reported in [26,27], three separated adsorbers are needed to be able to operate the chiller with R = 2, as the reallocation of ad-/desorption durations causes subsequent change in cooling cycle organization. This operation mode ensures a continuous cold production since, at any time, two adsorbers are connected with the evaporator while the third one is being regenerated against the condenser. The detailed description of the cooling machine is reported in [28] while in Table 2 its main features are summarized.

∫0 cycle ṁ ev ·cp·(Tin,ev−Tout,ev )·dτ τcycle

(2)

Volumetric Cooling Power (VCP), both referred to the adsorbers’ volume (AdHEx) and to the overall chiller volume (chiller): τ

VCPAdHEx =

∫0 cycle ṁ ev ·cp·(Tin,ev−Tout ,ev )·dτ VAdHEx ·τcycle

(3)

τ

VCPchiller =

∫0 cycle ṁ ev ·cp·(Tin,ev−Tout ,ev )·dτ Vchiller ·τcycle Fig. 1. Views of the 3 beds adsorptive chiller prototype.

1083

(4)

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deliver about 223 kW/m3 of cooling power at Te = 15 °C, Tc = 35 °C, Th = 90 °C that is quite close to results achieved for the 3 beds full scale adsorption chiller and even higher than results achieved employing a ratio R equal to 1 [27].

3.3. Testing conditions The test campaign was planned as to get a complete and exhaustive performance map of the studied cooling machine. In particular, a sensitive analysis aimed at assessing the influence of the following operating conditions and design parameters was carried out:

4.2. Influence of the temperature lift between the condenser and the evaporator

• cycle time; • temperature lift between the condenser and the evaporator (T − T ); • driving temperature T • Influence of heat transfer fluid (HTF) flow rate in the evaporator. M

The influence of the temperature lift between the condenser and the evaporator on the chiller performance was studied for a wide range of delta T (7–30 °C) keeping constant the driving temperature (TH = 90 °C), as reported in Table 3. In Fig. 4, the cycle time at which the tests were performed. was selected as the one maximizing the ACP, considering an application where high power density is required, while in Fig. 5 it was considered equal as the one needed to optimize the COP for applications asking for high efficiency. Figs. 4 and 5 show a quasi-linear correlation between both ACP and COP and the temperature lift. This is probably due to a quasi-proportional variation of amount of refrigerant processed at quasi constant cycle time while varying the T lift. This is a consequence of the adsorption equilibrium properties of the used sorbents (Silica Gel and Mitsubishi FAM Z02) at the specific tested boundary conditions [31–33]. The maximum ACP and COP decrease from 4.4 to 0.8 kW and from 0.35 to 0.11 respectively when the delta T between condenser and evaporator varies from 7 to 30 °C.

L

H;

Table 3 summarises all the operating conditions investigated. The testing conditions are reported in terms of low temperature (TL), medium temperature (TM) and high temperature (TH) that represent the 3 temperature levels at which the adsorption cycle operates. In particular, they refer to the average inlet temperature of the HTF flowing into each component of the chiller. The ratio (R) between the duration of adsorption and desorption phases was kept constant and equal to 2 for all the tests. Uncertainty analysis estimation was performed according to the extended uncertainty theory [30] and is reported in [28] for the experimental data obtained, with maximum values of 6% for the COP, VCP and ACP. 4. Results and discussion

4.3. Influence of the desorption temperature (TH) 4.1. Influence of the cycle time The effect of the desorption temperature on the ACP and COP was studied for 3 different boundary conditions (i.e. TM=30 °C, TL=10 °C; TM=28 °C, TL=15 °C, TM=25 °C, TL=18 °C), 4 driving temperatures (TH=60/70/80/90 °C) and several cycle times (from 2 to 40 min). The obtained results are shown in Figs. 6–8. For all boundary conditions the cooling power capacity (ACP) is strongly affected by the reduction of the desorption temperature probably due to the reduction of the heat transfer driving force which limits the kinetic of the process. At TM=30 °C and TL=10 °C, the ACP drops from 2.64 to 1.1 kW (−58%) reducing TH from 90 °C to 70 °C,

The influence of the cycle time on the chiller performance was studied for a wide range of evaporation (7–18 °C) and condensation temperatures (25–40 °C) keeping constant the driving temperature (TH = 90 °C) and the flow rates in the hydraulic circuits (see Table 3). For each boundary condition, 8/9 different cycle times were investigated, ranging from 2 to 40 min, to outline a performance map which covers both applications where high cooling power density is required as well as applications where the overall efficiency needs to be optimized. Figs. 2 and 3 depict the performance in terms of ACP and cooling COP as a function of cycle time. For all the tested operating conditions, the ACP shows a maximum between 5 and 8 min of cycle time: when the cycle time is lower, the performance decreases because the adsorption process is still not completed, while at higher times the ad/desorption processes becomes slower, negatively affecting the dynamic performance. As a consequence, the cooling power delivered is strongly reduced due to the useless protracting of the cycle. Fig. 3 shows the cooling COP versus the cycle time for the tested conditions. For all the boundary conditions, the COP presents a maximum between 20 and 30 min of cycle time: at lower cycle times, the performance decreases because the adsorption process is not completed while the energy losses due to the presence of inert masses negatively affect the performance. Outcome of this first analysis, as reported in Figs. 2 and 3, was the definition of the optimal cycle time according to the specific application (i.e. maximization of cooling power or COP) and the boundary conditions. At optimum cycle time, the volumetric cooling power for the chiller (VCPchiller), ranges from 1.76 to 9.36 kW/m3, while the VCPAdHEx ranges from 51.87 to 275.62 kW/m3. Interestingly, VCPAdHEX achieved for the full scale 3 beds chiller is in good agreement with the previous results obtained for a small scale AdHEx with similar design, as reported in [27], where the idea to reallocate the time management strategy was firstly proposed and verified. Indeed, in [27], a small scale AdHEx, using FAM Z02 as sorbent, showed an optimal adsorption to desorption time ratio of 2.5. Under this operation mode, it was able to

Table 3 Investigated operating conditions. PARAMETER 1: influence of the cycle time Driving temperature [°C] Condenser inlet temperature [°C] Evaporator inlet temperature [°C] Volumetric Flow rate adsorber heating/cooling [LPM] Volumetric Flow rate condenser [LPM] Volumetric Flow rate evaporator [LPM] Cycle time [min] PARAMETER 2: influence of the temperature lift Driving temperature [°C] Temperature lift [°C] Cycle time [min]

90 25/28/30/35/40 7/10/15/18 15.3/26 26 21 2/3/5/7/8/10/15/20/25// 30/40 90 7/10/13/15/18/20/25/28/ 30 Optimal cycle time

PARAMETER 3: Influence of the desorption temperature Driving temperature [°C] 60/70/80/90 Condenser inlet temperature [°C] 25/28/30 Evaporator inlet temperature [°C] 10/15/18 Cycle time [min] 2/3/5/7/10/15/20/30/40 PARAMETER 4: influence of heat transfer fluid flow rate in the evaporator Driving temperature [°C] 90 Temperature lift [°C] 24–18/26–15/28–15/30–10/ 35–7 Cycle time [min] Optimal cycle time Volumetric Flow rate evaporator [LPM] from 7 to 66

1084

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Fig. 2. Average cooling power versus cycle time at several boundary operating conditions and fixed TH = 90 °C.

5.0

Average cooling power [kW]

4.5 4.0 90-35-15

3.5

90-28-15 3.0

90-25-10 90-25-15

2.5

90-25-18

2.0

90-28-10

1.5

90-35-10

1.0

90-40-10 90-35-7

0.5 0.0 0

10

20

30

40

LJcle Ɵŵe [ŵin] 4.4. Influence of heat transfer fluid flow rate in the evaporator

while at TM=28 °C, TL=15 °C and TM=25 °C, TL=18 °C the reduction is respectively from 3.43 to 1.83 kW (−46%) and 4.41 to 2.72 kW (−38%). This shows that the reduction of ACP is more evident for a higher temperature lift values (between the condenser and the evaporator). Concerning the COP, at higher temperature lifts (Figs. 6 and 7), the performance achieved at TH of 80 °C and 90 °C are very similar, while a reduction of COP is observed at a lower desorption temperature (i.e. 70 °C). Indeed, the desorption temperature affects the ACP more evidently with respect to the COP: this is probably due to the specific layout of the adsorbers with a hybrid configuration based on two different sorbents. Zeolite provides the peak power, and, since it requires higher temperatures to be regenerated, the power decreases significantly going from 90 °C to 70 °C. On the contrary, Silica Gel provides a continuous, despite lower, cooling effect, but can be efficiently regenerated at 80 °C [31], thus limiting the reduction of the COP with decreasing desorption temperature within a narrow range. It is interesting to notice that, for lower condensation temperatures and higher evaporation temperatures, (TM=25 °C, TL=18 °C), even at a lower desorption temperature (60 °C) the chiller is able to work with high efficiency, reaching the best performance in terms of COP due to the reduction of the entropy generation [21] as well as the reduction of the heat losses to the surroundings.

In order to assess the influence of the heat transfer fluid flow rate trough the evaporator, a specific experimental campaign was been carried out by varying the boundary conditions, as listed in Table 3. The desorption temperature was kept constant at 90 °C, the cycle time was chosen as the one maximizing the ACP and the volumetric flow rate of the heat transfer fluid ranged from 7 to 66 LPM. Figs. 9 and 10 shows the performance, in terms of ACP and COP as a function of the flow rate. The tests were carried partially floating the evaporator with the aim of covering one of the two HEXs (half of the evaporator height) according to the layout described in [28]. For most of the boundary conditions, a reasonable increasing in performance is observed in the flow rate range 7–22 LPM while at higher values the improvement is less evident. Probably in the range of 7–22 LPM the internal convective heat transfer resistance affects strongly the evaporation process while at higher values another heat transfer mechanism limits the process.

5. Design analysis and performance improvement options Finally, a brief design analysis and a study of the performance improvement options were executed to estimate the effect of the inert masses and the feasibility of a heat recovery mode. Fig. 3. COP versus cycle time at several boundary operating conditions and fixed TH = 90 °C.

0.40 0.35

90-35-15 90-28-15

0.30

90-25-10

COP

0.25

90-25-15 0.20

90-25-18 90-28-10

0.15

90-35-10

0.10

90-40-10 0.05

90-35-7

0.00 0

10

20

30

40

Cycle Ɵŵe [ŵin] 1085

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Fig. 4. Maximum Average Cooling Power (at optimal cycle time) versus real temperature lift (TM − TL) at fixed TH = 90 °C.

5.0 4.5 4.0

ACP [kW]

3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0 0

5

10

15

20

25

30

35

Temperature liŌ [°C] 25-18

25-15

28-15

25-10

28-10

35-15

35-10

35-7

40-10

Fig. 5. Maximum COP (at optimal cycle time) versus real temperature lift (TM − TL) at fixed TH = 90 °C.

0.40 0.35 0.30

COP

0.25 0.20 0.15 0.10 0.05 0.00 0

5

10

15

20

25

30

35

Temperature liŌ [°C] 25-15

28-15

25-10

28-10

35-15

35-10

35-7

40-10

Consequently, vacuum chambers with flanged lids have been realised, suitable for the inspection of the adsorbers or the replacement of components. However, this is realised at the expenses of a heavy and bulky system, resulting in an overall weight of 270 kg. The effect of

5.1. Effect of inert masses

ACP [kW]

The prototype here presented was entirely realised in stainless steel, the lead principle in the design being the flexibility for testing purposes.

3.0

0.30

2.5

0.25

2.0

0.20

1.5

0.15

1.0

0.10

0.5

0.05 0.00

0.0 0

10

20

30

40

Cycle time [min] COP

90°C

80°C

70°C

ACP

1086

Fig. 6. Influence of heating temperature (TH) on the performance at TM = 30 °C, TL = 10 °C.

COP

25-18

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4.0

0.35

3.5

0.30 0.25

2.5

0.20

2.0 0.15

1.5

COP

ACP [kW]

3.0

Fig. 7. Influence of heating temperature (TH) on the performance at TM = 28 °C, TL = 15 °C.

0.10

1.0

0.05

0.5 0.0

0.00 0

10

20

30

40

Cycle Ɵme [min] COP

70°C

90°C

80°C

ACP

manifolds. Texchanger,max and Texchanger,min are the maximum and minimum temperatures of the heat exchangers during a cycle, considered equal to the temperature of the material inside the exchanger, which has been measured by means of a thermocouple.

such inert masses on the performance of the chiller, especially in terms of COP, can be relevant, since a significant amount of energy is cyclically wasted in heating and cooling the shells of the components. To identify and quantify this effect, an energy balance was performed for a “90-30-10 °C” test with a cycle time of 15 min, and all the contributions to the total heating energy to be supplied were calculated. In particular, the contributions listed below were considered.

– Sensible heat needed to heat up the adsorbent materials (zeolite and silica gel), calculated as:

Qsens,adsorbents = (mzeolite c pzeolite + msilicagel c psilicagel)

– Sensible heat needed to heat up the material of the chambers, calculated as:

Qsens,chambers = m chambers c pSS (Tchamber,max−Tchamber,min)

(Tadsorbent,max−Tadsorbent,min)

(5)

where mchambers is the total weight of the chambers, cpSS is the specific heat of the stainless steel, Tchamber,max and Tchamber,min are the maximum and minimum temperatures of the external wall of the chambers during a cycle, experimentally measured through a thermocouple. – Sensible heat needed to heat up the metal heat exchangers and the manifolds of the adsorbers, calculated as:

c padsorbent = c pdry + c pref Δw

(6)

where mexchangers is the total weight of the heat exchangers, mmanifold is the weight of the manifolds, cpAl is the specific heat of the aluminum of the heat exchangers, cpbrass is the specific heat of the brass of the 0.40

4.5

0.35

Fig. 8. Influence of heating temperature (TH) on the performance at TM = 25 °C, TL = 18 °C.

3.0

0.25

2.5

0.20

2.0

0.15

1.5

0.10

1.0

0.05

0.5 0.0

0.00 0

10

20

30

40

Cycle Ɵme[min] 90°C

80°C

70°C

COP

60°C

ACP

1087

COP

0.30

3.5

ACP [kW]

– Enthalpy of desorption, calculated as:

5.0

4.0

(8)

where cpdry is the specific heat of the dry material, taken from [34], cpref is the specific heat of the refrigerant (in this case, water), Δw is the uptake of refrigerant processed during the cycle, that has been calculated according to the isosteric diagram of the two materials employed.

Qsens,exchangers = (m exchangers c pAl + mmanifold c p brass) (Texchanger,max−Texchanger,min)

(7)

where mzeolite is the total mass of zeolite inside the adsorbers, msilicagel is the mass of the silica gel inside the adsorbers, Tadsorbent,max and Tadsorbent,min are the maximum and minimum temperatures of the adsorbent material during a cycle, experimentally measured through a thermocouple. The specific heat of the adsorbents has been calculated, in both cases, as [34]:

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Fig. 9. ACP versus flow rate of the heat transfer fluid in the evaporator at different boundary conditions (at optimal cycle time for ACP).

5.5 5.0 4.5

AC P [kW]

4.0 3.5 3.0 2.5 2.0 1.5 1.0 0.5 0.0 0

10

20

30

40

50

60

70

Volumetric Ňow rate [LPM] 90-24-18

90-26-15

90-28-15

90-30-10

90-35-7

ΔHdes = Hzeolite Δwzeolite + Hsilicagel Δwsilicagel

(9)

materials, a increment in the COP could be expected.

where the terms Hzeolite and Hsilicagel have been measured at ITAE [35]. 5.2. Evaluation of heat recovery

– Heat losses through the environment, calculated as:

Heat and mass recovery processes represent a design option for thermally driven cooling systems [37–40]. The feasibility of heat recovery has been evaluated. As explained in [28], the management of the prototype is quite different from the standard cycle of two-adsorbers systems and the cycle has been rearranged to operate the system properly. The estimation of the recoverable heat was done under the following assumptions:

−1

1 s Q losses = ⎛ + ins ⎞ S(Twall−Tamb) h λ air ins ⎠ ⎝ ⎜

⎟

(10)

where hair is the convective heat transfer coefficient for natural convection in air, taken from [36], sins is the thickness of the insulating material, λins is the thermal conductivity of the insulating material, S is the area exposed to the environment, Twall is the temperature measured at the wall of the chamber, and Tamb is the ambient temperature during the test. The results of the study are reported in Fig. 11: as visible, only 66% of the total heat supplied is due to the adsorbent materials (for sensible heating and desorption), the remaining 34% being almost equally divided among the heat losses and the heat wasted for the heating of the metal masses inside the system. In this context, the use of different materials for the chambers could be beneficial. Considering, for example, plastic materials suitable for high temperatures (e.g. nylon, PPA, PPS), their specific heat is of the same order of magnitude of stainless steel (0.4–0.7 kJ/kg K), while their densities are between 1.4 kg/dm3 and 2.4 kg/dm3, 7–3 times lower than stainless steel. In addition, their thermal conductivity is intrinsically lower than those of metals, thus reducing the need for high-thickness insulation. By using these

• Heat is recovered during the isosteric heating/cooling phases; • the duration of isosteric heating/cooling is fixed and equal to 20 s, •

this value being defined as the optimal one after a dedicated experimental campaign; the heat recoverable is the sensible heat due to the heating/cooling of the heat exchangers containing the adsorbent and the amount of the heat transfer fluid (water) in the pipes.

For a “90-30-10 °C” cycle, the heat recoverable under these conditions amounts to 200 kJ, representing only 7% of the overall heat supplied to the system under the examined boundaries, and therefore the effect on the COP would not be significantly relevant; as previously explained, by changing the design of the system or its assembly, greater Fig. 10. COP versus flow rate of the heat transfer fluid in the evaporator at different boundary conditions (at optimal cycle time for ACP and refrigerant level 0.5).

0.30

COP

0.25

0.20

0.15

0.10 0

10

20

30

40

50

60

70

Volumtric Ňow rate [LPM] 90-24-18

90-26-15

90-28-15

90-30-10

90-35-7

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Fig. 11. Shares of contributions to the total heat supplied to the chiller.

improvements can be achieved. [3]

6. Conclusions [4]

The main aim of this paper was to study a new management strategy for adsorption chillers based on a reallocated duration of adsorption and desorption steps of a basic adsorptive cooling cycle. The evaluation was experimentally carried out by the use of an innovative adsorption chiller prototype specifically designed to operate with adsorption to desorption ratio equal to two. The machine is based on three adsorbers, connected to a single evaporator and condenser, realized with a hybrid adsorbent reactor layout employing a coating of Mitsubishi AQSOA FAM Z02 and loose grains of a commercial Silica Gel as sorbents. The paper presented a wide experimental characterization campaign aimed at drawing an exhaustive performance map and at demonstrating the improvement achieved by the new management strategy. The results showed that the prototype is able to deliver high cooling power density. The volumetric cooling power (VCP), at optimum cycle time, ranged from 1.76 to 9.36 kW/m3 and from 51.87 to 275.62 kW/ m3 referred, respectively, to the volume of the entire chiller or to the volume of the only adsober. The here presented outcomes confirm, at full scale chiller level, the previous studies on the reallocation of the ads-/desorption steps carried out on reduced size systems.

[5] [6]

[7]

[8] [9]

[10] [11] [12]

[13]

[14] [15]

Acknowledgments The present work was partially funded by PON “Ricerca e Competitività 2007-13” n°02_00153_2939517, in the frame of the project T.E.S.E.O. (Efficient Technologies for Energy Sustainability and Environmental On Board) – P.O.N. (Research and Competition 2007/ 2013) and by PON “Ricerca e Competitività 2007-13” PON03PE_00206_2 S5 - Smart Small Scale Solar Systems.

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[19]

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