First level

The first level of the procedure permits to acquire basic information on cost and performance (on primary energy level) of the system with a limited number of sensors. Within this level 4 heat flow meters and the electricity counters for the measurement of the electricity consumption of the overall system is required. In Figure 3 the scheme with the measured energy fluxes is shown.

Figure 3. 1st level monitoring scheme including measured inputs and outputs of the solar assisted cooling

system

The primary energy ratio of the solar assisted cooling system can be calculated as shown in (Equation 1:

Where the heat and electricity fluxes are measured while the primary energy conversion factors for heat and electricity from fossil fuels have been set with the following values, being based on European Directives [4], Task 25 and Task 32 [7]:

• seiec = 0.4 (kWh of electricity per kWh of primary energy)

• sfossil = 0.9 (kWh of heat per kWh of primary energy)

• nboiler = 0.95 (boiler efficiency)

In general it has to been stated that the conversion and performance factors given in the present paper are based on literature and discussion agreements but have to be considered only as proposals for calculation and comparison as they depend on different aspects such us country specifications, technology and component size. The Primary Energy Ratio of a reference system (see Figure 2) can be calculated as shown in (Equation 2:

In this equation the heat and electricity fluxes are again the measured values. Within the calculation of the electricity consumption of the reference system the consumption of the pumps of solar circuit loop and of the absorption chiller have to be subtracted according to Figure 2 and as shown in (Equation 3.

Eelec, tot _ ref = Eelec, tot — (E1 + E2 + E6 + E7 + E8 + E10 + E11 + E14 + E18 + E1q)

The electricity consumption of the auxiliary E3 has to be corrected as described with (Equation 15, since the auxiliary in the reference has to deliver much more heat. The primary energy conversion factors for heat and electricity from fossil fuels have been listed before, the Seasonal Performance Factor (SPF) of the reference compression chiller has been set to:

• SPFref = 2.8 (compression chiller efficiency of the reference system)

• nboiler, ref = 0.95 (reference boiler efficiency)

From the financial point of view the overall cost per installed cooling capacity can be calculated following

Cost (€) includes the costs of all components shown in Figure 1 minus the appliances deployed in the corresponding conventional reference system (Figure 2) such us back up heating / cooling system and eventually installed cogeneration systems.

2.1. Second level

The second step consists within deeper partial monitoring of single parts of the system with an increased number of sensors in respect to the 1st level. In fact within this level 2 heat counters and a pyranometer have to be added to the 4 heat counters and the total electricity counters already present in the 1st level. The monitoring for this level is concentrating mainly on the solar thermal energy management (see Figure 4).

In the following several equations are shown which allow to calculate the amount of solar energy which the solar assisted cooling system is not able to exploit because of different losses such us within the solar collectors, heat exchangers or the storage.

The amount of solar energy that is not exploited because of losses in the solar loop, heat exchangers and collectors is calculated by the (Equation 5 — (Equation 6:

 Vcott. net _ Q Qsol (Equation 5) ^Qsol _ Qsol (1 — Vcoll ) (Equation 6) The amount of thermal energy that is not exploited because of losses in the storage the (Equation 7 — (Equation 8: is calculated by n _ Q3 + Q6 + Q4 иГа*Є Q1+Q2 (Equation 7) Qloss, storage _ (Q1 + Q2) — (Q3 + Q6 + Q4) (Equation 8) Finally, the amount of available solar energy that is not used in the thermal loop of the solar assisted cooling system (Qsolar, unex) is calculated by the (Equation 9 — (Equation 13: SF _ Q (Q, +Q2) (Equation 9) Q* _ SF • Q6 Q3* _… (Equation 10) Q* _ Q* + Q* + Q* (Equation 11)

 Qo — Q*

 (Equation 12)

 (Equation 13)

 Qsolar, unex

2.2. Third level

The third level consists in a full system monitoring based on the method of “fractional energy saving” as part of the FSC method which was elaborated in the IEA SHC Task 26 for solar combi systems and extended in the IEA SHC Task 32 for solar heating and cooling systems [7]. This file outlines the energy-flux components required to characterize the performance of solar heating and cooling systems with this method. The FSC method as such is not explained in this document, readers interested in a tutorial on the FSC method are referred to the cited literature [8] [9] [10].

 (Equation 14)

 sav, shc

(Equation 14 defines the “fractional solar heating & cooling savings” (f savshc) in terms of:

• energy consumption attributed to auxiliary devices required for the solar heating/cooling system. (numerator)

• energy usage allocated to a reference system with no solar energy-input (denominator)

The crux of the mater is to come up with a practical definition as to define the reference system, having in mind that the only accessible measurement object is the building with the SHC system. The proposed strategy is to derive the electricity consumption of the reference system “Eel, ref” from the SHC system, by adequate modifications in the measurement data analysis, following the below outlined scheme:

• define a maximal equipped SHC system (see Figure 1, labelled “SHC_max system”)

• the appliances deployed in the corresponding conventional reference system are depicted graphically in Figure 2 by skipping all solar assisted equipment from the SHC_max system.

• All thermal and electrical energies of the conventional reference system can be determined from measurements conducted in the SHC_max system using the following assumptions.

• All thermal energies (hot/cold) supplied by solar in the SHC-system are fully substituted by conventional heat/cold production in the reference system.

• Electricity consumption of pumps for DHW+SH and cold supply is equivalent in the SHC — systems and the reference systems.

For the measured solar heating and cooling system (numerator in (Equation 14) the energies according to Figure 1 are as following: The measured boiler energy supplied to the system Qboiler is equal to Q2 and the additional cooling provided from the compression chiller Qc0oiing, m[ssed is equal to Q8. The electricity consumption Eel is the sum of all electricity consumer except the compression chiller: Eel = (£ Ei)-E10-E12-E13-E15.

For the reference system (denominator in (Equation 14) the following calculations have to be done:

For the boiler the ratio of electrical energy to thermal energy is identical in the SHC-system and the reference-system.

Reference storage heat losses according to IEA SHC Task 26 and with reference to ENV 12977-1 (2000):

Qlossref… Reference storage heat losses [kWh/a]

VD.. .Average daily hot water consumption [Liter/day]

TT.. .Set point temperature of the hot water tank [°С], 52.5°C is used for this

Ta. Ambient temp. around the hot water tank [°С], 15°С is used for this in

The reference boiler energy supplied to the system therefore is:

Qboiler, ref = Q SHc(Q3) + QsDHcw (Q4) + Qf (Equation 17)

On inspection of Figure 2 the electrical energy consumption in the ref. system sums up to:

Er1f = E DfW, el(E5) + Ef (E4) + Ercrfsupply (E9) + E^”1" + E^ (Equation 18)

 Eel, boiler ref * aS

 = asHc * (QSHc(Q3) + QDHw (Q4) + Qross)

Where the electrical consumption of the boiler in the reference system is given through:

Fans’ electricity consumption for the conventional ventilation system is calculated based on the measured electricity consumption of the desiccant cooling system (SHC_max,) and corrected by the ratio of theoretical design pressure losses of the conventional ventilation system to the desiccant cooling system (based on datasheet of the DEC system). The electrical consumption of the two fans of the ventilation system can be estimated by:

Б™пРє1 = E™^1 • f (APREF, APdec ) (Equation 20)

Considering that for each fan of the system, the electrical power is given by (with V in [m3/s] and AP in [Pa]):

AP • V

EFan =——- W ]

Assuming the same n and flow rates for the fans of the reference and DEC system, the electricity power of the reference system can be calculated as following:

All the AP are known for the DEC system. The AP for the reference Air Handling Unit (AHU) can be estimated considering in the calculation only the components used, assuming that normally the pressure losses of each component of the DEC Air Handling Unit are known from the manufacturer of the AHU.

 + (E16 + E17) * f (AP)

These calculations result in the definition of the electrical consumption of the reference system in terms of data measured in the SHC-system:

Qcooling, ref = Q cHC= 6acm(Q7) + 6bup(Q8) + Qdec — Q10b 25)

Where QDEC is: Thermal cooling energy delivered from DEC system in the SHC-system in terms of enthalpy difference between ambient air and supply air (latent and sensible heat has to be taken into account!)

The Seasonal Performance Factor SPF for the compression chiller in the reference system can not be known exactly. The following possibilities are proposed:

• SPFmeas measured SPF in the monitored SHC system

• SPFref proposed SPF for a chiller in the reference system: SPFref = 2.8

3. Expected results

The target from the presented monitoring procedure is to have a common base for the monitoring of solar assisted heating and cooling systems, allowing a comparison of the performance of different systems and allowing the elaboration of a learning curve of the following years.

Status of the project

The developing process has comprised design and production of process equipment and processing forms for paddle vanes and turbine casing as well as internal heat exchangers.

Although there are substantial technological challenges to be solved related to the design, construction and test of the proto type, the primary challenge is to end up with a system that is commercially competitive.

Design and construction of prototype

The design of the turbines together with the bearing systems was fulfilled at the end of February. The conceptual design gave adequate data to produce the 4 different turbine parts and further design of the turbine house, the bearing system, the spindle etc. all to manage high speed on the spindles.

The turbine parts were delivered in the start of April and installed together with the external components (evaporator, condensers, fans etc.). The turbine sealings/washers were later changed to manage adequate higher temperatures.

Through the sampling of the parts several changes were made e. g on the sealing on the connections between external parts for keeping a specific vacuum in the pipelines.

Testing

The external cooling and heating components have been designed, produced and mounted/installed for test purpose e. g. prepared for vacuum, setting up for optimal measurement of temperature, pressure etc.

Furthermore a PC and a data logger has been installed to make the measurement on different temperatures and pressures around the turbines and external components. The RPM-counter uses light to measure the speed — the power used to heat the water (in stead of a solar collector) is measured; PC- programs to calculate the performance on the proto type using test data (temp., pressure, RPM, power) are completed.

Test series on the proto type are going on in august 2008. Report on the test results will firstly be available on www. ac-sun. com.

Newest: Logged data from the first tests on the proto type gives positive results and confirms the function in the design. The expander delivers the compressor capacity as expected and used in the cooling process for air condition.

Plans for further testing

The proto type test is followed up by making 3-5 test units mounted with solar panels and placed around in the surroundings mainly in the southern Europe for optimal test conditions.

3. Conclusion

The AC-Sun system is a new concept for solar driven air-condition.

It is expected to have much higher efficiency than other soar driven systems as well as it is expected to be manufactured to much lower cost than other solar driven systems.

If expectations are met the system should have the potential to overcome market barriers for solar driven air-condition.

The challenge has been the design and construction of the steam driven turbine which rotate at very high speed.

For the moment (August 2008) a prototype had been constructed and is being tested. The testing until now has detected problems which have been solved. If the further testing in the coming months perform successful further prototypes will be produced and tested under real conditions in Southern Europe.

The AC-Sun system is reported and modelled as part of the Danish participation in IEA SHC Task 38 Solar Airconditioning and Refrigeration

References

[1] S. A.Klein, (1992-2008). EES (Engineering Equation Solver) PC-program, © F-Chart Software 2008.

[2] TRNSYS (TRaNsient SYstems Simulation program), © 2002 The Board of Regents of the University of Wisconsin System.

Direct cooling (1)

 Figure 3 shows the temperatures in the rooms, the ambient temperature and the forerun temperature of the chilled ceiling. On this particular day (07/31/2008) no cooling energy was needed in the morning and the chiller started operation at 9:45. The ice-storage had not been discharged. Time [hh:mm:ss]

Fig.3: Temperatures in the rooms

The room temperatures did not exceed 26 °C until 17:00. The chilled ceiling system was almost not able to transfer the cooling energy to the rooms. Hence, the system ran on very low temperatures in the afternoon at 13:00 to 15:00. At 16:00 the cooling performance of the chiller was slowly dropping. The reason lies in the decreasing performance of the solar collectors when the azimuth is increasing. In addition the sky became cloudy. Hence, the heat gains dropped and the cooling power as well (refer Fig. 4 and 5). After 17:00 no notable cooling power could be generated.

Figure 5 shows the operational conditions of the chiller. The average COP over the day was 0.61.

 900

 □ Generator

 □ Evaporator

 800

 700

 600

 500

 400

300 200 100

09:00:00 10:00:00 11:00:00 12:00:00 13:00:00 14:00:00 15:00:00 16:00:00 17:00:00 18:00:00 19:00:00

Time [hh:mm:ss]

Fig. 4: Heat Fluxes (direct cooling)

 Time [hh:mm:ss] Fig. 5: Operating conditions absorption chiller (direct cooling)

Analysis tools

Objective analytical investment tools, such as Net Present Value (NPV) and Internal Rate of Return (IRR), were used to process the data and set them out to be compared.

Values of NPV and IRR at the end of the estimated service life of the system, 20 years, allow to deciding the economic viability of the installation. As soon as NPV becomes positive, and IRR higher than the banking interest rate, the system starts to be economically viable.

To measure the degree of viability of the installation, the period of time required to reach a positive NPV and an IRR higher than 5% was calculated.

Absolute values of cost and maintenance were also taken into account, showing the influence that each of the savings have, depending on the housing typology considered.

1.5. Results

As shown in figure 2, the cost of installation depends on the typology of housing and the configuration of the system, being the centralized compression system (which is identical for all the cases) the

cheapest one, and the centralized absorption system (also identical for the three housing typology) the most expensive one. Decentralized compression system is more expensive as far as the size of the individual house is smaller, because of the high number of splits required.

6-apartment-building decentralized system is the one which has the highest energy consumption, because of the high simultaneity of use. In absolute terms, the energy consumption cost of the absorption system is the cheapest, followed by the centralized system and the decentralized systems in decreasing surface order respectively.

Once shown the absolute initial and operation costs of all the systems, the results of the comparison between using an absorption cooling system and the rest of the considered are described.

The objective analysis of costs and savings data from the different technologies threw the results shown in the graphics below:

Fig. 4.a Period for NPV>=0 Fig. 4.b Period for IRR>=5%

As expected, the time period required to reach a balance between incomes and expenses of the absorption system, without external aids, exceeds its service life-time. Only considering environmental criteria, its installation could result profitable. The most probably scenario, SBS_CHT_SHW, is the one which gives the best results, with a minimum period of less than five years to get a positive NPV.

It is followed by the second scenario, which only considers the public subsidy. The complementary energy savings obtained by using the solar field to pre-heat sanitary and central heating water are not enough to justify the installation of such a big-surfaced solar field.

Related to the energy consumption of each housing typology, the time required to get a positive NPV grows with the size of the house, being proportionally inverse to the simultaneity of use.

Looking at Fig. XXX, the values of NPV and IRR in the second and third scenarios make the system economically attractive to be installed. It is clear that the solar cooling system will be more viable as far as the consumption of the electric system is higher. Maximum consumption occurs in the building of 6 apartments, where the bigger simultaneity factor increases energy needs.

Taking into account the second and fourth scenarios, the fact that the subsidy was granted or not, could be the angular key_that makes viable the project.

Fig. 5.a NPV in the 20th year Fig. 5.b IRR in the 20th year

advantage of the additional uses of the solar field. Despite the fact that the operation costs are very low, its high initial cost hinders a reasonable period of return.

Due to the small number of low-powered absorption chillers, it would be profitable to centralize them, refrigerating the air of a group of small houses. The smaller the house, the bigger the simultaneity of the use factor, and consequently, the bigger the energy consumption avoided.

By considering a public subsidy to the solar field, profitability could turn positive. That is not enough; profitability rate should be higher to counteract the big initial cost. By adding the effects of sanitary and central heating hot water pre-heating, positive results are obtained.

The most profitable housing typology to use the solar cooling systems is the 6 apartments building, because of its high simultaneity of use coefficient and the habitual splits overpowering.

To conclude, it is remarkable that the results of this analysis are very close to getting a real economic viability, but it requires public support and a large campaign to make people aware of environmental care. This would boost the number of installations, increasing the investment in R&D, which will cause the development of the solar cooling sector, promoting its economic profitability..

References

[1] ANAGNOSTOU, J., PRITCHARD, C., TSOUTSOS, T. et al. (2003). “Solar cooling technologies in Greece. An economic viability analysis”. Applied Thermal Engineering, 23, 1427-2439.

[2] FLORIEDS, G. A., KLGIROU, S. A., TASSOU, S. A. et al. (2002). “Review of solar low energy cooling technologies for buildings”. Renewable and Sustainable Energy Reviews, 6, 557-572.

[3] GARCIA CASALS, X. (2006). “Solar absorption cooling in Spain. Perspectives and outcomes from the simulation of recent installations”. Renewable Energy, 31, 1371-1389.

Energy equations

The energy equations related to the adsorber, which will be given next, correspond to a multi-tubular system, whose inner surface exchanges heat with the water coming from the hot storage tank or from the water supply network, depending on the stage of the cycle. The adsorbent occupies the space delimited by the external wall of the tube and the corrugated fins.

Fig. 2 — Fin-tube heat exchanger and the simplification by annular fins

For the heat transfer in the adsorbent medium, the following model assumptions have been considered: (a) the pressure is uniform; (b) the heat conduction is two-dimensional (axial and radial) as detailed in Fig. 2; (c) the adsorbent-adsorbate pair is treated as a continuous medium in relation to thermal conduction; (d) the convective effects and pressure drops are negligible; (e) the condenser and evaporator are ideal, i. e. they have a constant temperature during the isobaric phases; (f) all the adsorbent particles have the same properties (including shape and size); they are uniformly distributed throughout the adsorbent, and in local thermal equilibrium with the adsorbate and the surrounding gaseous phase; (g) the gaseous phase behaves as an ideal gas; (h) the properties of the metal and the gaseous phase are assumed to be constant; (i) the properties of the heat transfer fluid, as well as those of the adsorbate, are considered as temperature dependent. It results the following equation

[Pi (CPі + aCP2 )] ] = к V2T + qst p ^ (2)

d t d t

where Cp is the specific heat (indices 1 and 2 refer to the adsorbent and the adsorbate, respectively), p the specific mass and к, the conductivity of the adsorbent. The total derivation of the concentration, a, is given by

The da/dt term depends on the process that occurs in the adsorber. In the case of an isosteric process it is zero and for adsorption or desorption process, the term d lnp/dt is zero. Then, the energy equation for the adsorbent can be written as

where u is a function of the process, 0 for isosteric and 1 for adsorption or desorption process. The condition in the middle of the adsorbent material, between two fins or two tubes comprises the adiabatic boundary condition. Other boundary conditions are in the interface between the adsorbent and the wall of the tubes and the fins. To solve the Eq. (4), the temperature on the wall of the tubes and on the fins are considered known, recalculated for each simulation step by the energy equation

to the tubes and the following to the fins

where P is the perimeter, A is the area, Tt is the tube temperature, Tw is the water temperature, h is the conductance at the interface tube/adsorbent, and hfi is heat coefficients between the fluid and the tubes. The hfi is evaluated as the method described in [3]. The boundary conditions of Eq. (5) are adiabatic in both extremities and, to Eq. (6), they are adiabatic in the middle of the fin and known in the interface tube/fin. The temperature of the water is given by

were mw is the mass flow of water. To solve the system of equations formed by Eq. (1), (4), (5), (6) and (7) a mixed finite-difference and finite-volume method was used and the input data is chosen to be the temperature and the mass flow of the hot water, the number of tubes, the number and the thickness of the fins and the material of the fin. Additionally, the porous medium properties must be known, especially к and h. According to [4], for the AC-35 activated carbon: к = 0.19 W/mK and h = 16.5 W/m2K.

Evaluation of Solar Desiccant Cooling System for Field Test Office Building

The annual energy consumption for space heating and cooling in a office building of Japan is 359GJ/m2 [4]. About 40% of the annual energy consumption for space heating and cooling will be expected to be reduced using this solar thermal system for the similar type of the office building. Therefore, the reduced annual energy consumption for space heating and cooling in the field test office building is assumed by 57.6GJ/year. The reduced crude oil and the CO2 emission are estimated by 1.5kL and 4ton using the rates of 38.7GJ/kL and 2,649kg-CO2/kL, respectively.

The initial cost of this solar thermal system was 20 million JPY included the construction fee. If this solar system is used for 20 years, the CO2 reduction cost is 250 JPY/kg-CO2.

3. Conclusion

The solar desiccant cooling system was developed and the system performance was described in this paper. This system was developed as the passive solar thermal system using the renewable energy without the heat source equipment and the dehumidification cooling system for the fresh air.

From the field test results, it was found that the solar desiccant cooling system for office building

was effective throughout a year.

Acknowledgement

This study was supported by the research funds of NEDO project, Research and Development of

Technologies for New Solar Energy Utilization Systems for FY2005-F2007, and Grant-in-Aid for

Scientific Research (C)(19560598). The authors would like to express their sincere thanks to the

support.

References

[1] H. Roh, K. Suzuki, Research and Development of Air-based Passive Solar Dehumidification Cooling System, Part 1 Operation Test of Solar Dehumidification Cooling System, Proceeding of JSES/JWEA Joint Conference 2006 (Renewable Energy 2006 Japan Day), pp.309-312. (in Japanese)

[2] H. Roh, K. Suzuki, Research and Development of Air-based Passive Solar Dehumidification Cooling System, Part 2 Field Test of Solar Dehumidification Cooling System in Summer, Proceeding of JSES/JWEA Joint Conference 2007, pp.405-408. (in Japanese)

[3] S. Song, K. Suzuki, H. Roh, Study on the Performance of Dehumidification Cooling System with Solar Thermal and Well Water, Summaries of Technical Papers of Annual Meeting Architectural Institute of Japan 2008, D-2, pp.1209-1210. (in Japanese)

[4] Heat Pump & Thermal Storage Technology Center of Japan, White Paper of Heat Pump and Thermal Storage, pp.335, Ohmsha, 2007. (in Japanese)

Principal of Operation

Three external circuits are connected to the Millennium MSS Air Conditioning:

Thermal heat source (e. g. MSS solar collectors)

Air conditioning distribution system for cooling and heating (e. g. radiant floor, fan-coil units)

Heat sink for charging and discharging (e. g. swimming pool, cooling tower, air cooled condenser or geothermal holes)

Millennium Mss Air Conditioning System is a modular absorption machine that differs from the “standard” Lithium Bromide type absorption machines in three main aspects:

It has internal storage in each of the two accumulators. This allows the machine to store chemical energy with a very high density. This energy can subsequently be used both for cooling and heating. It is important to emphasize that this is chemical energy, not thermal energy that is stored.

It works intermittently with two parallel accumulators (Barrel A and Barrel B).

It is designed to use relatively low temperatures and is hence optimized for usage with solar thermal collectors. It also works with a stable temperature inside the accumulators, which in turn allows for an effective use of solar thermal collectors.

Millennium MSS Air conditioning system made up of two “barrels” each consisting of a reactor and condenser/evaporator. The two barrels can operate in parallel.

Cooling

The water returns from the distribution system at a higher temperature than when it left the condenser / evaporator (we have cooled the building). This heat causes the water in the evaporator to boil and the steam passes down to the reactor, where it condenses, since the reactor is relatively cooler. Steam that condenses into water in the reactor will dilute the LiCl solution. The diluted LiCl solution is then pumped through the filter basket, where it mixes with the salt and regains its saturation. The saturation is needed to continuously provide a temperature difference between the condenser/evaporator and the reactor.

680 mm 680 mm

Barrel ABarrel B

 Mode Storage Capacity * Maximum Output Capacity ** Electrical COP[7] Thermal Efficiency Cooling 60 kWh 10/20 kW 77 68% Heating 76 kWh 25 kW 96 160%
 * Total storage capacity (i. e. including both barrels)

** Cooling capacity per barrel: 10 kW cooling is the maximum capacity. If both barrels are used in parallel (double mode) the maximum cooling output is 20 kW and the maximum heating output is 25 kW.

Heating

Heating is just cooling in reverse, meaning that the charged energy is extracted as heat by connecting the condenser/evaporator to the heat sink and the reactor to the distribution system. Water returns from the distribution system at a lower temperature than when it left the reactor (we have heated the building). This water boils the water in the condenser/evaporator and steam passes down to the reactor. Steam condenses into water which dilutes the LiCl solution in the reactor. The diluted LiCl solution is pumped through the salt filter basket where it mixes with the salt and regains its saturation. The saturation is needed to continuously provide a temperature difference between the condenser / evaporator and the reactor. During discharging, the heating energy is extracted by connecting the evaporator to the heat sink and the reactor to the distribution system. Under charging, heat can also be extracted by connecting the condenser to the distribution system under charging mode.

5 m3 solar ammonia-carbon adsorption refrigerated container for food preservation

R. E. Critoph

University of Warwick, Coventry, CV4 7AL, UK.

R. E. Critoph@warwick. ac. uk
Abstract

An ‘alpha-version’ solar adsorption refrigerator for chilled food preservation is being developed jointly by the University of Warwick and Advanced Technology Materials Inc. The requirement is to cool a 5m3 insulated container in temperatures up to 40°. The prototype uses an ammonia — active carbon pair in a 2-bed cycle with heat and mass recovery. An ice-bank is created within the container during the day to act as a thermal store. The driving heat is supplied by 10m2 of evacuated tube collectors via a pressurised water loop. Simulations suggest that between 1 and 2 kW of cooling can be supplied given reasonable levels of insolation and that the adsorption cycle time may be made a simple function of insolation alone.

1. Introduction

Previous research at the University of Warwick on a mobile air conditioning system [1] has resulted in a patented concept [2] for a highly compact solid sorption reactor. In the work reported here the technology is applied to solar powered refrigeration. There is a requirement for maintaining chilled food at 0-5 °C in transportable containers in remote areas away from grid electricity. The conventional technology solution is to use vapour compression refrigeration powered from motor-generator sets. The University and Advanced Technology Materials Inc. (USA) are collaborating in the development of a solar thermal powered system, which will have parasitic power for controls etc. delivered by PV’s. The ‘alpha’ version, is due for field testing in Arizona from November 2008.

2. Specification

The standard insulated container, manufactured by CMCI (Figure 1) has external dimensions 2.4 m x 1.5 m x 2.1 m and internal volume of 4.7m3 . It is normally cooled by a conventional vapour compression chiller, rated at about 2kW cooling at 2°C. It is required to maintain normal use at ambient temperatures of 40°C using a solar thermal cooling system.

3. Design

Naturally, any solar powered system requires thermal storage and it has been decided to use an ice bank integrated with the flooded evaporator of the refrigerator. Approximately 50 kg of ice is needed and this is incorporated into a vertical wall within the container. The wall has enough fins extending into the cold space so that cooling within the container is achieved by natural convection. Figure 2 shows the complete evaporator/ice-bank assembly and Figure 3 shows the flooded evaporator alone.

The evaporator consists of approx. 40 vertical half inch diameter tubes with a large reservoir above and parallel feed below. Later versions will have direct expansion evaporators which will have the advantages of lower mass, lower cost and reduced refrigerant charge, but the overwhelming advantage of a flooded evaporator is that requires less development and is comparatively risk free. Each of the vertical tubes fits tightly between the fins of an aluminium extrusion that forms part of the ice-bank. Without this heat transfer enhancement, towards the end of the process of freezing the water, the evaporating temperature would drop significantly as heat from the freezing front had

to be conducted through an increasing thickness of ice, thereby reducing the system COP (cooling power / driving heat input).

The same aluminium extrusion is used on the outside of the ice tank to transfer heat to the cold space.

The refrigeration system is based on an adsorption cycle using ammonia as refrigerant and active carbon adsorbent. The plate heat exchanger (sorption generator) developed for this and other applications is shown in Figures 4 and 5. It is of low thermal mass (for good COP) and allows rapid cycling which reduces the physical size of the sorption generator for a given cooling power. Two of these generators, about 200mm on one side have been used in ‘TOPMACS’, an EU project to create a car air conditioning system that is driven by the waste heat of the engine. Laboratory operating results are given in Figure 6, which shows a mean cooling power of 1.6kW. The solar ice-making application is similar and design work is based on two similar but improved units. In the original generator, granular carbon in 4mm thick layers was sandwiched between stainless steel shims containing numerous water channels for heating and cooling the

carbon. The new design will utilise a more highly conductive (~2.0 W m K-1) carbon developed by ATMI

which will enable the use of 12mm carbon layers, reducing cost, complexity and thermal mass. The two beds will be operated in a simple cycle with both mass and heat recovery, with typical cycle times of 2 minutes.

 Water channels

 Figure 5: Solar refrigerator plate sorption generator core with 12 mm slots

The original design for car air conditioning was heated or cooled by unpressurised water. The solar collectors are expected to operate at well above 100°C and so the choice had to be made between using a heat transfer fluid or pressurised water. The heat transfer properties of water are so superior that pressurised water was selected.

A schematic of the whole refrigeration system is shown in Figure 7. Hot water from the solar collectors is pumped to either G1 or G2, heating the carbon within and desorbing ammonia. The ammonia flows through a check valve to the condenser where heat is rejected, through the float type expansion valve to the flooded evaporator / ice bank, where it boils and produces useful cooling. From there it passes through a check valve and into the other generator, where it is adsorbed. It is necessary to remove the heat of desorption via another pressurised water loop to an

air cooled heat exchanger. At a suitable time (optimised for maximum cooling power) the cycle is reversed and the bed that was desorbing becomes the adsorber and vice versa. At the change-over, the performance can be improved, firstly by briefly opening the mass recovery line to equalise pressures in the two generators and then by using the heat recovery loop to pre-heat one bed with the reject heat of the other.

Given the high collector temperatures, the only commercially available options are evacuated tubes. The collectors used in our simulation models are Thermomax DF100 2m2 panels which feature direct flow of the fluid (water at up to 8 bar pressure) through the tubes. Figure 8 shows the manufacturer’s performance data. 10m2 collector area will be used to obtain a peak cooling power of up to 2 kW at an ambient temperature of 40°C in a desert environment.

A critical area of design is the waste heat rejection from the condenser and adsorbers. This is done using conventional fan coils and with attention being paid to minimising the fan power. The design compromises are critical. A small compact heat exchanger may have higher temperature differences which lead to lower COP and hence more heat to be rejected. It may also need more fan power and since parasitic electrical power will be met from PV, this must be minimised. Conversely, very large heat exchangers could be both impractical and costly. The compromise chosen uses a direct condenser measuring 650 h x 900 l x 570 w and with a 66W fan motor and a cooler measuring 650 h x 900 l x 470 w with a 102W fan motor.

 Collector efficiency, G=1000W/m2, Tamb=40°C Figure 8: Performance curves of chosen evacuated tube collectors

4. Simulation

The operation of the complete system has been modelled in Matlab to assist the design. The operation of the chiller has to be modelled at a timestep of about 0.001 s which is obviously impractical for modelling several days of operation. This problem has been overcome by deriving a pseudo-dynamic model in which the chiller is assumed to respond much more quickly (within minutes) than changes in the load or ambient conditions.

 Figure 9 : Performance envelope for ambient temperature 30°C and evaporating temperature -10°C

An example of the approach is given in Figure 9 in which each point (derived from detailed simulation every 0.001 s) corresponds to a balance between the heat input from the collectors at the particular insolation and ambient temperature, together with a particular cycle time (control

parameter) and evaporating temperature (corresponding to the state of the load). The envelope of the points (linear and quadratic are shown) gives the instantaneous cooling power corresponding to the best control strategy for that particular evaporating temperature and ambient temperature for the full range of insolation.

A set of these correlations for a range of evaporating and ambient temperatures may be combined empirically to yield a polynomial function for optimum cooling power under any conditions which can act as input to the model of the ice-bank and cold box. Preliminary examination of these results implies that the cycle time can be made a simple function of insolation only (i. e. ignoring ambient and evaporating temperature) with comparatively little penalty. Figure 10 illustrates this for the particular case of an ambient of 30°C and evaporating at 10°C.

Basic geometries

As comes out of preceding relations, guidelines for dimensioning of air-soil heat exchangers relate to careful examination of convective and diffusive heat exchange coefficients, which themselves depend on geometrical parameters (eq. 5-6).

As is given by eq. 5, the convective air/pipe heat exchange ha increases almost linearly with air velocity, with a lesser dependence on pipe diameter, reaching values between 4 and 16 W/K. m2 for velocities between 1 and 4 m/s (fig. 3, left column).

Diffusive heat exchange on its turn depends on the available soil layer around the pipes and the way heat diffusion can actually take place. We therefore will consider two distinct geometries, with quite different

the airflow may be charged and discharged in the soil in the same radial way as before (fig. 2, right). To the contrary, charge and discharge of the yearly heat wave saturates the immediate vicinity of the pipes and can hence only take place in plane mode, downwards into the ground, on a surface S defined by the system size (length by width). Accounting for the reduced available capacity, the diffusive coefficient h now reduces to 0.6 W/K. m2, which forces the resulting amplitude-dampening coefficient to some 0.5 W/K. m2, independently of air velocity (fig. 3). As a consequence, such a geometry enables to reach daily amplitude-dampening with the same pipe length as before, but almost without dampening of the yearly amplitude (see also fig.5). Note that such a compact geometry may also be realized in a multi-layer configuration, in which case the slight effect on yearly amplitude will be all the more negligible.

3. Guidelines

For both the above defined geometries it is now possible, via equation 3, to give dimensioning rules in terms of necessary pipe length for “complete” amplitude-dampening, which we define as a residual amplitude of e-2 ~ 15%. The resulting guidelines are as follows:

• If the point of interest concerns dampening of the yearly oscillation, it is necessary to work with distant of each other and deeply buried pipes, with approximately 3 m soil around each. In this case, a rough dimensioning rule for complete amplitude-dampening is 30 m pipe per 100 m3/h airflow (fig. 4, left), for diameters yielding an air velocity in the 1 to 4 m/s range. As pointed out above, the daily amplitude will have vanished on approximately half that distance, with the same rules as hereafter.

• If the point of interest concerns dampening of the daily oscillation, only with slight or negligible dampening of the yearly component, it is sufficient to work with close to each other and shallow buried pipes, with approximately 15 cm soil around each. Typical required length is roughly 15 m per 100 m3/h airflow (fig. 4, right), a relation which however isn’t as linear anymore and slightly depends on the pipe diameter and associated air velocity.

Note that preceding guidelines for pipe length concern residual amplitude of e-2, but may be extrapolated by way of the exponential form of eq. 3. For example, the guidelines for the compact configuration (fig. 4) indicate that a 200 m3/h airflow in a 22 cm diameter pipe (~ 1.5 m/s) needs 55 m for complete daily dampening; a 20 m pipe hence would yield a residual amplitude of e — 2-(20/55) = 48%.

It should finally be stressed that the model on which these guidelines have been established does not take into account perturbation of the storage mechanism by thermal drive from upper surface (ambient or building), nor the possible effects of transient airflow or latent heat exchanges. As a complement to the basic guidelines elaborated in this study, these effects can however be investigated by way of numerical simulation.

ha : convective exchange (air/pipe) hs : diffusive exchange (pipe/soil) h : total exchange (air/soil)

Storage of day/night oscillation, extensive or compact geometry

Storage of summer/winter oscillation, extensive geometry

20 18 16 14 12 10 8 6 4 2 0

 1 2 3 4 0 0.1 0.2 0.3 0.4 0 1 2 3 4 20 cm — 25 cm
 diametre

Fig. 4: Guidelines for dampening of daily or yearly amplitude.

4. Validation

Validation as well as illustration of these behaviors and guidelines will now be given on hand of an extensively validated finite element numerical model for buried pipe systems [4], which accounts for fully three dimensional heat diffusion in soil and flexible border conditions, as well as for sensible and latent heat exchanges (latter not used within this study).

Both the considered configurations consist of a layer of 20 cm diameter pipes, each swept by a 200 m3/h dry airflow, with hourly input given by the standard annual meteorological temperature for Geneva. The pipes are buried in a soil with same thermal characteristics as above, in one case at 310 cm depth (center of pipe) and with 620 cm inter-axial distance, in the other case at 25 cm depth and with 50 cm inter-axial distance. Superior border conditions are in both cases taken as adiabatic, whereas a sufficiently important soil layer is taken into account for the heat wave to expand as deep as necessary (adiabatic border conditions 15 m below pipes). The arrays are supposed to consist of a sufficient amount of pipes for lateral border effects to be negligible, so that the system may be described by way of a unique pipe (adiabatic conditions at inter-axial distance).

Simulated hourly temperatures are depicted in form of daily minima and maxima profiles at 0, 20 and 50 m distance (fig. 5, top), and confirm preceding analysis: deep and wide apart buried pipes induce combined daily and yearly amplitude-dampening, eventually reaching the constant annual average, whereas the compact pipe configuration essentially allows for daily amplitude — dampening, with a constant output over the day/night period but an almost unaltered seasonal trend.

The hourly data at every 10 m is Fourier analyzed, so that residual yearly and daily amplitudes may be compared to the values given by the above dimensioning rules (fig. 5, bottom): a very good correlation is manifest in either case, theoretical guidelines only slightly overestimating the residual amplitudes given by numerical simulation.

[1] Hollmuller P., Lachal B. (2001) Cooling and preheating with buried pipe systems : monitoring, simulation and economic aspects. Energy and Buildings, 33(5), p. 509-518.

[2] Hollmuller P. (2003) Analytical characterisation of amplitude-dampening and phase-shifting in air/soil heat-exchangers. International Journal of Heat and Mass Transfer, vol. 46, p. 4303-4317.

[3] Gnielinski V. (1975) Neue Gleichungen fur den Warme — und Stoffubergang in turbulent durchstromten Rohren und Kanalen, Forsch. Ing.-Wes., Vol.41(1), pp. 8-15.

[4] Hollmuller P., Lachal B. (2005) Buried pipe systems with sensible and latent heat exchanges : validation of numerical simulation against analytical solution and long-term monitoring, in : Building Simulation, proceedings of the 9th International Building Performance Simulation Association, 15-18 August 2005, Montreal, Quebec, Ecole Polytechnique de Montreal, Vol.1, p. 411-418.

http://www. unige. ch/cuepe/html/biblio/detail. php? id=362

Effects of regeneration on low-temperature solar organic Rankine cycles

1Tchanche Fankam Bertrand*, 1George Papadakis, 1Gregory Lambrinos and 2Antonios

Frangoudakis

Department of Natural Resources and Agricultural Engineering,

1 Laboratory of Agricultural Engineering,

2 Laboratory of Agricultural Constructions,

Agricultural University of Athens, 75 Iera Odos Street 11855 Athens, Greece Tel.: +30 (210) 529 4046; Fax: +30 (210) 529 4036 * Corresponding author, e-mail address: tfb@aua. gr

Abstract

This paper deals with the study of the effects of regeneration on low-temperature solar organic Rankine cycles utilizing R134a as working fluid. The configurations investigated are: basic ORC (organic Rankine cycle), regenerative ORC with internal heat exchanger, regenerative ORC with open feedwater heater and regenerative ORC with closed feedwater heater. The effects of the regeneration is assessed using a method of thermodynamic exergetic analysis called exergy — topological method based on the combination of exergy flow graphs, exergy loss graphs and thermoeconomic graphs. For a designed output power of 2 kW, the incorporation of an open feedwater is the most efficient way to increase the system performance. The thermal and exergetic efficiencies obtained were 3.95% and 9.16% respectively; this represents an increase of 6.83% and 6.93% in thermal and exergetic efficiencies respectively compared to the simple Rankine engine. The configuration with closed feedwater heater produces similar results but is less efficient. The incorporation of an internal heat exchanger produces negative effects. Keywords: organic Rankine cycles, exergy-topological method, regeneration, solar energy

1. Introduction

The lack of electricity or means of power generation seriously limits the development of many regions around the world. The increasing fossil fuels prices and the pollution derived from their combustion in power plants are two important reasons for which the attention should be shifted towards renewable energy sources. As can be seen today, the price of oil will still increase and the access to grid electricity will become more and more difficult for populations leaving in remote areas. In this context, solar energy can play a major role in sunny areas. Recent works [1-3] on low-temperature organic Rankine cycles show that this technology could be a solution for providing fresh water and electricity, and perhaps thermal comfort. Although the system’s components are well known technologies, the integration to a fully and efficiently working system is a challenge. A serious drawback of the low-temperature organic Rankine cycle is its low efficiency. An examination of the basic Rankine cycle reveals that the heat is transferred to the working fluid at relatively low temperature; this reduces the system thermal efficiency. A solution to remedy this shortcoming is to use a regenerator or a recuperator, which gives birth to several organic Rankine cycle configurations.

In this paper, the effects of the regeneration on the modified cycle with internal heat exchanger and regenerative cycles with open and closed feedwater heaters are analyzed using a new exergy approach called exergy-topological method [4-7].