negst5_d1-report solar cooling etc

NEGST – NEW GENERATION OF SOLAR THERMAL SYSTEMS is a project financed by the European Commission DG TREN within FP6

SUMMARY The overall objective of NEGST project is the development and market introduction of more cost-efficient solar thermal systems, particularly for domestic hot water preparation and / or space heating and cooling. This “new generation” of solar thermal systems represents a further development of today’s system technology with respect to the improvement of thermal performance and the reduction of system costs. This is essential for contributing to the European Union's Action Plans with regard to the reduction of greenhouse gas emissions and the cost effective supply of renewable energies. In this framework, NEGST Work Package 5 aims to assess the potential of solar thermal systems for advanced applications, such as cooling and desalination. An enlargement of the utilization field, traditionally limited to hot water preparation and/or space heating, could in fact positively contribute to sustain a further diffusion of solar thermal systems and, at the same time, to promote innovative concepts. In this context, an overview on all cooling and desalination systems, which can be suitable coupled with low to medium solar collectors, constitutes a preliminary step towards the aforesaid objective. Thus, the first step of Work Package 5 is an investigation, relevant to each considered process, which main issues concern: the brief description of the technologies under investigation, the assessment of the global energy requirements, economics, the main advantages and drawbacks in particular in view of the coupling with solar systems, the development status and their possible commercial diffusion. The result of this investigation is a “Technical Status Report”, which main target is to support the selection of the most promising solar technology for cooling and desalination, worth to be subject to a further analysis in the following steps of the NEGST WP5. For this purpose, a preliminary screening criterion, based on energy saving approach, is introduced. Internet More information are available on Project website: http://www.swt-technologie.de/html/negst.html

WP5.D1: Technical status report on solar desalination and solar cool ing

Dissemination level: Public

Authors: V. Sabatelli, G. Fiorenza, D. Marano (ENEA)Reviewer: P. Oliaro (PoliMI)

November 2005

CONTENTS: INTRODUCTION Description of main objectives and expected results. SOLAR COOLING Overview on cooling technologies which can be in principle coupled with low to medium solar collectors. Focus area: � Overview of traditional cooling

systems and their markets in European countries

� Description of the main cooling technologies: - Absorption chillers - Adsorption chillers - Thermo Chemical

Accumulators (TCA) - DEC systems - Other cooling technology

� Energy comparison between different solar cooling technologies based on primary energy saving.

SOLAR DESALINATION Overview on solar desalination plants currently available. Focus area: � Overview of traditional

desalination technology in European countries

� Description of the main desalination plants: - Multiple Effect Evaporation - Multi-stage Flash Evaporation - Application for small users

� Comparison between different technologies under investigation.

CONCLUSIONS Description of main results and open question in view of the subsequent selection of the most promising solar technology for cooling and desalination. ANNEX A Country data on conventional air-conditioning systems. Information about market trends.


WP5.D1: Technical status report on solar desalination and solar cool ing Date: November 2005 page 2 of 102 pages

TABLE OF CONTENTS

LIST OF CONTRIBUTIONS ___________________________________________________________ 4

1. INTRODUCTION ________________________________________________________________ 5

2. SOLAR COOLING ________________________________________ _______________________ 6

2.1 OVERVIEW OF TRADITIONAL COOLING SYSTEMS AND THEIR MARKETS IN EUROPEAN COUNTRIES________________________________________________________________ 6

2.1.1 Demand of AC in Europe ____________________________________________________ 6 2.1.2 Energy consumption________________________________________________________ 7 2.1.3 The market: present situation and trends ________________________________________ 8 2.1.4 Solar cooling installations in Europe___________________________________________ 10

2.2 ABSORPTION CHILLERS___________________________________ _________________ 11 2.2.1 Physical principles ________________________________________________________ 12 2.2.2 Classification of absorption chillers ___________________________________________ 13 2.2.3 Cycle Performance and Maintenance _________________________________________ 15 2.2.4 Coupling of the Absorption Chillers with a Solar Heating System ____________________ 16 2.2.5 Investment costs _________________________________________________________ 19 2.2.6 Operating plants __________________________________________________________ 20

2.3 ADSORPTION CHILLERS___________________________________ _________________ 21 2.3.1 Physical principles ________________________________________________________ 21 2.3.2 Classification of adsorption chillers ___________________________________________ 21 2.3.3 Operating plants __________________________________________________________ 23

2.4 THERMO CHEMICAL ACCUMULATOR (TCA) ____________________ _______________ 24 2.4.1 Physical principles ________________________________________________________ 24 2.4.2 Classification of TCA machines ______________________________________________ 25 2.4.3 Operating plants __________________________________________________________ 27

2.5 DESICCANT COOLING FOR AIR-CONDITIONING ________________ ________________ 27 2.5.1 Physical principles ________________________________________________________ 28 2.5.2 Coefficient of performance __________________________________________________ 29 2.5.3 Limits of the thermo dynamical principle _______________________________________ 30 2.5.4 Configuration of Desiccant cooling systems_____________________________________ 32 2.5.5 Desiccant Wheel _________________________________________________________ 33 2.5.6 DEC-system with liquid sorbent materials ______________________________________ 37 2.5.7 Examples of DEC-System plants _____________________________________________ 38

2.6 OTHER COOLING SYSTEMS _________________________________________________ 40 2.6.1 Organic Rankine Cycle_____________________________________________________ 40 2.6.2 Ejector air-conditioning system_______________________________________________ 42

2.7 COMPARISON OF THE DIFFERENT TECHNOLOGIES UNDER INVESTIGATIO N _______ 46 2.7.1 Description of the methodology ______________________________________________ 46 2.7.2 Reference locations _______________________________________________________ 47 2.7.3 Heating and cooling needs__________________________________________________ 47 2.7.4 Configuration of solar cooling systems_________________________________________ 51 2.7.5 Results _________________________________________________________________ 53

3. SOLAR DESALINATION ____________________________________ _____________________ 58

3.1 OVERVIEW OF TRADITIONAL DESALINATION TECHNOLOGIES IN EURO PEAN COUNTRIES_______________________________________________________________ 58

3.2 MULTIPLE EFFECT EVAPORATION _______________________________ ____________ 62 3.2.1 Physical principles ________________________________________________________ 62 3.2.2 Research and demonstration systems _________________________________________ 65

3.3 MULTI-STAGE FLASH EVAPORATION ___________________________ ______________ 72 3.3.1 Physical principles ________________________________________________________ 72



3.3.2 Energy Consumption and Costs______________________________________________ 73 3.3.3 Integration of Solar Heat in Multiple Stage Flash Evaporation Plants _________________ 73 3.3.4 Operating plants __________________________________________________________ 74

3.4 APPLICATIONS FOR SMALL USERS ____________________________ ______________ 76 3.4.1 Solar still________________________________________________________________ 77 3.4.2 Humidification-dehumidification ______________________________________________ 81

3.5 COMPARISON OF THE DIFFERENT TECHNOLOGY UNDER INVESTIGATION _ ________ 84 3.5.1 Selected solar desalination systems __________________________________________ 85 3.5.2 Calculation methodology ___________________________________________________ 86 3.5.3 Results _________________________________________________________________ 86

4. CONCLUSIONS________________________________________________________________ 88

ANNEX A – Country data on conventional cooling systems __________ _____________________ 91

REFERENCES_____________________________________________________________________ 98



LIST OF CONTRIBUTIONS

Arguments Authors Institutions

INTRODUCTION V. Sabatelli, G. Fiorenza ENEA

SOLAR COOLING

Overview of traditional cooling systems and their markets in European countries V. Sabatelli ENEA

Absorption chillers U. Jordan, R. Shahbazfar UniKassel

Adsorption chillers H. Drueck, E. Streicher ITW

Thermo Chemical Accumulator C. Bales SERC

Desiccant cooling for air-conditioning T. Selke ARSENAL

Other cooling systems (Rankine Cycle, Ejector air-conditioning system) G. Fiorenza ENEA

Comparison of the different technology under investigation V. Sabatelli, D. Marano ENEA

- Heating and cooling needs (§ 2.7.3) P. Oliaro, R. Adhikari PoliMI

SOLAR DESALINATION

Overview of traditional desalination technologies in European countries G. Fiorenza ENEA

Multiple Effect Evaporation M. J. Carvalho, G. Buchinger INETI - AEE

Multi-stage Flash Evaporation J. Buchinger, C. Isaksson, D.Jähnig AEE

Applications for small users (Solar still, Humidification-dehumidification) E. Mathioulakis, G. Panaras Demokritos

Comparison of the different technology under investigation G. Fiorenza, D. Marano ENEA

CONCLUSIONS V. Sabatelli, G. Fiorenza ENEA



1. INTRODUCTION

The overall objective of the NEGST Project is to support the market of solar thermal systems, as a contribution to the European Union’s Action Plans with regard to the reduction of CO2 emissions and the cost effective supply of renewable energies. In order to achieve this goal, the project provides a framework for the development of the next generation of solar thermal systems and their diffusion. In particular, the Work Package 5 aims to assess the potential of solar collectors for advanced applications, such as cooling and desalination. The enlargement of the utilization field is a measure that cannot be disregarded to facilitate a further diffusion of solar thermal systems and to stimulate the production of innovative concepts. In effect, solar systems for domestic hot water preparation represent a mature technology which application already reaches a satisfactory figure in many European countries. In any case the margins for a further market growth are rather narrow and the impact on global energy consumption limited. In addition, in South European countries use of solar systems for space heating is not much attractive. The main reason is that the over sizing of the collectors area in order to meet heating requirements too has to be paid off in a very reduced period of utilization. Therefore, as a general rule, such systems turn out to be affected by an excessive cost. On the contrary, the application of solar collectors for air conditioning or generically production of refrigerated water can have a higher potential. In fact, even though in many southern locations also the heating load is greater than the cooling load, the period of utilization of the solar system is wider. The explanation is that the cooling cycle can be reversible or, in any case, the collectors can be directly used for heating purposes during the cold season, while the opposite process is not possible without the presence of the cooling machine. A second potential application in southern regions is represented by the desalination of seawater, which can be particularly valuable in isolated areas, where the supplies of fresh water with traditional means is usually very onerous. It is to be noticed that cooling and desalination processes require in general a driving temperature higher than the average working temperature of collectors for domestic hot water production and space heating. Therefore the diffusion of these advanced applications can act as an incentive to improve the efficiency of low temperature collectors and introduce to the market more cost-effective collectors able to operate at relatively higher temperatures (up to 200 °C). The overview on all cooling and desalination systems, which can be in principle coupled with low to medium temperature solar collectors, constitutes a preliminary step towards the aforesaid objective. The main topics relevant to each considered process will be:

� brief description of the working principle � accurate assessment of the global energy requirements � short statement about economics � advantages and drawbacks in particular in view of the coupling with solar systems � development status and possible commercial diffusion

Differently from other works in the same field, aspects such as the exhaustive description of working cycles, components, operating systems and any other feature of the cooling or desalination process not closely related to the coupling with the solar source are considered as secondary issues. In fact this work is not intended as a handbook for planner and thus design criteria and examples will not be deeply investigated. On the contrary the main target is to support in the selection of some solar cooling or desalination systems worth to be subject to a further analysis in the following steps of the NEGST WP5. For this purpose, a preliminary screening criterion is introduced according to an energy saving approach. The obtained results are then to be analysed considering the system capital cost, the level of commercial maturity, the presence of technological barriers and any other key factor in order to reach a final assessment of the most promising solar systems for cooling and desalination.



2. SOLAR COOLING

2.1 OVERVIEW OF TRADITIONAL COOLING SYSTEMS AND THE IR MARKETS IN EUROPEAN COUNTRIES

Commonly, conventional air conditioning systems are divided into two main categories: � Room Air Conditioners (RAC) characterized by an individual and autonomous

appliance for household use, with a cooling capacity usually less than 12 KW. These systems include split, multi-split, single-duct and single-packaged units.

� Central Air Conditioners (CAC) are instead systems with more than 12 KW of cooling capacity, having a central refrigerating unit that use a fluid (typically water or air) to transport the “cold”.

As illustrated in Figure 2.1.1, for “central air conditioning units” exist a large variety of systems and technical options, according to the type of refrigerant employed and/or the equipments used to distribute the air-conditioned.

Figure 2.1.1: AC systems classification (Sources: EECCAC and EERAC studies).

2.1.1 Demand of AC in Europe

In the last years, the sales of conventional air-conditioning systems is growing rapidly in Europe, as a result of the increased living standards and the need to improve comfort conditions both in household and workplace. Actually, in a study of European Commission, the “Energy Efficiency and Certification of Central Air Conditioners” Report (/EECCAC/), the increase of annual additional air-conditioned floor area in buildings has been estimated from 1980 to 2000. Figure 2.1.2 shows this rapid trend in growth.



Even if the expansion of AC demand is principally related to the climatic conditions and the requirement of better comfort conditions, the development of the tertiary sector has produced in some middle European countries (Germany for example) a larger rate of growth in AC demand than in some Southern countries such as Portugal or Greece. Figure 2.1.3 shows this tendency in each national market of AC systems. In the analysis are included the RAC units, which are also predominantly used in

trade and office buildings. At present time just two countries, Spain and Italy, account for more than 50% of the entire EU market, as emphasized in the same Figure 2.1.3 that shows the distribution of additional floor area by country.

France12%

Germany11%

Greece5%

Italy25%Portugal

2%

Spain24%

United Kingdom

8%

Other13%

Figure 2.1.3: (a) Evolution of apparent annual additional building floor area conditioned from 1980 to 2000 in different EU countries – (b) Present distribution of apparent annual additional building floor area by country (Source: EECCAC study).

2.1.2 Energy consumption

Energy consumption in European domestic and tertiary sectors represents about 40% of the annual EU-15 final energy use and about a third of greenhouse gas emissions. Among these, about two-thirds are concentrated in residential sector, the remaining part in commercial buildings (Source: EU – Energy & Transport in figure, 2004). The household sector represents about 70% of total energy consumption in buildings sector. Table 2.2.1 gives an estimation of total and per capita energy consumptions, both for central and room air conditioners, in EU-15 countries.

Figure 2.1.2: Apparent annual additional building floor area

conditioned by CAC from 1980 to 2000 (Source: EECCAC study).



Total energy consumption [GWh/y] COUNTRY

CAC Systems (EECCAC Report, 2000)

RAC Systems (EERAC Report, 1996)

Per capita consumption

[KWh/y]

Austria 469 121 73 France 5010 1782 116 Germany 2286 672 36 Greece 2909 1007 371 Italy 16209 4494 361 Portugal 1020 714 175 Spain 19689 2496 553 United Kingdom 2359 446 48 Belgium 274 34 Denmark 71 17 Finland 206 51 Ireland 127 44 Luxemburg 11 0 Netherlands 605 49 Sweden 391

444

56 TOTAL 51636 12176 81 Table 2.2.1: Estimation of total and per capita energy consumption by country.

Presently, and in particular in South European countries, there is a well-established correlation between the growth of peak power electricity demand in summer season and the growth of air-conditioning sales in the small and medium size market. The use of conventional cooling equipment has introduced several drawbacks such as frequent peak electric loads, increase of electrical energy consumption and environmental problems resulting from the use of refrigerants and the increase of installed electric power generation.

2.1.3 The market: present situation and trends

The same EECCAC and EERAC studies present the market share of air conditioning systems by equipment type (not only for CAC but also for RAC systems), and their distribution for economic sector, including residential part (see Figure 2.1.4 and Figure 2.1.5).

Figure 2.1.4: Market share by equipment type for different economic sector

(Source: EECCAC Report - 1998).



Figure 2.1.5: Market share by economic sector (Source: EECCAC Report - 1998).

As it can be noted from Figure 2.1.5, the household sector is entirely characterized by RAC systems while economic sectors such as hospitals, hotels, offices and education are prevalently characterized by chillers with a modest percentage of RAC systems, packaged and multi-split units. A more uniform distribution of different cooling systems, with a weak prevalence of RAC systems, characterizes the trade sector. Regarding the equipments, as depicted in Figure 2.1.6, vapour compression chillers and split units dominate the CAC and RAC markets, respectively. Among these, thermally driven chillers represent only a small percentage with respect to total plants in operation (e.g. no more than 1% in southern European countries such as Italy, Greece and Spain). More recent data confirm this trend in the market share of technology, even though a major attention for thermally driven chillers is rising. For more details, about the present market situation (typical application, size, cost, trends, etc.) in some EU countries, refer to data and references included in Annex A.

(a) Central Air-Conditioning systems

Chillers71%

Packages8%

Rooftops7%

Splits >12KW11%

VRF3%

(b) Room Air-Conditioning systems

Multi-split7%

Single-duct18%

Single-packaged

6%

Split69%

Figure 2.1.6: Market share of CAC and RAC systems by equipment type (% of floor conditioned area).



Finally, as shown in Figure 2.1.7, it is to be expected a high growth of the AC market especially for room air conditioners, with a consequent remarkable increase of the electrical consumptions during the summer season especially for Southern Europe countries. It is to be noted that, in some cases (Italy for example), such data are underestimated in comparison with the current diffusion of air conditioning systems.

Figure 2.1.7: Evolution of cooled floor area by equipment type (Source: EECCAC Report).

2.1.4 Solar cooling installations in Europe

In contrast to the significant electric consumptions attained by cooling systems currently in operation, the recourse to solar energy, as suitable choice in opposition to conventional energy sources, is still scarcely investigated. In fact, at the present time the total cooling capacity of solar powered air-conditioning systems installed in Europe is estimated in only 6.3 MW.

(a) Installations

27

19

6

4

3

3

2

2

0 5 10 15 20 25 30

Germany

Spain

Greece

France

Portugal

Italy

Austria

Netherlands

# of plants

(b) Employed technology

DEC Systems28%

Adsorption chillers

12%

Absorption chillers

60%

Figure 2.1.8: Installations (a) and distribution by technology (b) of solar cooling systems in Europe.



Presently, in the EU-15 countries solar energy contribution to air-conditioning in buildings is really modest and is principally characterized by demonstration plants. Spain and Germany account for about 70% of the total installations in Europe, as can be noted in Figure 2.1.8. The same figure illustrates the distribution of solar cooling systems by technology, in which thermally driven absorption chillers play the role of leading technology, followed by desiccant air conditioning systems (DEC) using solid or liquid sorbent materials. _________________________________ Finally, some interesting remarks about the current trend in national markets are summarized in the following table (for more details refer to Annex A). COUNTRY Some information about market trend

AUSTRIA In recent years there is a great attention to realize new products more environmental friendly, more energy efficient, reliable and robust. Concerning existing technologies (mainly compression chiller), they should be improved for reduction of the energy consumption, reduction of refrigerant leakages and volume per cooling capacity, improvement of “Life Cycle Climate Performance”. With regards to R&D actions for technology beside compression chiller, there is an interest in ab/adsorption cooling machines driven by district heating network.

GERMANY In recent years there is a growing interest for thermal driven chillers. Global trend for Germany shows a huge increase of air-conditioned areas. Concerning construction trend, an increase of split systems to more than 50% at the expense of single duct systems.

GREECE There are no significant changes expected with regard to the present situation. Nevertheless, it has to be expected an increase of penetration of thermal driven systems (due to natural gas penetration), and the domination of inverter type systems. As regards the market, a gradual but constant increase of Chinese low-cost products penetration has to be pointed out.

ITALY A continuous market growth has to be expected for small compression chillers (up to 40% for room air conditioners and in particular for split and multi-split units). Conversely, a small reduction of sales has been observed, instead, for large scale systems. Most recently, there is a significant increase in the interest towards thermally driven chillers.

SPAIN The market of inverter type systems is increasing in spite of these systems are more expensive, but the energy consumption is lesser. Due to high water consumption and to the possible contamination of “legionella bacteria” in the cooling towers, the market of water condensed chillers is decreasing in favour of air condensed chillers.

SWEDEN

Increased use of air-air heat pumps in domestic and small office locations. Increased installation of comfort cooling in commercial and public buildings due to perceived warmer summers and increased internal loads. District cooling has increased from 180 to 600 GWh cooling in the period 1998-2002. The estimated market for district cooling is 2 TWhcooling by 2010. Increased interest in providing local cooling using thermally driven cooling processes, with heat supplied from district heating.

Table 2.1.2: Market trends in some EU countries (Source: country data delivered by NEGST WP5 participants)

2.2 ABSORPTION CHILLERS

Most thermally driven cooling systems and solar assisted air conditioning systems installed today are based on absorption chillers /Hen02/. Flat plate collectors, vacuum tubes, as well as concentrating collectors are used for the heat supply of solar assisted air conditioning systems. Absorption chillers are using heat to drive a refrigerant. A working pair, i.e. the refrigerant and a solvent are used to drive the process in two loops.



Absorption chillers work either in a continuous or intermittent mode. For solar driven absorption chillers continuously driven cycles are most suited. An overview of solar assisted absorption chillers is given, for example in /Hen04c/, /Hen04a/ and /Hen04b/.

2.2.1 Physical principles

Figure 2.2.1 shows schematic of the working principle of absorption chillers. Similar to a compression cooling machine, the main loop, shown in blue, is composed of an evaporator, a compressor, a condenser, and a throttle-valve. In the evaporator heat is extracted from the

room to be chilled ( coolingQ& ) and transferred to the refrigerant on a low pressure level. In the

thermal compressor (Figure 2.2.1, green colour) the evaporated refrigerant is compressed by an absorption/desorption cycle, using heat e.g from a solar collector. In the condenser heat is removed by a heat sink ( 1AQ& ), e.g. a cooling tower. Afterwards, the pressure of the refrigerant condensate is reduced in the expansion valve and flows back to the evaporator. Instead of a mechanical compressor used in compression cooling machines, a thermal compression cycle is used (Figure 2.2.1). It consists of an absorber, a pump, a heat exchanger, a generator, and a second throttle valve. A concentrated hygroscopic fluid absorbs the refrigerant vapour in the absorber. Heat released during the exothermic absorption process

( 2AQ& ) is transferred to a heat sink, i.e. a cooling tower.

The mixture of the two fluids, the diluted solution, is pumped from the absorber to the regenerator. In the regenerator the mixture is separated again by increasing the pressure due to heat supply, e.g. from a solar collector. Due to the fact that the boiling point of the mixture is higher than the boiling point of the pure hygroscopic solution, the refrigerant vapour can be released at high pressure to flow to the condenser. The concentrated hygroscopic solution returns to the absorber. Heat recovery is applied from the hot concentrated solution to the diluted solution in a heat exchanger. The positions of the components shown in Figure 2.2.1 indicate the pressure and temperature levels of the fluids in the system.

2.2.1.1 Advanced cycles: AHE (Absorber Heat Exchange) and GAX (Generator-Absorber heat eXchange)

According to Engler et al. (/Eng97/) and Corallo et al. (/Cor03/), additional heat exchangers can be installed in the absorber and generator for heat recovery purposes, in order to increase the performance of absorption chillers. As shown in Figure 2.2.2 a) and b), the heat recovery can be realized either with internal loops inside both, the absorber and the generator, so-called Absorber Heat Exchange (AHE), or with an additional loop between the absorber and regenerator (Generator-Absorber heat eXchange, GAX).

condenser

evaporator

throttle valve

pump

thermal compressor

solutionheat exchanger

cooling

driveA1

throttle valvefor the solvent

A2Q Q

QQ

generatorpres

sure

temperature

absorber

Figure 2.2.1: Absorption chillers, physical principle

(Source: /Hen04a/ (changed)).



condenser

evaporator

cooling

driveA1

A2Q Q

QQ

gene-rator

ab-sorber

(a)

P

CONDENSER

RECTIFIER

EVAPORATOR ABSORBER

GENERATOR

EXPANSION VALVE

PRECOOLER

SOLUTION VAPOUR REFRIGERANT LIQUID REFRIGERANT

(b)

Figure 2.2.2: (a) Absorber Heat Exchange (AHE) (b) Generator-Absorber heat eXchange (GAX).

2.2.1.2 Diffusion Absorption Chillers

In diffusion absorption chillers a gas bubble pump is used instead of the solution pump. During heating, gas bubbles are produced which carry the rising solution along to the regenerator. In opposite to the absorption chillers described in the last section, there is a uniform pressure all over the diffusion absorption chillers. The density difference in the absorber and evaporator is realized by a pressure equalizing auxiliary gas (usually Helium). The main advantages of absorption chillers compared to conventional compression cooling machines are their low electrical energy consumption and maintenance needs. On the other hand, variable operating conditions (temperatures, etc.) have a large impact on the reliability of the systems.

2.2.2 Classification of absorption chillers

Absorption chillers can be classified by their working pairs, the number of stages the chillers consist of (single, double, triple effect), and according to developments obtainable from coupling the absorption chillers with solar collectors of advanced cycles.

2.2.2.1 Working Pairs

The working pair consists of the refrigerant and a hygroscopic solvent. The fluid materials used depend on:

� the chilling set temperature; � the temperature supplied in the generator; � the thermodynamic properties of the fluids relevant for the chilling process.

Suitable working pairs are chosen according to the absorption capability of the solvent and the capability to release as little heat as possible in the absorption process. Furthermore, the absorption capability of the solvent shall only marginally depend on temperature and the solution shall be liquid after the absorption of the refrigerant. In order to guarantee a sufficient separation of the refrigerant and the solvent in the generator, the solvent shall have a distinctly higher boiling temperature at generator pressure than the refrigerant.



So far, absorption chillers operating with the working pairs H2O/LiBr und NH3/H2O have been proven to be reliable. In order to achieve temperatures below 4°C the working pair NH 3/H2O is used, whereas for air conditioning purposes mostly H2O/LiBr is used. Moreover, research and development projects are being carried out using NH3/H2O/He for diffusion absorption chillers. Prototypes of these systems are being developed, for example, by Jakob et al. /Jak03/. The working pair H2O/LiBr (refrigerant/solvent) is used for air conditioning and water chilling down to a temperature of solely 4°C, in order to pr event the refrigerant to freeze. The operating pressure of the refrigerant is typically about 0.01 bar (evaporator, absorber) to 0.1 bar (generator, condenser). Crystallization of LiBr that occurs at high concentrations must be prevented. With the working pair NH3/H2O (refrigerant/solvent) cooling temperatures down to about –20°C can be generated. In order to achieve these low temperatures, supply temperatures of 120 to 150°C are necessary in the regenerator. The working pressure of the refrigerant is typically about 2 bar (evaporator, absorber) to 10 bar (generator, condenser). At a temperature in the regenerator of more than 120°C and at such a pressure the solvent (water) evaporates as well. In this case it must be separated from the Ammonium by means of a rectification column (dephlegmator).

2.2.2.2 Number of cycles

Absorption chillers are built with either one, two or three cycles as shown in Figure 2.2.3. Additional generators and condensers are used on different temperature and pressure levels, in order to increase the amount of refrigerant to be evaporated in the regenerators and therewith the COP of the system. The heat removed from the condenser of the higher cycle is sufficient to be used in the lower-temperature regenerator. Double and triple effect absorption chillers are usually less suitable for solar energy assisted absorption chillers since higher working temperatures are needed. Parabolic trough collectors are applied for double effect absorption chillers recently, developed by the company Solitem (Austria).

ABSOR BEREVAPOR ATOR

GENERATORCONDENSOR

(a)

ABSORBEREVAPORATOR

LOW-TEMPERATURGENERAT OR

MIDDLE-TEMPERATURGENERAT OR

LOW-TEMPERATURCONDE NSOR

MIDDLE-TE MPERATURCONDENSOR

(b)

ABSORBEREVAPORATOR

LOW-TEMPERATURGENERAT OR

MIDDLE-TEMPERATURGENERAT OR

HIGH-TEMPERATURGENERATOR

LOW-TEMPERATURCONDE NSOR

MIDDLE-TE MPERATURCONDENSOR

HIGH-TEMPERATURCONDENSOR

(c)

Figure 2.2.3: Single-effect (a), double-effect (b), and triple-effect (c) absorption chillers (Source: Southern California Gas Company /Mah98/).



2.2.3 Cycle Performance and Maintenance

2.2.3.1 Cycle Performance

The coefficient of performance for absorption chillers is defined as the useful cooling power related to the heat necessary to drive the process. An overview of COP values, the suitable driving temperature range, and the range of cooling capacity is given in Table 2.2.1 for absorption chillers operated with one and two cycles with the working pairs LiBr/H2O and H2O/NH3.

Number of cycles 1 2 1 Solvent LiBr LiBr H2O Refrigerent H2O H2O NH3 Driving temperature 80°C - 110°C 140°C - 160°C 80°C - 120°C

Driven by hot water steam

hot water steam

directly burned

hot water steam

directly burned COP 0,6 - 0,8 0,9 - 1,2 0,3 - 0,7

Power range market available

few producers 20 to 100 kW

many producers more than 100 kW

few producers 50 to 100 kW

more producers more than 100 kW

*********

Producer

Broad, Carrier, Century, EAW, Ebara, Enropie, LG Machinery, Sanyo-McQuay,

Sulzer-Escher Wyss, Trane, Dunham-Bush, Yazaki, York

ABB, Colibri, Mattes, Robur

Table 2.2.1: Absorption chillers. Table by /Hen04a/ (modified, producers added).

If additional heat exchangers are included in both, the absorber and generator (Absorber Heat Exchange, AHE cycle, section 2.2.1.1) of single effect LiBr/H2O absorption chillers, higher driving temperatures Tdrive of at least 150°C are required compared to the bas ic cycle. Meanwhile, the COP increases from 0.6 (for the basic cycle) to about 0.75. For GAX cycles (Generator-Absorber heat eXchange), a minimum driving temperature of 160°C is necessary, to generate a COP of about 0.75. The COP increases rapidly to about 1, when the driving temperature approaches 200°C.

Ammonia/Water Advanced Cycles Parameter AHE GAX

Typical size 10-90 KW 10-90 KW COP 0.7 – 0.75 0.8 – 0.9 Driving temperature 150 – 200°C 160 – 200°C Table 2.2.2: Advanced cycles: Absorber Heat Exchange (AHE) and

Generator-Absorber heat exchange (GAX). Typical sizes and COP.

2.2.3.2 Maintenance and lifespan

The system pressure and the chemical stability of solvent or refrigerant need to be inspected at least once or twice a year. Due to incomplete sealing of the system, the solvent degrades gradually and expanded maintenance and analysis of the solvent has to be carried out about every 5000 operation hours. Moreover, for operating temperatures of more than 100°C special inspections are mandatory e.g. in Germany (TÜV). For H2O/LiBr absorption chillers the pressure needs to be checked daily or weekly. Corrosion protection needs to be checked at the beginning of operation.



The lifespan of absorption chillers is about 15 to 25 years. Absorption chillers with the working pair H2O/LiBr tend to have a longer lifespan than systems with a NH3/H2O working pair.

2.2.3.3 Typical Sizes

Most single effect absorption chillers available have a cooling capacity of more than 100 kW. Only few products are on the market with a smaller cooling power. Companies offering absorption chillers are predominantly from USA and Asia (Japan, Korea, China, India). Among small absorption chillers the system WFC 10 by Yazaki (35 kW) has the largest market share. Moreover, small single effect LiBr systems are supplied by the companies Broad Air (20kW), Robur Corporation (17kW - 88kW) and by EAW (Westenfeld, 15KW). Additionally, absorption chillers with a fairly low cooling power are being developed by small companies, for example Phönix (Germany) and SolarFrost (Austria), as well as by a number of research institutes (University of Applied Science Stuttgart, Germany), Joanneum Research (Austria), INETI (Portugal), UPC (Spain), etc. There are few manufacturers of double effect absorption chillers. Most products are directly fired. Most two-stage absorption chillers are available with a cooling power of more than 100 kW. An overview of 36 absorption chillers installed in Europe is shown in section 2.2.5, Table 2.2.4 (source: Henning /Hen04b/ and http://www.ocp.tudelft.nl/ev/res/sace.htm). Location, application, cooling capacity, and solar collector type are listed for each system. The mean cooling capacity of these systems is about 87 kW, and the mean collector area installed is about 250 m², which corresponds to a specific collector area of about 2.9 m² per kW cooling power. However, the data are not completely comparable, since the collector area is not defined identically, some of the plants are used as heating systems as well, and some of these absorption chillers are stand-alone, others are solar assisted systems /Hen04b/.

2.2.3.4 Advantages and Drawbacks with Respect to Conventional Technologies

According to Henning et al. /Hen02/ the main problems and expected developments in order to achieve a further penetration of absorption chillers in solar-assisted air conditioning systems are the following:

� The absorption chillers on the market are mainly intended for large-scale applications. Nevertheless, there is a demand for smaller solar-assisted air-conditioning systems.

� For LiBr absorption chillers a cooling tower is needed, which leads to increasing investment costs.

� For low driving temperatures only small efficiencies and capacities can be achieved. � More expensive collector types (e.g. vacuum tubes, CPCs) are required to guarantee

sufficient efficiencies. Nevertheless, it has been shown in the projects listed in table 4 that absorption chillers have been operating successfully and are established on the market.

2.2.4 Coupling of the Absorption Chillers with a Solar Heating System

Solar energy driven absorption chillers can be installed either with an auxiliary energy source (solar assisted system) or without a back up system (stand-alone system). A schematic of absorption chillers coupled with a solar heating system and auxiliary energy supply is shown in Figure 2.2.4. The solar heating system consists of solar collectors, a storage tank, hydraulics, and a back-up heating supply system. The total coefficient of performance (COP) of solar assisted absorption chillers is defined as:

collchillersol COPCOP η⋅= (2.2.1)



C

EA

GB

CT

HS

CS

ABC

SC

A absorber ABC absorption chillers B thermal backup system C condenser CS cold storage CT cooling tower E evaporator G generator HS heat storage SC solar collector

Figure 2.2.4: Coupling of an absorption chillers with a solar heating system.

As shown in Figure 2.2.5, the higher the supply temperature delivered by the solar collector, the lower the collector efficiency and the higher the COP of the absorption chillers. Thus, high efficiency collectors even at high temperatures are decisive for an economic and effective operation of absorption chillers. Collector types suitable for single and double-effect absorption chillers are shown in Figure 2.2.6. In addition to the driving temperature the overall system performance depends on the reference conditions like solar irradiation available and desired cooling temperature and demand.

0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

50 60 70 80 90 100 110 120TG [°C]

ηη ηη co

llect

or ,

ηη ηη

0

0,1

0,2

0,3

0,4

0,5

0,6

0,7

0,8

CO

P

00,10,20,30,40,50,60,70,80,9

1

0 0,05 0,1 0,15 0,2 0,25 0,3 0,35

(T f -Tamb ) / G

ηη ηη col

lect

or

CPC stationary compound parabolic concentrator

EDF direct-contact evacuated tube collector

EHP heat-pipe evacuated tube collector

FPC flat-plate collector

SAC solar air collector

SYC evacuated tube collector (Sydney type)

Figure 2.2.5: Collector efficiency, COP and the

resulting overall efficiency of the system over the temperature in the regenerator.

Figure 2.2.6: Collector efficiency curves of various collector types, source: Henning et al. /Hen04d/ (changed).

2.2.4.1 Solar Collector

According to Henning et al. /Hen04c/ for stand-alone sorption chillers the specific collector area, defined as the collector area per nominal cooling capacity, was roughly chosen according to the following equation (rule of thumb):



COPG

1²m/A

collectorspec,coll ⋅η⋅

=⊥ (2.2.2)

with:

G⊥ global irradiation; ηcollector collector efficiency at design conditions; COP coefficient of performance.

Thus, for G⊥ = 800 W/m², ηcollector = 50%, and COP = 0.7, the specific design collector area results to Acoll,spec = 3.6 m² for 1 kW cooling capacity. While the specific collector area for sorption chillers varies between about 1 and 6 m² per kW of installed cooling capacity, in average 2.9 m² of collector area per kW were installed for the absorption chillers listed in Table 2.2.4).

2.2.4.2 Storage Tank

In order to store surplus energy in periods with high solar irradiation, storage tanks can be installed. They can either store surplus heat or cold, depending on the application and reference conditions, like additional solar heat supply (e.g. for domestic hot water, space heating), accepted temperature tolerances, solar fraction, performance of the absorption chillers, storage characteristics (size, insulation,..), supply temperatures, etc. The required storage volume per kWh stored cold is shown in Figure 2.2.7

over the useful temperature difference. Cold can be stored as Eutectic salts and water, ice, or chilled water. The working principle, advantages and disadvantages of these three storage methods are listed in Table 2.2.3.

Method Functioning Advantage Disadvantage

Eutectic salts / H2O

Eutectic salts are a combination of inorganic salts. A mixture with water and few other elements make freezing possible at the desired temperature (typically near 8°C). The mixture is encapsulated and placed in a water store.

Small stores, small heat losses.

Storage temperatures of 8 – 10°C, only suitable for air conditioning, the stores are still being developed.

Ice store Ice is stored in the tank.

Small stores, 10..20% of the size compared to a cold water store, 30..50% compared to stores with eutectic salts; for direct air flow drying is not necessary.

Low temperatures are needed.

Cold water store

Cold water is stored with typical temperatures of about 6°C.

Inexpensive, no critical chemicals

Large storage sizes.

Table 2.2.3: Characteristics of cold stores.

chilledwater

hot water COP 0,5

hot water COP 0,8

0,000,020,040,060,080,100,120,140,160,180,20

5 10 15 20 25 30 35 40

useful temperature difference [K]

requ

ired

stor

age

volu

me

[m3 /k

Wh

cold

]

Figure 2.2.7: Required storage volume per kWh cold (Henning

et al. [2]).



2.2.4.3 Back-up System

In order to guarantee the functioning of absorption chillers also in times of low irradiation, back-up heating or cooling systems are installed. For back-up cooling usually conventional compression cooling machines are used. These are connected to the cooling loop of the absorption chillers. Oil, gas, or pellet burners are usually used as back-up heating devices. The back-up heating system can either be connected to a heat store or it can be installed between the heat store and the absorption chillers. For both, cooling and heating back-up systems, modulation is decisive in order to insure a high overall efficiency of the cooling system.

2.2.5 Investment costs

2.2.5.1 Collector Area

A rough estimation of the investment costs for the collector area can be calculated from the collector equation (equation 2.2.3, Duffie and Beckman /Duf91/) and the specific collector price. The heat supplied by the solar collector per m² collector area is described as a function of the optical and thermal losses:

)²TT(k)TT(k)(GA

Qambav2ambav1

coll

coll −−−−τα⋅=&

(2.2.3)

For typical reference conditions and a heat delivery of collQ& = 1KW, the required collector area per kW heat delivery Acoll,1kW can be calculated. The costs for the solar collector for 1kW heat delivery for typical reference conditions is then calculated by the product of Acoll,1kW (in m²/KW) and the specific collector price (in €/m²).

In Figure 2.2.8 (Source: Henning et al. /Hen04c/) investment costs for the solar collector area of various collector types per kW cooling capacity are plotted against the required collector area per kW cooling capacity for different collector parameters and the following reference conditions: ambient temperature: Tamb = 20°C, average fluid temperature in the collector Tav= 95°C, solar irradiation incident on the collector area G = 800 W/m², and incident angle modifier of 1. The resulting specific investment costs range from about 400 €/kW for flat plate collectors to about 2000 €/kW for evacuated tube collectors.

Specific collector areas between about 3 and 5 m²/kWheat and about 2 m²/kWheat are taken into account for these calculations.

Figure 2.2.8: Investment costs over the required absorber area

(Source: Henning et al. /Hen04c/) for typical reference conditions of single-effect absorption chillers (Tamb = 20°C, Tav = 95°C, G = 800 W/m2).

ETC evacuated tube collector

FPC flat-plate collector



2.2.5.2 Absorption chillers

Specific costs for absorption chillers per kW cooling capacity with a COP of about 0.7 (driving temperatures 88°C/83°C) range from about 400 EUR/kW (for Pcooling ≈ 35 kW) and about 700 EUR/kW (for Pcooling ≈ 100 kW) (York).

2.2.6 Operating plants

Table 2.2.4 shows some solar assisted absorption plants operating in Europe.

Land Anlage / Ort Anwendung Leistung [kW]

Kollektortyp Kollektorfläche [m 2]

Wolfferts / Köln Büroräume 70 ETC / 176 Ott & Spies / Langenau Büroräume 35 ETC / 22 Bundespresseamt / Berlin Büroräume 70 ETC / 244

Fraunhofer-Institut Umsicht / Oberhausen Büroräume Laborräume

58 ETC / 108

Bundesverkehrsministerium / Berlin Kältenetz 70 FPC / 209

ZAE Bayern / Garching Büroräume und Laborräume

7 ETC / 20

M + W Zander / Stuttgart Büroräume 143 ETC / 260

Deutschland

Technologiezentrum / Köthen Büroräume 15 ETC / 79 American College I / Athen Ausbildungsräume 168 ETC / 615

Solar Lab Demokritos / Athen Büroräume und Laborräume 35 FPC / 160

Rethymno Village Hotel / Rethymno Hotel 105 FPC / 450 Grichenland

Lentzakis Crete / Rethymno Hotel 105 FPC / 450 Social and Cultural Centre Clara Campoamor / Barakaldo

Auditorium, Ausstellungsraum 229 FPC / 150

Education Department Regional Goverment / Toledo

Büroräume 252 ETC / 750

Fabrica de Sol Building / Barcelona Büroräume 105 ETC / 120 Fundación Metrópoli Building / Madrid Büroräume 105 ETC / 72 Daoiz y Velarde Sports Centre / Madrid Sportcenter 170 ETC / 507 Head Offices of Inditex / Aretixo (La Coruna) Büroräume 170 FPC / 1500 Old Peoples Home / Fustinana (Navarra) verschiedenes 105 ETC / 102 University Rovira i Virgili / Tarragona Büroräume 35 ETC / 96 Head Offices of Viessmann Spain / Pinto (Madrid)

Büroräume 105 FPC / 105 ETC / 6

Belroy Palace Hotel / Benidorm (Alicante) Hotel 125 ETC / 345 University of Sevilla (School of Engineers) / Sevilla Laborräume 35 FPC / 151

University Carlos III / Leganés (Madrid) Laborräume 35 ÊTC / 50 FPC / 50

Laia Hotel / Derio Hotel 105 FPC / 160

CARTIF, Boecillo Technology Park / Valladolid Büroräume Laborräume 35 FPC / 37,5

ETC / 40 Siemens Contromatic / Corneliá del Vallés Büroräume 105 CPC / 214 National Institute of Airospecial Techniques INTA / Huelva

Laborräume 10 FPC / 25 ETC / 18

FONTEDOSO / El Oso Industriegebäude 105 FPC / 504

Spanien

Stella-Feuga Building, Santiago de Compostela Büroräume 115 FPC / 60

Portugal Verkehrsleitzentrale Büroräume und Leitzentrale

70 CPC / 663

Italien Baxter / Trento Büroräume, Ausstellungsfläche

108 FPC / 108

Österreich Weinbetrieb Peitler / Steiermark Weinlager 10 FPC / 100 CSTB-Gebäude/ Sophia Antipolis Laborräume 35 ETC / 63 DIREN-Gebäude / Guadeloupe Büroräume 35 ETC / 61 Frankreich Weinkeller / Banuyls Weinlager 52 ETC / 130

Table 2.2.4: Solar assisted absorption chillers installed in Europe.(Source: Henning /Hen04b/).



2.3 ADSORPTION CHILLERS


Adsorption means the binding of molecules or particles to a surface. Contrary to absorption the adsorption process represents a physical process in which the molecules of one substance are adsorbed on the internal surface of another substance. The binding to the surface is weak and reversible. Suitable adsorbents are porous materials that are insoluble in water, which have enormous surface areas per unit weight where they can physically bind water or molecules. Typical substances used are silica gel, activated carbon or alumina. During the adsorption process condensation heat is released (exothermic reaction). The reverse process is called desorption, for which evaporation heat must be applied.

2.3.2 Classification of adsorption chillers

Concerning solar cooling it can be distinguished between closed and open sorption-chilling processes. Application areas for closed systems are the production of cold water which is either used in central ventilation stations (dehumidification) or for decentral air conditioning e.g. the

cooling of building elements. In the following the functional principle of adsorption chillers, which belong to the group of closed systems, is explained. An adsorption chiller consists of two separate chambers, an evaporator and a condenser (Figure 2.3.1). Each of the chambers contains the adsorbent (e.g. silica-gel) and a heat exchanger. The main difference to a general chiller is that the solid sorbent cannot be circulated. The adsorption chiller works discontinuously. In one phase the adsorption process is linked with the evaporation and in another phase the

desorption process is linked with condensation. During one cooling cycle the following processes take place:

� The refrigerant adsorbed is driven off through the use of firing water in the desorber (right chamber).

� The refrigerant condenses in the condenser and heat of condensation is removed by cooling water.

� The condensate is sprayed in the evaporator, and evaporates under low partial pressure. This step produces the useful cooling effect. Heat is driven off from the chilled water which is cooled to the required temperature.

� The refrigerant vapour is adsorbed in the adsorber (left chamber). Heat is removed by cooling water.

Once the adsorber is charged and the desorber regenerated, their functions are interchanged. Solar energy is used for the regeneration of the absorbent: regeneration is done by passing hot water through the chamber that is heated by a solar system. Typical cooling powers are between 70 – 400 kW. A characteristic performace parameter of adsorption chillers is the coefficient of performance (COP). This is the quotient of heat transferred in the evaporator to the heat required for regeneration, which is delivered by the solar collectors (electrical energy use is not considered). In general the COP of adsorption chillers range from 0.5 to 0.7.

cooling water condenser

evaporator

adsorber desorber

cooling water

firing water

chilled water

50…90 °C

25…35 °C

25…35 °C

5…12 °C

Figure 2.4.1: Schematic of the internal chambers of an

adsorption chiller.



There are only two manufacturer of this kind of adsorption chillers worldwide. Both are Japanese manufacturers: Nishiyodo (adsorption chiller type: NAK) and Mayekawa (adsorption chiller type: ADR). In the following the adsorption chillers are classified according to their working pair (cooling agent/absorbent).

2.3.2.1 Water/silica-gel

This adsorption chiller uses water as its cooling agent. Water evaporates in a vacuum at room temperature and because of that it extracts heat from the surroundings (evaporation energy). A cooling of the “chilled water” takes place because of this process. Compared to open systems the evaporated water is not released as steam into the surroundings, but re-condensed within the chiller. The adsorption chiller is a closed system. A direct condensation of the evaporated water is energetically not possible for thermodynamic reasons (temperature of the evaporated water is lower than temperature level in the condenser). Therefore, first the water is adsorbed by a solid carrier material, the adsorbent. In this case silica-gel, a material related to quartz or sand, is used as adsorbent. As the adsorption process is an exothermic reaction, condensation heat is released during the adsorption process. It has to be ensured that this heat is driven off the chamber through cooling water, as the maximum degree of possible adsorption depends on the temperature level of the adsorbent as well as on the pressure level in the chamber. The higher the temperature of the adsorbent, the lower is the relation between adsorbed water mass [kgH20] and mass of the adsorbent (silica-gel) [kgads]. That means that with higher adsorbent temperature less water vapour can be adsorbed until saturation is reached. With the firing water the adsorbed water vapour on the silica-gel (carrier material) is desorbed again and the silica-gel is regenerated. The desorbed water vapour can now directly be condensed in the condenser, as the resulting temperature level in the right chamber (desorption chamber) is higher than the temperature level in the condenser. The following factors are essential for the process:

� Silica-gel can easily take up water without causing a structural change or volume expansion.

� It can release easily the stored water due to a temperature increase. This process is reversible and unlimitedly repeatable.

� The evaporation process depends on temperature and pressure. Under common atmospheric pressure water evaporates at 100 °C. If the pressure drops, the evaporating temperature of the water also decreases.

� By creation of a sufficient vacuum the water evaporates at a lower temperature. For the purpose in the adsorption chiller, a vacuum of 13 – 26 mbar is sufficient.

� If water is sprayed or injected into a vessel under vacuum, it evaporates spontaneously and extracts energy from the surroundings.

2.4.2.2 Ammonia/activated carbon

The disadvantage using water as cooling agent is that it is not a suitable refrigerant for sub-zero temperature application. For applications below 0 °C ammonia NH3 can be used as an appropriate cooling agent. Activated carbon is produced by roasting organic material to decompose it to granules of carbon – coconut shell, wood, and bone are common sources. Spent activated carbon is regenerated by roasting but the thermal expansion and contraction eventually disintegrate the structure so some carbon is lost or oxidized. But the ammonia/activated carbon pair requires high temperature (>120 °C) input heat for regeneration.



2.3.2.2 Methanol/silica-gel

The disadvantage using water as cooling agent is that it is not a suitable refrigerant for sub-zero temperature application. For applications below 0 °C methanol can be used as an appropriate cooling agent. Silica gel is a matrix of hydrated silicon dioxide. The advantage of using methanol/silicagel as adsorbate/adsorbent pair is that it may get activated even at a temperature of 60 - 70 °C.


Until now only few plants are realised because adsorption chillers show some disadvantages like heavy weight, big volume and limited market choice as there are existing only two manufacturers of adsorption chillers up to now. Another reason is that adsorption chillers are more expensive than absorption chillers. The big volume of adsorption chillers results out of the necessity to use large heat exchangers, as the heat conductivity of silica-gel is very low. The following will show an extract of some plants that realize solar cooling with adsorption chillers.

2.3.3.1 University hospital in Freiburg/Germany

The university hospital of Freiburg operates a solar assisted chilled water production using an adsorption chiller. The cold water is used for air conditioning of a laboratory building. The adsorption chiller used is type GBU NAK 70 with a chilling capacity of 70 kW. As solar collectors, 170 m² vacuum tube collectors from Seido are used.

2.3.3.2 Storage of agricultural products in Haipur/India

This solar-hybrid adsorption cooling system for decentralized storage of agricultural products in India was jointly designed by TERI (Tata Energy Research Institute) and DLR (German Aerospace Establishment). As working pair methanol/silicagel has been used in the adsorption refrigeration unit, as sub-zero temperatures have to be produced. As solar collectors, 25 m² CPC collectors are used. The adsorption cooling system is applied for the decentralized cold storage of agricultural products in India where electricity is not available. The system has only a COP of about 0.3. This is for the reason that the evaporation enthalpy of methanol is lower than that of water and that the temperature level in the evaporator is lower than it is at climatic applications.

2.3.3.3 Hospital in Kamenz, Dresden/Germany

The system is a combination of fuel cell, solar system and an adsorption chiller with a chilling capacity of 105 kW (type: Mycom ADR 30 from Mayekawa). As working pair water/silicagel is used. The system is realised with a collector area of 115 m².

2.3.3.4 Others

� Sarantis cosmetic factory, Aharnes/Greece � Office „An der Loge“, Dresden/Germany (Nyshiodo NAK 70; water/silica-gel; 70 kW; 156

m² flat plate collectors) � Office building Remscheid/Germany (MYCOM ADR 30; water/silica-gel; 105 kW; 150 m²

collector area) � Office building Bremen/Germany (MYCOM ADR 15; water/silica-gel; 50 kW; 120 m²

collector area)



� University Dortmund/Germany (MYCOM ADR 30; water/silica-gel;105 kW) � Office building Augsburg/Germany (Nyshiodo NAK 100; water/silica-gel; 350 kW;

collector area 2000 m² flat plate collectors) � Office building Würzburg/Germany (Nyshiodo NAK 70; water/silica-gel; 70 kW; collector

area 60 m² flat plate collectors) � Cuernavaca/Mexico (MYCOM ADR 50; water/silica-gel; 176 kW)

2.4 THERMO CHEMICAL ACCUMULATOR (TCA)


The thermo-chemical accumulator (TCA) is an absorption process that uses a working pair, not only in the liquid, vapour and solution phases but also with solid sorbent /Ols00/, and was patented in 2000. This makes it significantly different from the traditional absorption processes in that it is a three phase process (solid, solution and vapour). All other absorption processes are two phase processes with either solution + vapour or solid + vapour. Figure 2.4.1 shows the schematic of a single TCA unit, where the solution is pumped over a heat exchanger via a spreader arm to increase the wetted area and improve heat transfer. The process has been developed by the Swedish company ClimateWell AB.

During desorption the solution comes closer and closer to saturation, and when it reaches saturation point further desorption at the heat exchanger results in the formation of solid crystals that fall under gravity into the reactor vessel. Here they are prevented from following the solution into the pump by a sieve, thus creating a form of slurry in the bottom of the vessel. This gives the TCA the following characteristics:

� High energy density storage in the solid crystals.

� Good heat and mass transfer, as this occurs with solution.

� Constant operating conditions, with constant temperature difference between reactor and condensor/evaporator.

For discharging, where the process is reversed, saturated solution is pumped over the heat exchanger where it absorbs the vapour evaporated in the evaporator. The heat of evaporation is provided either by the building (cooling mode) or from the environment (heating mode). The solution becomes unsaturated on the heat exchanger, but when it falls into the vessel it has to pass through the slurry of crystals, where some of the crystals are dissolved to make the solution fully saturated again. In this way the solution is always saturated and the net result is a dissolving of the crystals into saturated solution. The heat of condensation and binding energy release is transferred to the environment (cooling mode) or to the building (heating mode). Thus there is a flow of energy from the evaporator at low temperature to the reactor at moderate temperature.

Figure 2.4.2: Schematic of a single unit thermo-

chemical accumulator. (Source: ClimateWell AB).



The first TCA units have been built using water/LiCl as the active pair. The physical properties of the working pair have been summarised in the literature /Con04/ and empirical equations have been created for them based on data from a large number of studies over the last 100 years. ∆Tequ, the maximum theoretical temperature lift between the reactor and condensor/evaporator, is constant for a given set of boundary conditions resulting in constant operating conditions, such as solution temperature (Tsol), during charging / discharging. However, ∆Tequ does vary considerably over the operating range of the machine. Figure 2.4.2 shows the relationship of ∆Tequ to the temperature of the saturated solution based on Conde’s

equations (solid line) and as measured by ClimateWell AB (filled squares). Note that ∆Tequ is the temperature difference between the evaporator and reactor and that the temperature difference between the liquids in the external circuits is greater than ∆Tequ for charging and smaller for discharging (both cooling and heating). This technology has been developed to the demonstration phase with extensive lab and field testing during 2004. Note that there is an inbuilt conflict between the energy density for heat storage and the COP for cooling. A higher binding energy for water to the salt will result in higher energy density storage, as well as higher COP for heat recovery, but lower COP for cooling.

2.4.2 Classification of TCA machines

Two different types of prototype TCA machines have been developed and tested. In prototype 2 (Figure 2.4.3) the main unit was used only for charging and the slave unit only for discharging. By pumping saturated solution and water from the main unit to the slave unit while pumping back dilute solution, semi-conitnuous operation was possible despite the fact that the process is by nature a batch process. This makes it similar to a standard single-effect LiBr absorption chiller. However, it has not been possible to include a solution heat exchanger as crystallisation occurs very easily there. This limits the potential COP for the machine. In prototype 7 (Figure 2.4.4) there are two identical units that work in batch mode, with one undergoing charging while the other can provide cooling.

0 20 40 60 80 10015

20

25

30

35

40

45

50

55

60

Tsol [°C]

∆∆∆∆T

equ

[°C]

Conde 2004

ClimateWell 2004

Figure 2.4.3: Relationship of ∆Tequ to the saturated solution

temperature.

Huvudkondensor/evaporator

Huvudreaktor Slavevaporator

Slavreaktor

Vatten tillslavevaporator

Substans tillslavreaktor

Substans frånslavreaktor

Pump

PumpPump

Pump

Main Condenser

Water to slave evaporator

Main Reactor

Slave Evaporator

Slave Reactor

TMC,in

TMCout

TMR,in

TMR,out TSE,in

TSE,out

TSR,in

TSRout

Figure 2.4.3: Schematic of TCA prototype 2.



When either the unit being charged is fully charged or the other unit is fully discharged, the units are swapped using a number of valves connected between the internal heat exchangers and the external circuits. The machines that have so far been made are best viewed as prototypes. There is no commercial production in 2004, although prototype machines have been sold to a number of parties. The design COP for cooling of the prototype 7 machine, called ClimateWell DB220, is 70% with no heat recovery. With 40% heat recovery during the swapping of units, a COP of 75% should be possible to achieve. The integral storage capability allows the machine to provide up to 50 kWh cooling if both units have been fully charged, although this figure is partially dependent on the cooling rate. The design COP for recovery of heat from the machine is 87%, but in heat pump mode a minimum temperature of 5°C is required for the evaporation of water. Costs are not available.

Reactor – Heat exchanger

Internal pumps

Tubes connected to the external circuits

Barrel

Salt filter basket

Condensor / evaporator – Heat exchanger

2000 mm

700 mm 700 mm

180 mm Connection unit

8 valves for controlling internal processes

Figure 2.4.4: Schematic of TCA prototype 7, ClimateWell DB220.

The main advantage of the technology is the integral heat storage that makes it very suitable for intermittent heat sources such as solar. In addition, the regeneration temperature is relatively low, being roughly 50-65°C above ambient, depending on the charging rate and ambient temperature. This is possible to achieve with flat plate collectors, although other types might be more economically viable in certain cases. The main disadvantages of the technology are the limited temperature lift during discharge. Here the performance is very dependent on the heat exchange with the ambient and with the air inside the building. It is not in general possible to provide dehumidification by having a low



delivery temperature, and a delivery system with 12/15°C or even higher is preferable. Another disadvantage is the cost of the LiCl in the machine. Other, less expensive, salts are being investigated. The reliability is still an open question as the technology is still in relatively early stage of development.


This is a new technology as stated previously. However, there are a number of solar cooling plants that have been built during 2004 with the prototypes. These are located in Sweden, Italy and Spain. Measurement data are not at present available. Prototypes have also been installed with heat supplied from district heating in Sweden.

2.5 DESICCANT COOLING FOR AIR-CONDITIONING

In general air-conditioning equipment has to fulfil following objectives: � Compensation of external loads (thermal transmission through the building envelope

and solar gain through window) and of internal loads (latent and sensible heat of persons, machines and other thermal heat sources)

� Dehumidification/ Humidification of supply air � Cooling/ Heating of supply air � Supply of fresh air according to hygienic needs

Traditional air-conditioning systems are more or less a combination of a ventilation system and a cooling device which is normally a conventional compression chiller. The ventilation system supplies the fresh air in accordance to the hygienic needs. The compression chiller provides chilled water to cool the supply air. The control of humidity and temperature of the supply air depends on the evaporator temperature of the compression cycle, i.e. dehumidification is realised by evaporator temperatures below the dew point of the supply air. Desiccant cooling systems (DEC-systems) operate as well as a fully air-conditioning unit; supplying fresh air and controlling humidity and temperature of supply air. DEC-systems use sorption based air dehumidification with the help of liquid or solid sorption materials and the evaporative cooling effect. Consequently the air-treatment in DEC-systems is based on two physical principles: dehumidification and evaporation. Accordingly the technical equipment of DEC-systems abandons totally a use of refrigerant medium with high potential of global warming.

1

2 3

4

5

678

9

heat recoverywheel

desiccantwheel

return airhumidifier

filter

supply airhumidifier

fan

return air

supply air

filterfanexhaust air

ambient air

heating coil

1

2 3

4

5

678

9

heat recoverywheel

desiccantwheel

return airhumidifier

filter

supply airhumidifier

fan

return air

supply air

filterfanexhaust air

ambient air

heating coil

Figure 2.5.1: General scheme of a desiccant cooling air-handling unit.



A standard desiccant cooling system consists on several different technical components which is shown in Figure 2.5.1. This most common DEC-system is generally separated by a supply air and a return air stream. Furthermore it is composed on standard components of air-conditioning units such as filters, fans, heat recovery, heating or cooling coils and humidifiers. In comparison to standard air-handling units the desiccant wheel is additionally implemented which dehumidifies the supply air to enlarge the potential of evaporative cooling. To guarantee a continuous operating air-conditioning process the desiccant material which is permanently charged with water molecules has to be discharged/ regenerated also constantly. The regeneration process of the desiccant material - whether liquid or solid – can be realised by providing regeneration heat. The required temperatures for an efficient regeneration of the desiccant wheel are in a range of 45°C up to 90°C. Due to this low driving temperatures economic advantages arise particularly for DEC-systems when it is coupled with district heating or heat supplied from a combined heat and power (CHP) plant. Of particular interest is the coupling with thermal solar energy.


Contrary to thermally driven chillers producing chilled water which can be supplied to any type of air-conditioning equipment the open desiccant cooling cycle produce directly conditioned air. Therefore the open air-conditioning process of a DEC-system can be described and explained by a psychometric chart for moist air. Representative air states which appear in standard desiccant cooling cycles using a rotating desiccant wheel are shown in Figure . DEC-Systems

according to this scheme are typically applied in moderate climates. The supply air fan sucks ambient air into the DEC-system passing primarily a filter unit. The filtered ambient air is dehumidified by the rotating desiccant wheel (2). The dehumidification is based on sorption which is an exothermal process effecting an air temperature increase. In Figure 2.5.2 the psychometric chart for moist air the thermo dynamical process is approximately an adiabatic one. This warm and dry air is pre-cooled by the rotation heat recovery wheel (3). The cooling potential is enabled by using the return air humidification. That means the heat of the sorption process which causes a temperature decrease of the supply air is transferred to the return air. To provide the required supply air state, the supply air humidifier can be used to control humidity and temperature (4). Following the thermo dynamical principle the temperature and the humidity can not be controlled separately. There is a strict dependency between temperature and humidity due to the adiabatic humidification process. In practise there is another temperature increase of about 1 Kelvin to the rejected heat of the electrical driven motor of the fan. The supply air fan sucks ambient air into the DEC-system passing primarily a filter unit. The filtered

0

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20

30

40

50

60

70

80

0 2 4 6 8 10 12 14 16 18 20 22

1

2

3

4

6

5

7

8

9

ϕ = 100 %

ϕ = 50 %

ϕ = 30 %

ϕ = 20 %

ϕ = 10 %

ϕ = 5 %

Figure 2.5.2: Psychometric chart for moist air

showing the state changes for a desiccant cooling process (Source: arsenal research/Austria)



ambient air is dehumidified by the rotating desiccant wheel (2). The dehumidification is based on sorption which is an exothermal process effecting an air temperature increase. In Figure 2.5.1 this fact is not illustrated. A proper design of the fan is recommended in such a way that the heat added to the supply air is minimised. The internal and external loads of the air-conditioned room normally cause an increase of temperature and humidity (5). The return air passes the second humidifier of the DEC-system which is almost humidified very close to the saturation point (6). This so called evaporative cooling of the exhaust air is an indirect method to pre-cool the supply air. The heat recovery wheel only transfers sensible heat between exhaust air and supply air. Regarding the overall performance of the DEC-system especially the heat recovery efficiency has to be as high as possible. The heat recovery from (6) to (7) leads to a temperature increase. In the following process step the exhaust air is heated by a heating coil up to the regeneration temperature (8). The provided regeneration heat effects the desorption process, i.e. the water molecules bound in the pores of the desiccant material of the sorption wheel is desorbed by means of the hot air (9). To enforce the required flow rate the exhaust air fan is finally implemented. The DEC-process can be summarised as follows: 1 → 2 sorptive dehumidification of supply air; the process is almost adiabatic and the air is

heated by the adsorption heat and the warmed wheel matrix coming from the regeneration side

2 → 3 pre-cooling of the supply air in counter-flow to the return air from the building 3 → 4 evaporative cooling of the supply air to the desired supply air humidity by means of a

humidifier 4 → 5 supply air temperature and humidity are increased by means of internal an external

loads 5 → 6 return air from the building is cooled using evaporative cooling close to the saturation 6 → 7 the return air is pre-heated in counter-flow to the supply air by means of a high

efficient air-to-air heat exchanger, e.g. a heat recovery wheel 7 → 8 regeneration heat is powered by a heating coil; this heating coil is driven by hot

water; for instance by hot water generated by solar thermal collectors 8 → 9 regeneration process of the desiccant material; the water bound by pores of the

desiccant material of the sorption wheel is desorbed by means of regeneration air

2.5.2 Coefficient of performance

The definition of the Coefficient of Performance COP for open air-conditioning methods is based on a thermal energy balance on the system. The COP - we consider - is generally spoken the quotient of the thermal cooling output and the driving heat input. In difference to the definition of efficiency the COP can be greater than 1. The thermal cooling output or use is the enthalpy difference between ambient air and room inlet air. The driving heat input of the DEC-system corresponds to the enthalpy difference between outlet and inlet air of the regeneration heating coil. According to Figure 2.5.2 the thermodynamic term of the COP is defined as:

Coefficient of performance [-] inputheatDriving

outputCoolingCOP =

( )

( )78Re

14Pr

hhm

hhmCOP

Airgeneration

ocessAir

−−

=&

&

The definition of refrigeration capacity and room cooling capacity are: Refrigeration capacity [kW] ( )14Pr hhmQ ocessAirRC −= &&



Room cooling capacity [kW] ( )45 hhmQ AirRCC −= &&

The Room cooling capacity corresponds to sensible and latent cooling load of the room. Regarding a total energy balance the COP should also take into account all terms of energy consumptions, e.g. electrical and thermal. With open systems, the electrical energy powering the fans is of particular importance as a high number of additional components are usually installed compared to conventional ventilation systems. For such reason DEC-systems entail greater loss of pressure and therefore more electricity to move the air.

2.5.3 Limits of the thermo dynamical principle

2.5.3.1 General ventilation systems

The general design of a DEC-system is based on the same design conditions, which are guilty for conventional air-conditioning system. For Central European countries the ambient air is normally defined by a temperature of 32°C and a hum idity of 40% which corresponds to 12 g/kg absolute humidity. According to the occupant requirements or standard guidelines the air-handling unit has to provide supply air (temperature, humidity and flow rate) to meet the required comfort conditions. In general a air-conditioning system based only on ventilated air (supply and return air ventilation) compensates the thermal loads - e.g. external loads like transmission through the building envelope and solar gains through windows or internal loads like person, lighting or technical equipment - by the ventilated air itself. Thus, there are practical and thermo dynamical limits regarding the cooling power. Considering a design case where air-conditioning is required for human occupied rooms – e.g. hotels, offices or seminar rooms – for comfort reasons the supply air temperature should not fall below 15°C. Furthermore it is unacceptable to force indoor air velocity by increased flow rates of the installed ventilation system. Ignoring these elementary design aspects the ventilation system would lead sooner or later to a sick building syndrome (SBS). Figure 2.5.3 shows an example for a typical design for Central Europe. During the cooling period the air-conditioning system provides the supply air with a temperature of Tsupply = 17°C and a humidity of Xsupply = 10.57 g/kg. Due to latent and sensible loads of the room the return air temperature is Treturn = 26°C and the humidity is X return = 11.896 g/kg. Considering a normalised flow rate of 1000 m³ per hour the air-handling unit treats the ambient air by a refrigeration power of 6.448 kW per 1000 m³/h to blow the required supply air into the room. A cooling load of 4.130 kW per 1000 m³/h can be covered by the ventilation following the defined design conditions. Figure 2.5.4 lists a variation of supply air temperatures and the calculated refrigeration power which is needed to cover a certain cooling load. The return air is assumed as constant and defined as mentioned above (Treturn = 26°C; X return = 11.896 g/kg). This sensitivity analysis takes only in account a variation of sensible load, e.g. a constant latent load is assumed. If the supply air temperature verifies between 15°C and 21°C the air-conditioning system covers a specific cooling load from 4.8 kW per 1000 m³/h down to 2.8 kW per 1000 m³/h which corresponds to a refrigeration power range from 7.1 kW per 1000 m³/h to 5.1 kW per 1000 m³/h. A DEC-system is an overall air-conditioning system which also compensates cooling loads by using ventilated air. Therefore the achievable maximal values regarding cooling power or ventilated cooling loads are analogue valid for desiccant cooling systems and limit their application. Consequently the cooling load strongly determines the design of the air-conditioning system.



DEC Design

sealevel 200.00 m

T 26.00 °CRH 55.00 %X 11.896 g/kgEnthalpy 56.378 kJ/kg

T 32.00 °C T 17.00 °CRH 40.00 % RH 85.00 %X 12.251 g/kg X 10.570 g/kgEnthalpy 63.447 kJ/kg Enthalpy 43.776 kJ/kg

Return Air

Supply AirAmbient Air

exhaust air

Figure2.5.3: Design conditions for air-handling unit in Central Europe.

0.0

1.0

2.0

3.0

4.0

5.0

6.0

7.0

8.0

15 16 17 18 19 20 21

Temperature (supply air) [°C]

Spe

c. c

oolin

g po

wer

or

cool

ing

load

[k

W/1

000

m³/

h]

Cooling power (machine)

Cooling load (room)

Calculationtemperature [°C] 15 16 17 18 19 20 21humdity [g/kg] 10.57 10.57 10.57 10.57 10.57 10.57 10.57relative humdity [%] 96.6 90.6 85.0 79.8 74.9 70.4 66.2enthalpy [kJ/kg] 41.73 42.75 43.78 44.80 45.83 46.85 47.87

Enthalpy difference Ambient - Supply [kJ/kg] 21.72 20.70 19.67 18.65 17.62 16.60 15.57Enthalpy difference Supply - Return [kJ/kg] 14.65 13.63 12.60 11.58 10.55 9.53 8.50

Cooling power (machine) [kW/1000 m³] 7.119 6.783 6.448 6.112 5.776 5.440 5.104Cooling load (room) [kW/1000 m³] 4.802 4.466 4.130 3.795 3.459 3.123 2.787

Supply air

Figure2.5.4: Variation of supply air temperatures and the calculated refrigeration power which is needed

to cover a certain cooling load. The ambient air is assumed by a temperature of 32 °C and a humidity of 40% which corresponds to 12 g/kg absolute humidity. The return air is assumed as constant and defined as mentioned above (Treturn = 26°C; X return = 11.896 g/kg).

2.5.3.2 Desiccant cooling systems

For moderate climates (warm and humid; humidity values lower than 15 g/kg) DEC-systems are a capable application for air-conditioning (/Eic01/). Especially in extreme dry climates air-conditioning can be based on evaporative cooling without any sorption wheel. Assuming that the dehumidification performance of the sorption wheel is more or less ideal, i.e. dehumidification capacity of around 6 g/kg, the DEC-system provides supply air with a

WP5.D1: Technical status report on solar desalinati on and solar cooling Date: November 2005 page 32 of 102 pages


temperature of 15°C and a humidity of 9.5 g/kg. Thi s consideration take into account the standard design conditions for air-conditioning systems in Central Europe (Tamb = 32°C; RHamb = 40%). Without any dehumidification by sorption the air-handling unit (supply air humidifier and adiabatic cooling of the exhaust air) only obtains a supply air temperature of around 20°C. The humidity values exceed already the required supply air humidity of maximal 11.6 g/kg. DEC-systems in operation provide supply air temperatures in the range of 17°C and 19°C (standard design conditions). The humidifier efficiency is lower than 95% and the efficiency of the heat recovery wheel is in practice between 70% and 75%.

2.5.4 Configuration of Desiccant cooling systems

Analogue to the immense diversity of conventional air-conditioning system configuration DEC-system design are also manifold. DEC-systems consist on different components, e.g. heat recovery wheel, heating or cooling coil, desiccant wheel, humidifier - and their position in the system itself verify. Figure 2.5.5 illustrates promising DEC-system configuration designed for climates of Central Europe (/Eic97/). DEC-Systems I, II, VII and VIII in Figure 2.5.5 are classified as single-stage system configuration. Two heat recovery wheels in sequence are implemented in system configuration VII. In two steps the exhaust air is humidified in order to achieve an improved overall heat transfer performance of both heat recovery wheels, IV, V and VI in Figure 2.5.6 are classified as cascade system configuration. In such systems an additional heat pump contributes to transfer sensible heat from the supply air to the exhaust air. Only DEC-system configuration VI represents a combination between air-conditioning by ventilated air and a cooling ceiling which is provided by cold water of the heat pump. Cooling ceilings only cover sensible cooling loads.

heating coil supplied by solar energy

Figure 2.5.5: Configuration of DEC-systems considered as single-stage or cascade DEC-systems (source: Heinrich/ Franske »Solargestützte Klimatisierung«/ Germany)



Concerning the applicable solar collector (heat sources) for the regeneration process all DEC-system configurations - except for DEC-System VIII - are suitable configurations for solar collectors with liquid fluids. The solar heat provides the heating coil mounted in the exhaust air stream just in front of the desiccant wheel. The heating coil in the supply air stream can also be driven by heat from the solar collector. DEC-System configuration VIII is designed for a direct regeneration air heating. During the cooling period ambient air is heated by means of the solar air collector and is used to regenerate the desiccant wheel. The advantage of this DEC-system configuration is that there is no heat transfer from water to air. Water-air heat exchangers operate normally with a temperature level decrease which would be a disadvantage for the required high regeneration temperatures. A detailed description of such a DEC-system configuration takes places in chapter 2.5.7 where examples for solar driven DEC-system are presented. The application of direct heated regeneration air results in improved collector performances but the installation and assembly of the air ducts is more complicate and extravagant. For regions with a humid and warm climate DEC-system with cascade configuration are recommended. DEC-system cascades with integrated heat pumps and cooling ceilings are more or less the most sufficient system configurations to meet the required comfort demand. The overall DEC-system performance is improved when the design of the heat pump condensator covers the required regeneration heat. In this case the DEC-system performance should benefit from adiabatic evaporative cooling in the exhaust air stream because the design evaporator temperature of the heat pump can be higher. Consequently the design of the heat pump leads to smaller capacities which do have two positive impacts. On the one hand the primary energy consumption of the DEC-system is reduced and on the other hand the investment cost of the overall DEC-system is lower. Finally DEC-systems consume less primary energy or operate more efficient than conventional air-conditioning systems.

2.5.5 Desiccant Wheel

In practise DEC-systems are commonly equipped with desiccant wheels. The permanent rotation of the desiccant wheel facilitates supply air dehumidification which results in a continuously operating air-conditioning process.

Figure 2.5.6: Schematic drawing of a desiccant wheel compound (left). The rotation is driven via a

belt drive. The regeneration heat is supplied by a water-air heat exchanger (Source: University Hamburg Harburg / Germany). An Example of a desiccant wheel integrated into a cassette is on the right (Source: Klingenburg GmbH / Germany)



The desiccant wheel is quite similar constructed to a heat recovery wheel. As a basic principle a rotating matrix is passed through two air streams in counter-flow. Figure 2.5.6 shows the general construction. In comparison to heat recovery wheels the matrix of desiccant wheels is additionally coated with solid desiccant material. Typically applied solid materials are silica-gel or zoelite and other hygroscopic chemical compounds. The basic material which forms the supporting structure is a mix of different fibres, including glass, ceramic binders and heat resistant plastics. An example of a typical desiccant wheel performance under design conditions is shown in Figure 2.5.7. This Figure 2.5.7 represents performance data for a desiccant wheel which is applied in a DEC-system of the European Project DESODEC in Armenia (INCO-COPERNICUS Programme of Commission of the European Communities). The shown desiccant wheel is designed for a volume flow of 8 500 m³/h. In such system the desiccant wheel is regenerated by a reduced air-flow. The regeneration air-flow is designed by 80% of the process air-flow. The regeneration heat is provided by flat-plate collectors which generate a regeneration air temperature of 85°C in the design point.

IN85.00 T [°C]15.04 x [g/kg]

OUT 3.91 r [%]T [°C] 48.05 7898 v [m³/h]

x [g/kg] 27.87 7157 m [kg/h]r [%] 36.73

v [m³/h] 7226 OUTm [kg/h] 7157 59.96 T [°C]

5.57 x [g/kg]4.27 r [%]

IN 9649 v [m³/h]T [°C] 32.00 9543 m [kg/h]

x [g/kg] 15.19r [%] 48.00

v [m³/h] 8975m [kg/h] 9543

regeneration air

process air

IN85.00 T [°C]15.04 x [g/kg]

OUT 3.91 r [%]T [°C] 48.05 7898 v [m³/h]

x [g/kg] 27.87 7157 m [kg/h]r [%] 36.73

v [m³/h] 7226 OUTm [kg/h] 7157 59.96 T [°C]

5.57 x [g/kg]4.27 r [%]

IN 9649 v [m³/h]T [°C] 32.00 9543 m [kg/h]

x [g/kg] 15.19r [%] 48.00

v [m³/h] 8975m [kg/h] 9543

regeneration air

process air

cleaning section

IN85.00 T [°C]15.04 x [g/kg]

OUT 3.91 r [%]T [°C] 48.05 7898 v [m³/h]

x [g/kg] 27.87 7157 m [kg/h]r [%] 36.73

v [m³/h] 7226 OUTm [kg/h] 7157 59.96 T [°C]

5.57 x [g/kg]4.27 r [%]

IN 9649 v [m³/h]T [°C] 32.00 9543 m [kg/h]

x [g/kg] 15.19r [%] 48.00

v [m³/h] 8975m [kg/h] 9543

regeneration air

process air

IN85.00 T [°C]15.04 x [g/kg]

OUT 3.91 r [%]T [°C] 48.05 7898 v [m³/h]

x [g/kg] 27.87 7157 m [kg/h]r [%] 36.73

v [m³/h] 7226 OUTm [kg/h] 7157 59.96 T [°C]

5.57 x [g/kg]4.27 r [%]

IN 9649 v [m³/h]T [°C] 32.00 9543 m [kg/h]

x [g/kg] 15.19r [%] 48.00

v [m³/h] 8975m [kg/h] 9543

regeneration air

process air

IN85.00 T [°C]15.04 x [g/kg]

OUT 3.91 r [%]T [°C] 48.05 7898 v [m³/h]

x [g/kg] 27.87 7157 m [kg/h]r [%] 36.73

v [m³/h] 7226 OUTm [kg/h] 7157 59.96 T [°C]

5.57 x [g/kg]4.27 r [%]

IN 9649 v [m³/h]T [°C] 32.00 9543 m [kg/h]

x [g/kg] 15.19r [%] 48.00

v [m³/h] 8975m [kg/h] 9543

regeneration air

process air

cleaning section

Figure 2.5.7: Performance data for a desiccant wheel which was proposed for a DEC-system in the frame of the European INCO-Copernicus Project DESODEC in Armenia (Source: Fraunhofer ISE / Freiburg / Germany)

Table 2.5.1 lists performance data of a desiccant wheel which is designed for an air-flow of 3000 m³/h. The temperature and humidity values of both process air and regeneration air differ, so that different dehumidification rates appear. The listed data are provided by Klingenburg.

Inlet air Outlet air Dehumidification rate

[°C] [g/kg] [°C] [g/kg] [g/kg] Process air 32 12.0 53.2 6.9 A Regeneration air 70 14.0 48.7 19.1

5.1

Process air 32 12.0 47.4 7.4 B Regeneration air 60 12.0 44.5 18.6

4.6

Process air 35 15.0 50.9 9.4 C Regeneration air 70 15.0 54.1 19.6

5.6

Process air 35 15.0 46.2 10.0 D Regeneration air 60 15.0 48.8 18.9

5.0

Table 2.5.1: Performance data of a desiccant wheel design for an air-flow of 3000 m³/h (Source: Klingenburg 1999)



In dependence of the rotation speed the desiccant wheel operates in two different modes – on one hand the dehumidification mode and on the other hand the enthalpy recovery mode . Using the desiccant wheel in the dehumidification mode the rotation speed is generally in the range of 6 - 12 rotations per hour. If the desiccant wheel rotates with 8-14 rotations per minute, it performs as an enthalpy recovery wheel .

Figure 2.5.8 illustrates both the dehumidification mode and the enthalpy mode . The supply air is dehumidified (state change 1 - 2) on the process side of the wheel. The return air is heated up to a sufficient regeneration temperature which flows through the regeneration side of the wheel (state change 8 - 9). The desorption of the water bounded in the desiccant material on the process side is activated since the vapour pressure of the water bound in it exceeds the partial pressure of the water vapour in the warm regeneration air. The energy associated with the sorption and desorption processes is equal to the latent heat of condensation plus a differential heat of sorption. It is beneficial to have a low total heat of sorption. In addition, the state change is also affected by the heat stored in the rotor matrix on the regeneration side. As mentioned above at high rotation speeds the desiccant wheel performs as an enthalpy recovery wheel. Especially in wintertime heat and moisture recovery from the return stream is required thus the overall performance of the DEC-system benefits from the desiccant wheel only by increasing the rotation speed. The enthalpy recovery process is represented on the psychometric chart along a line which connects both inlet air points for the two streams – see again Figure 2.5.8.

The performance of desiccant wheels is mainly characterised by the dehumidification capacity which describes a number of bounded water per kilogram dry air. According to statements of desiccant wheel manufactures typical dehumidification capacities are in a range of 4 to 6 gH2O/ kg Dry Air for a regeneration temperature of 70°C. To operate the desiccant wheel with an optimal performance the rotation speed is one of the important parameter. Depending on the values of humidity and regeneration temperature the dehumidification capacity is a function of rotation speed. Figure 2.5.9 illustrates some measurement results1 of a market available desiccant wheel. During the test the rotation speed was verified by values of 10, 15 and 20 rotations per hour. The maximal dehumidification capacity occurs with 15 rotations per hour. Providing a regeneration temperature of 70°C and rotating the wheel by 15 rotations per hour the dehumidification leads to values of round about 6 gH2O/ kg Dry Air.

1 Test facilities at University of Applied Sciences Stuttgart, Germany

0

10

20

30

40

50

60

70

80

0 2 4 6 8 10 12 14 16 18 20 22

8

ϕ = 100 %

ϕ = 50 %

ϕ = 30 %

ϕ = 20 %

ϕ = 10 %

ϕ = 5 %

2

1

9dehumdification

enthalpy recovery

Figure 2.5.8: Psychometric chart for moist air

showing the state changes for dehumidification of air in a desiccant wheel (Source: Arsenal research)



Some manufactures of desiccant wheels offer additionally a cleaning sector located between the supply air and exhaust air stream. This particular zone is passed by fresh ambient air and reduces the percentage of exhaust air transfer to the supply air, e.g. there is a hygienic benefit. The cleaning sector is used only if it is intended to prevent soiling or entrained olfactory impacts. In addition the heat transfer from the regeneration side to the supply air side is also minimised. Stored heat in the matrix on the regeneration side is pre-cooled by the cooler ambient air. Consequently the temperature increase of the supply air is lower which results in an

improved performance of the overall desiccant system. Figure 2.5.9 is a schematic illustration of such a system with an additional fresh air zone.

Focusing on the general design of a DEC-system the desiccant wheel diameter strongly determines the geometric dimension of the air-handling unit itself. Figure 2.5.10 shows the influence of the required air-flow on the diameter of the desiccant wheel. For example the DEC-system is designed for an air-flow of 12 500 m³ per hour the diameter of a sufficient desiccant wheel is around 2000 mm. The dimension of a DEC-system is an important issue during planning phase especially for an installation in existing buildings.

Company Country of origin Desiccant Wheel Size

Munters USA USA SiGel, AlTi, Silicates, New Proprietary 0.25 – 4.5 m Munters AB Sweden SiGel, AlTi, Silicates, New Proprietary 0.25 – 4.5 m Seibu Giken Japan SiGel, Am, Silicates, New Proprietary 0.10 – 6.0 m

Nichias Japan SiGel, Mol, Sieves 0.10 – 4.0 m DRI India SiGel, Mol, Sieves 0.30 – 4.0 m

Klingenburg Germany Al oxide, LiCl 0.60 – 5.0 m PorFlute Sweden SiGel, Mol, Sieves 0.50 – 3.0 m

Rotor Source US SiGel, Mol, Sieves 0.50 – 3.0 m NovelAire US SiGel, Mol, Sieves 0.50 – 3.0 m

Table 2.5.2: Manufacturers and product description of sorption dehumidifiers. The list does not claim to be exhaustive (Source: Handbook IEA SHC Task 25, 2003).

Regeneration temperature [°C]

Deh

umid

ifica

tion

[g/k

g]

Regeneration temperature [°C]

Deh

umid

ifica

tion

[g/k

g]

Figure 2.5.9: Influence of rotation speed and regeneration

temperature on the dehumidification capacity of a desiccant wheel (Source: University of Applied Sciences Stuttgart/ Germany)

-

500

1 000

1 500

2 000

2 500

3 000

3 500

4 000

-

5 00

0

10 0

00

15 0

00

20 0

00

25 0

00

30 0

00

35 0

00

40 0

00

45 0

00

air-flow [m³/h]

mat

rix d

iam

eter

[mm

]

Figure 2.5.10: Influence of the required air-flow on the diameter

of the desiccant wheel (Source: Klingenburg / Germany)



World wide there are more than 9 manufactures of desiccant wheels. Table 2.5.2 shows a list of desiccant wheel manufacturers provided by the handbook of IEA SHC Task 25. It is important to mention that desiccant wheels are not an independent market product but are purchased by manufacturers of air-handling units. Prices for desiccant wheels differ strongly on the size of the air-handling unit, on the applied materials and on the manufactures itself.

2.5.6 DEC-system with liquid sorbent materials

Open desiccant air-conditioning systems with liquid sorbent materials work according to the same principle as all open processes: ambient air is dehumidified by means of sorption and cooled by water evaporation. In such liquid desiccant system water serves as the refrigerant. A large number of working fluid pairs are available for closed absorption refrigerating machines but there are only a small number of suitable materials for open liquid-based systems. According to the strict limitations for ventilation systems in which materials come in direct contact with the environment the used solutions should be non-toxic and environmentally friendly. They also should not contain any volatile material other than water. In practice, liquid sorbent agents which consist principally of salts dissolved in water are mainly used, e.g., lithium chloride or calcium chloride. These hygroscopic salts lower the vapour pressure of water in solution sufficiently to absorb humidity from the air. In contrast to the case of the solid sorbents, the water bonding mechanism is not adsorption, but absorption.

Figure 2.5.11 shows a DEC-system with liquid sorbent material optimised for solar operation. Ambient air is dehumidified in the absorber, where cooled contact surfaces are humidified with a concentrated liquid sorbent material using the falling film technique. The sorption heat is transferred to the exhaust air through a composite circuit system and an indirect evaporation cooler so that the ambient air is dehumidified and cooled at the same time. A downstream cooler cools the dry air below room temperature. The sorbent material is diluted during dehumidification of the air. In an air-flow regenerator it is heated up to 60-80°C and re-concentrated. Heat recovery from the air and the sorbent material increases the efficiency and saves collector area. Energy can be stored by storing diluted and concentrated sorbent

separately. When using the usual aqueous lithium chloride solution as a sorbent, it is possible to achieve an energy storage density of up to 280 kWh/m² by using a special internally cooled absorber without diminishing the dehumidification potential of the concentrated solution. In practise there are different ways to assemble such DEC-systems with liquid sorbent material. Figure 2.5.11 is a schematic drawing of the system with liquid sorption and solar regeneration which is developed by ZAE Bayern/ Germany. Another concept is focused on a compact central evaporative cooling unit. In cooperation Menerga and the University of Essen/ Germany developed such compact DEC-system which is shown in Figure 2.5.12. Since 2003 a pilot DEC-system design for a volume flow of 1500 m³/h is implemented at Fraunhofer Solar Building Innovation Centre (SOBIC) in Freiburg/Germany. The tested system performance allows an optimistic market perspective for such desiccant cooling system with liquid sorption material. Particularly with regard to the coupling with solar thermal

Figure 2.5.11: Schematic drawing of a desiccant

cooling unit using liquid sorption and solar(source: Menerga/Germany) driven regeneration (Source: ZAE Bayern - Germany)



energy the desiccant cooling is an attractive technical concept to reduce primary energy consumption for air-conditioning.

ambient airsupply air

exhaust airreturn air

heat recovery andevaporative cooling regenerator

absorberwater pump solution pump

solutioncooler, heater

ambient airsupply air

exhaust airreturn air

heat recovery andevaporative cooling regenerator

absorberwater pump solution pump

solutioncooler, heater

Figure 2.5.12: Schematic drawing of a central evaporative cooling unit operating on the principle of a

liquid sorption and solar regeneration (Source: Menerga / Germany)

Desiccant cooling systems with liquid sorbent materials are more complex than systems with rotors and are not yet available on the market. Some fundamental advantages such as potentially higher overall efficiency due to greater potential of heat and refrigeration recovery, lower possible regeneration temperature with the same dehumidification potential due to cooling of the sorption process, and the potential of efficient energy storage and, not least, the physical separation of supply and exhaust air-flows could help establish them in combination with solar systems.

2.5.7 Examples of DEC-System plants

2.5.7.1 Freiburg/ Germany

In Figure 2.5.13 is shown the Chamber of Trade and Commerce for the south-west region of Germany (IHK Südlicher Oberrhein). The building is located in Freiburg and was constructed in 1992. The top floor of the building is designed with large glazed facades to the east, the west and the south. Due to this architectural design the cooling load of both - seminar room and small cafeteria located on this top floor – correlates with solar gains during sunny and warm days. Both seminar room and cafeteria are mainly used for meetings or other events and both were originally only equipped with an air ventilation system without air-conditioning function. However, conditions in the rooms during summer were very often

uncomfortable, so it has been decided to install an air-conditioning system. The DEC-system was designed as a solar autonomous DEC-system. The air-handling unit is directly ducted to the solar air collector system without any back-up heat and thermal storage. The design values for the desiccant air-handling unit are listed in Table 2.5.3. In summer the ambient air is heated by means of the solar air collector and is used to regenerate the sorption wheel.

Figure 2.5.13: View of the IHK building in

Freiburg / Germany (Source: Fraunhofer ISE/ Germany).



The DEC-design requires an additional fan compared to standard systems, in which return air is used for regeneration. During operation in winter the solar air collector is used for pre-heating of fresh air before it enters the heat recovery system. In case of high solar gains the heat recovery wheel is not operated. The real performance of the system for the period between July 2001 and January 2002 was investigated by Fraunhofer ISE/ Freiburg. It can be mention that the maximum room temperature corresponding to the comfort zone is exceeded only during a very few hours. The humidity value is exceeded more often, but except on two days the design value of 60%

relative humidity is met. According to the design study it was estimated that the system would save about 30% of primary energy compared to a reference system. The absolute primary energy consumption of the reference system was calculated as 25922 kWh and that of the desiccant cooling system with solar air collectors as the heat source as 18162 kWh. An air-handling unit with heat recovery and a conventional compression chiller for dehumidification and temperature control was assumed, as reference system. The energy figures are valid for the entire year, including the heating season.

2.5.7.2 Hartberg/ Austria

The research house is located in the Ökopark Hartberg which is about 3.5 km in the south of Hartberg in Styria (South/East of Austria). The research house is used for seminars and conferences on the one hand and on the other hand, it consists of an office infrastructure. It is a building with two floors (each about 140 m²) with a glass facade in the south (in the lower part there are 11 vacuum tube collectors). The DEC-system is designed for an air flow of 6000 m³/h which corresponds to a cooling capacity of 30 kW. The heat to regenerate the desiccant wheel is supplied by two options. Solar collectors or a mobile pellet biomass heater provide the regeneration energy. A tank of 2000 l water puffers the energy for cooling and heating. Using a biomass/solar driven DEC-system the office building is air-conditioned in

summer, is heated in winter and is supplied by fresh air in the intermediate season. Concerning the cooling this DEC-system operates in two modes:

1. Adiabatic evaporative cooling mode; electrical power for the fans and liquid water to humidification is required

2. Desiccant mode; additional heat to regenerate the desiccant wheel is required Based on the results of a measurement campaign and the experience with such a system it could be mentioned that the operation mode “adiabatic evaporative cooling” is quite sufficient. In many days in summertime of the south-west Austria the cooling load was covered by the

Table 2.5.3: Design values for the desiccant air-handling unit at IHK building in Freiburg (Source: Fraunhofer ISE / Germany)

Figure 2.5.14: The DEC-system is installed

outdoor, in the western part of the building (Source: Arsenal research / Austria).



operation mode “adiabatic evaporative cooling”. Days with higher values of humidity require the necessity of operation mode “desiccant cooling”. Due to the scientific analyses there are important results on the control strategy development. Optimised control strategy minimises significantly the heat demand to drive the DEC-system and additionally it leads to an improved system operation with a high efficiency and reliability.

2.5.7.3 Mataró/Spain

The building was equipped with four conventional air conditioning units, where the cooling energy was provided by heat exchangers from a central electrical compressor chiller. One of the air conditioning units for the children’s reading and multimedia room with 510 m² surface areas was replaced by the desiccant cooling plant and provides cool fresh air via 15 ceiling air outlets. The building had a ventilated photovoltaic façade (244 m²) and shed roofs (330 m²) with a total electrical power of 55 kip. A desiccant cooling plant with a process air volume flow of 12000 m³/h was installed in a public library building in Mataró (Spain) with 3500 m² used surface. The desiccant cooling

system is coupled with combined solar thermal energy system. On one hand a PV-solar air pre-heating system is installed and on the other hand solar air collectors provide the required regeneration air temperature. The heat produced by the photovoltaic modules is transferred into a 14 cm wide air gap, which is exhausted by the desiccant cooling regeneration fan. Two additional air collector fields in the façade (50 m²) and roof (105 m² at 34° tilt angle) increase the temperature level to the required regeneration air temperature. The common regeneration ventilator is volume flow controlled to provide a regeneration temperature between 50 and 70°C. Concerning the performance of the plant: with a yearly irradiance of 1020 kWh/m² on the vertical south façade and 1570 kWh/m² on the shed roofs the combined solar thermal energy system is calculated to produce nearly 70000 kWh useful thermal energy (April to October). This corresponds to 93% covering of the cooling demand of 44000 kWh. The thermal efficiency of the ventilated photovoltaic system is rather low (12-15%), because flow velocities in the many parallel large air gaps reach only 0.3 m/s and the maximum temperature level increase is between 10-15K. The complete volume flow through the ventilated PV system of 3000-9000 m³/h passes three parallel air collector fields so that flow velocities in the 9,5 cm air channels are between 3 and 9 m/s and efficiencies are considerably higher (in the range of 50%).

2.6 OTHER COOLING SYSTEMS

2.6.1 Organic Rankine Cycle

In principle, amongst the cooling systems that can be coupled with low to medium temperature solar thermal collectors, also the traditional mechanical vapour compression chiller shall be included. The compressor can be operated by the solar driven engine either directly or via a preliminary conversion into electrical energy. In any case, in the range of collectors working temperature under investigation, a low global efficiency of the cooling system has to be expected. On the other hand, this topic is part of the wider problem regarding the exploitation of solar energy for power generation. Consequently if the obtained results should be somehow

Figure 2.5.15: Public library building in Mataró

(Source: University of Applied Sciences Stuttgart, Germany)



encouraging, cooling would simple one of the several potential applications. Despite this general framework, a brief description of a solar driven vapour compression chiller has been considered useful in order to get a complete overview of solar cooling systems and to give an estimation of the advantages related to the use of thermally driven machines. First of all, it must be assessed the technical practicability of this technological option. Concerning this, the use of low grade heat for the production of mechanical / electrical energy in small size plants has been

experimented since ’70, by means of the Organic Rankine Cycle (ORC) engines (see Figure 2.6.1). A number of solutions have been designed for a variety of heat sources (solar energy, geothermal fluids, biomass, process waste heat), within an ample range of power outputs (from few KW to some MWe) and of maximum operating temperatures (from 70 to 350 °C). The use of working fluids, such like hydrofluorocarbon, perfluorocarbon, etc., in place of water allows the achievement of the main following advantages:

� successful exploitation of low temperature heat sources with an elevated turbine efficiency (up to 85%), due to the high molecular mass and the appropriate critical parameters of the working fluid;

� low mechanical stress of the turbine and direct drive of the generator or the compressor (without reduction gear), due to the reduced RPM of the turbine and the consequent moderate peripheral speed;

� no erosion of blades, due to the absence of moisture in the vapour nozzles, with a quiet operation, long life and minimum maintenance requirements (considering also the reduced fatigue damage previously stated);

� simple start-stop operations and good part load performance (especially beneficial for the coupling with a discontinuous source such like solar energy).

For very low driving temperature, the effect of regeneration is negligible, so the real cycle tends to approach a Carnot cycle, except essentially for the irreversibility due to the expansion in the turbine. Differently for a top operating temperature around 300°C, the ratio between the Rankine and the corresponding Carnot cycle decreases to some 0.6. According to this premise, the overall efficiency for the conversion from the thermal energy of the carrier fluid to the mechanical energy available at the drive shaft, grows with the maximum cycle temperature as reported in Figure 2.6.2. In conclusion, the use of low to medium temperature collectors to drive vapour compression machines is certainly viable

HEAT SOURCE

CONDENSER

HEAT EXCHANGER

CONDENSER

EVAPORATOR

COMPRESSOR

EXPANSION VALVE

REGOLATION VALVES

TURBINE

VAPOUR COMPRESSION CHILLERORGANIC RANKINE CYCLE

Figure 2.6.1: Schematic of an Organic Rankine Cycle (ORC).

0,0

0,1

0,2

0,3

0 50 100 150 200 250 300

Maximum cycle temperature (°C)

Effi

cien

cy

Figure 2.6.2: Efficiency in function of maximum cycle

temperature.



from the technical point of view. In fact the range of working temperature of ORC is compatible with these type of heat source and adequately small size engines (equivalent to 10-500 KW of cooling capacity) are commercially available. Furthermore, it is to be noted that in this case the losses of the diathermic oil boiler, which amount to more than 15% of the heat of combustion of the fuel, are eliminated, so the values shown in figure are fairly realistic. On the other hand, hypothesizing a maximum working temperature up to about 200°C for the collectors un der investigation, the global efficiency of an ORC can range from little less than 10% to little more than 20%. Therefore, even though the coupling between solar collectors and vapour compression chillers is possible, a number of drawbacks with respect to thermally driven machines are present:

� low COP within the range 0.2-0.5, according to the selected solar technology: the inferior limit is referred to conventional collectors working at 70°C and the upper value to parabolic throughout collectors working at 200°C;

� remarkable additional investment cost, due to the introduction of the ORC turbine, which impacts also on running costs and dramatically increases the system complexity;

� especially high cost per unit and heavy management in case of required cooling capacities smaller than 100 KW, which on the other hand cover the most important segment of the cooling market;

� fully based on mature technologies with limited expectations of development for both the ORC turbine and the vapour compression chiller.

Specific positive aspects of these cooling systems are the full reversibility with a significantly higher COP, when working as a heat pump and the opportunity of using the electric grid as a storage device. However, even if solar assisted systems for the production of cold via a preliminary conversion into electrical energy should be considered encouraging, other technologies are to be investigated, such as high temperature solar thermal systems and photovoltaic, but this type of application is beyond the objectives of the present study.

2.6.2 Ejector air-conditioning system

The operation principle of steam jet ejectors is the entrainment and successive compression of vapour via the energy supplied by a heat source. Therefore this device can completely replace the conventional mechanical vapour compressor within a refrigeration cycle: really it allows to boost the pressure of the working fluid coming from the evaporator, so that condensation temperature overtakes evaporation temperature and heat flows from the cold reservoir to the hot reservoir. The grade of the energy required by this process makes it to be in theory suitable for the coupling with low to medium temperature solar thermal collectors, but the current interest in this application is scarce, due to the limitations that will be briefly discussed in this section. On the other hand steam ejectors are valuable for several other uses, including seawater desalination, removal of non-condensable gases, transport of solids, and gas recovery. The function of the ejector may differ considerably according to the type of process considered. Details about its use in refrigeration and air conditioning cycles will be given afterwards. The basic components of this system are the ejector, the generator, and the circulation pump, in addition to the evaporator, the condenser, and the expansion valve, which, together with the mechanical compressor, constitute the conventional refrigeration cycle. In Figure 2.6.3 the arrangement of these components is illustrated. The cycle is carried out as it follows. A heat source delivers low grade thermal energy to the generator, where the liquid refrigerant is vaporized at high pressure. The resultant driving vapour is accelerated through the primary nozzle till the flow becomes supersonic, causing at the exit a low pressure region.



Hence the vapour from the evaporator is entrained into the ejector and combined with the driving vapour in the mixing section. The resulting stream undergoes a transverse shock inside the constant area section of the diffuser, due to the back pressure resistance of the condenser, which induces both a sudden static pressure rise and a reduction in velocity of the mixture to subsonic conditions. A further pressure recovery, until the value pertinent to the condenser is reached, occurs in the diffuser section. Then the mixed steam flows to the condenser where it

condenses, rejecting the corresponding heat to the environment. The condensate is divided into two streams again: one is pumped back to the generator, while the other expands via a valve to a low-pressure state and enters the evaporator, where it is evaporated in order to produce the required cooling effect. The obtained vapour is finally entrained by the ejector, thus completing the cycle. The ejector configuration and the variations of pressure inside it are shown in Figure 2.6.4.

MIXINGSECTION

NOZZLE SECTION DIFFUSER SECTION CONSTANT AREA SECTION

VAPOUR MIXTURETO THE CONDENSER

VAPOUR FROM EVAPORATOR

VAPOUR FROM GENERATOR

GENERATORCONDENSER

EVAPORATOR

SHOCK

PRESSURE

POSITION Figure 2.6.4: Ejector internal structure (a) and associated pressure profile (b).

Figure 2.6.3: Schematic representation of a steam ejector refrigeration system.

EVAPORATOR

GENERATOR

EXPANSION VALVE

PUMPEJECTOR

HEATSOURCE

CONDENSER



The working fluid is normally water, owing to its several advantageous properties, such like the low cost and non-toxicity or flammability. On the other hand the freezing point and the specific volume are rather high. However the last aspect is not so critical for steam ejectors, where moving parts are not present and higher volumetric flow rates can be processed simply by increasing the diameters of the tubing. When the motive steam is generated by a low temperature heat source, such like solar thermal collectors, the use of water requires to operate the whole loop below the atmospheric pressure. Therefore it can be valuable to apply to different working fluids with a lower boiling point, such as halocarbon compounds (/Sun99/). The main limitation of these substances is the relatively fast deterioration as temperature grows. Nevertheless in the range of temperatures below 100 °C numerous compounds (HCFC-141b, HFC-245fa, etc.) are available.

As it can be drawn from the previous description, the total energy input in an ejector refrigeration cycle consists of the low grade heat required by the generator and the work consumed by the circulation pump. The last term is very small compared to the first one, hence the COP of the cycle can be assumed as in direct proportion to the ejector entrainment ratio (ER), which characterizes its performance (/Laz83/, /Eam95/). This parameter is defined as the mass flow rate ratio of the entrained vapour to the motive steam and can be estimated from the expansion and the compression ratios, defined as the pressure ratios to the entrained

vapour of the motive and the mixed steam respectively. In Figure 2.6.5, the reciprocal of the ER as a function of the compression and the expansion ratios is reported (/Pow94/). The curves are obtained by solving the balance equations for the one-dimensional steady-state model of the ejector. The working fluid is water and all processes are considered isentropic. It can be noticed that the ER increases with the generator and evaporator temperatures, while decreases with the condenser temperature. Provided that in refrigeration and air-conditioning cycles the upper and lower temperature limits can not be varied too much, the entrained ratio and, as a result, the COP of the cycle can be improved basically by rising the motive steam temperature. On the other hand, when the driving heat is supplied by solar thermal collectors, a significant increase in the working temperature leads to a dramatic reduction in efficiency, unless a more advanced and costly technology is adopted.

Figure 2.6.5: Example of graph for the calculation of ideal values of the

ER (Source: El-Dessouky et al. /Des02/).



In Figure 2.6.6 the theoretical ER has been plotted as a function of the motive steam temperature, via the above-mentioned chart assuming an evaporator and a condenser temperature of 8 and 30 °C respectively. As it can be observed, the improvement of the ejector performance, when the motive steam temperature grows, is modest. Therefore, it seems to be effective to operate with basically low heat source temperature, when solar thermal collectors are used to drive the cooling system. Furthermore, considering the useful range of temperatures, the ER can

reach at the most a theoretical value not much higher than 0.6. Actually the COP of a real ejector refrigeration cycle is markedly lower, as a result of the irreversibility of the ejector functioning, mostly related to the friction in nozzle and diffuser, the mixing process between two streams at different velocities, and the formation of the shock wave. In effect the efficiency of both the nozzle and the diffuser of a well designed steam ejector does not exceed 0.9, while a value of about 0.95 can be achieved for the mixing chamber. Consequently the COP of an ejector refrigeration cycle is normally not higher than 0.4. In conclusion, the use of steam ejectors in air-conditioning and refrigeration cycles is supported by numerous favourable aspects:

� the configuration of the cycle is remarkably simplified, since ejectors are formed of a single unit connected to tubing of motive, entrained, and mixture steam;

� no valves, rotors, or other moving parts are present, thus the system works without vibration and noisiness and consequently the deterioration is reduced;

� ejectors are available commercially in an ample range of sizes; � capital and maintenance costs are lower in comparison with other processes.

Nevertheless, the following critical drawbacks limit the importance of this technology, in particular when the driving heat must be supplied by solar thermal collectors:

� ejectors are designed to operate at a single optimum point, a deviation from it results in a dramatic deterioration of the system performance, thus the coupling with a fluctuating source, such as solar energy, is problematical;

� the COP is very low, even if compared to that of an Organic Rankine cycle under the same driving temperature, as a result the required collector area can dramatically grow;

� the increase in ejector efficiency with the motive steam temperature is moderate, therefore to operate at low generator temperature can be valuable for solar driven cycles but the COP is further reduced.

Various solutions are available to improve the performance of an ejector refrigeration cycle: 1) the inclusion of a pre-heater and a pre-cooler, to heat the condensate before it flows into

the generator, via the ejector exhaust at the exit of the circulation pump, and to cool the condensate before it enters the expansion valve, via the vapour from the evaporator;

2) the use of a variable position nozzle, to preserve the optimum operating point for the ejector, even if external conditions, such like temperature and flow rate of the different streams, are changed;

3) the development of multi-ejector cycles, to expand the functioning range of the system. Clearly, against a limited improvement of efficiency, these modifications of the basic cycle lead to a loss in simplicity, which is really the main attractive of these systems.

0,0

0,2

0,4

0,6

0,8

60 80 100 120 140

Motive Steam Temperature [°C]

Ent

rain

men

t Rat

io

Figure 2.6.6: Theoretical ER as a function of motive steam

temperature (Tev = 8 °C, T cond = 30 °C).



2.7 COMPARISON OF THE DIFFERENT TECHNOLOGIES UNDER INVESTIGATION

Solar systems have the main purpose of replacing fossil fuel based systems so as to attain a primary energy saving. In this context, the recourse to solar energy for air conditioning in buildings represents a suitable choice not only to reduce the final energy consumptions but also to enlarge the utilization field of solar energy, opening new perspective of development of this technology. But, on the other hand, different technical solutions can be profitably employed. Since the main objective of this work is to support the selection of the most promising solar cooling technologies for a further analysis in the following steps of this task of NEGST project, a classification method, which can be helpful for a rational choice between different options, is therefore necessary. Thus, in this paragraph, a preliminary screening criterion, based on an energy saving approach, is introduced. The calculations have been restricted to commercially available technologies where there is reliable performance data for cooling machines. The obtained results will be, subsequently, analysed considering the system capital cost, the level of commercial maturity, the presence of technological barriers and any other key factor in order to reach a final assessment of the most promising solar systems for cooling.

2.7.1 Description of the methodology

The approach adopted to perform the comparative analysis between the different solar cooling systems can be synthesised in the following main steps:

� Estimation of the cooling and heating loads for two typical reference buildings (both for residential and office application) located in sites representative of different European climatic conditions.

� Calculation of the related primary energy consumption attained by an electric compression chiller, functioning as a heat pump during the winter season, and assumed as reference system for the analysis.

� Assumption of a reasonable fraction of consumed primary energy to be replaced by means of solar energy. This choice could be based on techno-economic considerations, possible limit for accessing to subsidies, etc.

� Choice of suitable thermally driven cooling technologies to be simulated and definition of performance data based on commercially available machines.

� Evaluation of the requested collectors area to attain the desired primary energy saving by means of a suitable simulation model.

� Comparison of obtained results. PHIBAR f-Chart method , applied to a general solar heating system, has been used to perform the thermal analysis of each solar assisted air-conditioning system under investigation (/Duf91/).

Figures 2.7.1 shows the schematic of this type of solar heating system, which can be used in principle for a wide range of thermal applications including space heating and air conditioning. In this system, solar energy is collected and stored in a liquid storage tank via a heat exchanger. In the closed-loop configuration, when required by the user, the heated liquid is pumped, through a

second heat exchanger, from the storage to supply thermal energy to the load. Alternatively, the fluid can be pumped directly to the user in an open-loop system. For this kind of solar heating systems, the thermal energy is delivered when the temperature is above a specified minimum useful temperature (Tmin). The value of Tmin depends on the type of

Figure 2.7.1: Schematic of a general closed-loop solar heating

system (Source: F-Chart - User manual).

T < Tmin



application: for residential space heating, Tmin is the indoor temperature of the building; for thermally driven air-conditioning applications, it depends on the particular installation. An auxiliary heater provides to supply thermal energy when the solar input is insufficient to meet the load. The simulation model adopted in this study is well appropriate to analyse thermally driven chillers (in particular absorption and adsorption cooling cycles), which are characterized by a minimum driving temperature under which they cannot operate properly. Nevertheless, this method can be applied, in principle, to solar assisted DEC systems too, even if in this case there is a strong dependence of the functioning temperature from the external air conditions. To overcome this difficulty and to perform however the performance analysis, average values of COP and driving temperature have been assumed over the entire year.

2.7.2 Reference locations

The comparative analysis has been carried out considering four reference locations, representative of different European climatic conditions. The selected European sites are:

� Palermo (Italy), representative of hot humid climate � Athens (Greece), representative of hot dry climate � Milan (Italy), representative of warm humid climate � Frankfurt (Germany), representative of temperate climate

Table 2.7.1 shows the monthly average value of global solar irradiance on horizontal plane, the mean ambient temperature and the relative humidity of each site.

Site #1 Palermo - IT

Site #2 Athens - GR

Site #3 Milan - IT

Site #4 Frankfurt - DE

Month H

[KWh/m²] Ta

[°C] RH [%]

H [KWh/m²]

Ta [°C]

RH [%]

H [KWh/m²]

Ta [°C]

RH [%]

H [KWh/m²]

Ta [°C]

RH [%]

JAN 67 13.6 80 66 11 71 33 1.6 86 23 1.8 85

FEB 85 13.2 78 74 10.8 71 50 3.2 78 43 2.8 80

MAR 133 12.5 77 104 11.4 68 96 7.1 71 74 6.6 75

APR 170 13.9 74 147 14.4 62 129 10.5 75 114 9.9 69

MAY 204 16.9 73 183 18.8 59 162 15.5 72 150 14.5 69

JUN 218 20.3 71 201 22.8 53 179 19.3 71 152 17.8 69

JUL 220 23.1 72 213 25.5 48 188 22.1 71 162 19.7 68

AUG 194 24.5 74 200 25.8 49 164 21.6 72 134 19.2 71

SEP 153 23.9 75 155 23.8 56 117 18.2 74 93 15.7 75

OCT 114 21.3 78 106 19.5 65 73 12.1 81 55 10.6 81

NOV 73 17.4 80 66 15.9 72 36 6.0 85 28 5.7 84

DEC 59 15.2 80 52 13 72 28 2.1 86 18 3.1 86

Year 1690 18.0 76 1567 17.7 62 1255 11.6 77 1046 10.6 76

Table 2.7.1: Monthly average climatic data (H = monthly global solar irradiance on horizontal plane – Ta = average ambient temperature – RH = average relative humidity).

2.7.3 Heating and cooling needs

2.7.3.1 Architecture

For the loads calculation, carried out by means of EnergyPlus simulation software (/EERE/), two typical reference buildings corresponding to residential and office sectors have been



considered. It is a three-story building with a total effective floor area of 1080 m2. Each storey consists of three apartments (two of 122 m2 and one of 98 m2) and has a height of 3 m. The total air-conditioned area of the building is 1020 m2. The effective glazed area of the building is 135 m2 (1/8 of total floor area). The building is oriented along east-west axis. The architectural design of the building is shown in Figure 2.7.2.

Figure 2.7.2: The architectural design of reference building.

2.7.3.2 Construction

In a recent study (/Pet04/), carried out on the Mitigation of CO2 emission from the building stock, the characterization of the European building stock and the U-values of building types are defined according to climatic zone, building age group and different insulation standards. In the present simulations, the U-values for building envelope components, corresponding to a building age from 1975 to 1990, are used. Table 2.7.2 shows the selected values.

U-value [W/m 2°C] Surfaces Cold climate Moderate climate Warm climate

Roof 0.20 0.50 0.80 Façade 0.30 1.00 1.20 Floor 0.20 0.80 0.80 Windows 2.00 3.50 4.20

Table 2.7.2: U-value selected for the simulations.

2.7.3.3 Internal Loads

The internal loads for people occupancy, lighting, equipment etc. are shown in Table 2.7.3.

Value Parameter Residential Office

People occupancy 25 m2/person 8 m2/person Lighting 10 W/m2 10 W/m2 Equipment 1 W/m2 5 W/m2 Air-change 40 m3/h person 40 m3/h person

Table 2.7.3: Internal loads.

2.7.3.4 Schedules

Residential Building The schedule for lighting and people occupancy are shown in Figure 2.7.3. The operation of equipments is assumed to 100% over the day.



(a) People occupancy

(b) Lighting

Period: 15 April – 14 October

Rest of the year

Figure 2.7.3: Time schedule of people occupancy (a) and lighting (b) in residential building.

Office building The schedule for lighting and people occupancy are shown in Figure 2.7.4. The operation of equipments is same as lighting schedule.

(a) People occupancy

Week days

Saturday

(b) Lighting

Week days

Saturday

Figure 2.7.4: Time schedule of people occupancy (a) and lighting (b) in office building.



2.7.3.5 Energy loads

Table 2.7.4 and Table 2.7.5 show the results of simulations for each type of building under investigation. The cooling and heating loads are indicated in KWh per square meter of conditioned floor area. The simulations have been conducted considering internal set point temperatures of 20°C and 26°C for winter and summer season respectively.


Site #2 Athens - GR

Site #3 Milan - IT

Site #4 Frankfurt - DE Month

Cooling [KWh/m²]

Heating [KWh/m²]

Cooling [KWh/m²]

Heating [KWh/m²]

Cooling [KWh/m²]

Heating [KWh/m²]

Cooling [KWh/m²]

Heating [KWh/m²]

JAN 0.00 -12.34 0.00 -17.04 0.00 -41.38 0.00 -42.83

FEB 0.00 -10.69 0.00 -15.82 0.00 -30.58 0.00 -34.72

MAR 0.00 -13.28 0.00 -16.28 0.00 -21.37 0.00 -31.33

APR 0.00 -9.64 0.00 -8.45 0.00 -11.88 0.00 -15.89

MAY 2.03 0.00 3.01 0.00 2.66 0.00 0.00 -10.38

JUN 6.83 0.00 9.61 0.00 8.11 0.00 1.65 0.00

JUL 15.88 0.00 17.25 0.00 13.33 0.00 4.66 0.00

AUG 22.13 0.00 18.44 0.00 12.17 0.00 4.89 0.00

SEP 18.75 0.00 13.86 0.00 7.90 0.00 0.00 0.00

OCT 10.65 0.00 5.01 0.00 0.00 -12.22 0.00 -21.42

NOV 0.00 -4.39 0.00 -6.72 0.00 -25.40 0.00 -30.49

DEC 0.00 -8.82 0.00 -14.19 0.00 -37.63 0.00 -40.16

Total 76.27 -59.16 67.17 -78.50 44.17 -180.46 11.20 -227.21

Table 2.7.4: Energy loads for residential building.


Site #2 Athens - GR

Site #3 Milan - IT

Site #4 Frankfurt - DE Month

Cooling [KWh/m²]

Heating [KWh/m²]

Cooling [KWh/m²]

Heating [KWh/m²]

Cooling [KWh/m²]

Heating [KWh/m²]

Cooling [KWh/m²]

Heating [KWh/m²]

JAN 0,00 -12,71 0,00 -20,21 0,00 -53,24 0,00 -58,39

FEB 0,00 -11,27 0,00 -19,32 0,00 -38,43 0,00 -46,54

MAR 0,00 -14,60 0,00 -20,25 0,00 -26,22 0,00 -40,92

APR 0,00 -10,08 0,00 -9,65 0,00 -13,63 0,00 -19,53

MAY 2,75 0,00 4,16 0,00 3,32 0,00 0,00 -11,61

JUN 9,43 0,00 13,05 0,00 11,20 0,00 2,20 0,00

JUL 19,06 0,00 22,81 0,00 17,87 0,00 4,92 0,00

AUG 27,44 0,00 24,97 0,00 15,52 0,00 6,19 0,00

SEP 21,93 0,00 18,03 0,00 9,45 0,00 0,00 -10,95

OCT 12,63 0,00 6,32 0,00 0,00 -13,52 0,00 -25,69

NOV 0,00 -4,39 0,00 -7,29 0,00 -29,60 0,00 -38,46

DEC 0,00 -8,78 0,00 -16,88 0,00 -46,50 0,00 -52,22

Total 93,26 -61,83 89,34 -93,61 57,36 -221,14 13,31 -304,31

Table 2.7.5: Energy loads for office building.



2.7.4 Configuration of solar cooling systems

On the basis of initial overview of the market of air-conditioning systems in Europe, it is quite simple to identify as reference conventional cooling technology the electric compression chiller, on the contrary numerous configuration of solar thermal assisted air-conditioning systems do exist, which can differ from each other both for the type of cooling cycle and the type of solar collectors. From the survey of air-conditioning systems, which can be in principle coupled with low to medium temperature solar collectors, the following main technologies have been taken into account for the present analysis:

� The absorption chiller represents the first cooling technology to be analysed. The configurations considered are: single and double-effect cycles using water/LiBr as working pair and the advanced AHE cycles using ammonia/water as working fluid. Each of systems is coupled with the more suitable low to medium temperature solar field to supply thermal energy.

� The second type of cooling system is represented by an adsorption chiller using water/silica-gel as carrier material (the most common sorption material used for this application). Considering the fact that this type of chiller is very similar, from an energetic point of view, to the previous systems and, moreover, a low driving temperature characterizes it, in the present analysis we consider only the coupling with low temperature collectors (flat-plate and evacuated-tube collectors).

� Finally, the third configuration is represented by a desiccant air-conditioning system using solid sorption materials. Given that the desiccant regeneration temperature is quite low, common flat-plate or evacuated-tube collectors can be used. Since the air to be conditioned is directly treated by this type of system, air collectors could be favourable employed too. Nevertheless, in the present analysis, this type of solar collector will not be taken into account, owing to the limitations of the simulation model to consider this type of collector.

Other applications, such like organic rankine cycle (OCR) or ejector air-conditioning systems, are characterized by a modest coefficient of performance (typically less than 0.4); thus a large amount of collector area per m² of conditioned floor area is required for driving the system, making these technologies scarcely attractive for the applications examined in this study. Finally, an interesting application but still at a development stage, is represented by Thermo-Chemical Accumulators (TCA). This technology, even though has not been explicitly considered in the present analysis, is characterized by design parameters comparable to other low temperature cooling systems previously considered; thus its performance figure can be drawn from the results obtained for other similar technologies. Table 2.7.6 shows the nine configurations of solar driven cooling systems identified for the present analysis.

System # Cooling technology Solar collector 1 Flat-plate 2

Water/LiBr - Single Effect Evacuated-tube

3 Stationary CPC 4

Water/LiBr - Double Effect Parabolic trough

5

Absorption chiller

Ammonia/water - AHE cycle Parabolic trough 6 Flat-plate 7

Adsorption chiller (Water / silica-gel) Evacuated-tube

8 Flat-plate 9

Desiccant air-conditioning systems (Desiccant wheel) Evacuated-tube

Table 2.7.6: Solar cooling system configurations considered for the investigation. For each configuration, typical values for the main parameters (COP and driving temperature) have been assumed on the basis of information reported in the previous chapters. Table 2.7.7



shows these values. In consequence of the strong dependence from external, supply and return air conditions, for DEC systems average values for the above-mentioned parameters, corresponding to common design conditions, have been assumed over the entire year. Since the analysis is performed on annual basis, for cooling cycles working only as refrigerator, the heat supplied by the solar field is directly used to meet the load during the winter season. This is accomplished by means of common low temperature heat distribution systems for space heating, assuming a value of 45°C for the supply te mperature. In the present analysis AHE-Cycle using ammonia-water is the only machine simulated as a heat pump. Other absorption technologies, including TCA, as well as adsorption chillers can in principle also be used for heat pumping in winter, but, quite the opposite of ammonia/water absorption chiller, are not at present commercially available for such an application. Design parameters of different solar cooling system s

Absorption chillers Type of cycle H2O/LiBr

Single Effect H2O/LiBr

Double Effect NH3/H2O

AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)

Cooling COP 0.7 1.1 0.8 0.6 0.7 Driving temperature 85°C 150°C 180°C 75°C 65°C Heating COP - - 1.6 - - Heat supply temperature 45°C 45°C 180°C 45°C 45°C

Table 2.7.7: Main design parameters of cooling machines adopted for the present analysis. For the reference electric compression heat pump, COPs of 2.5 and 3.0 have been assumed in the present analysis for cooling and heating respectively. Moreover, in order to assess the primary energy consumption, an average value of 0.4 for the conversion efficiency from primary energy to electricity has been assumed. Finally, Table 2.7.8 summarize the main parameters adopted in the calculation with PHIBAR f-Chart method. With regards to collector efficiency curve parameters, test intercept and slope have been obtained first by linearizing the respective efficiency curves (which values have been carried out from average test values or from mean data available in literature) and then by applying the corresponding formulas for converting test results from European to United States format. Collector parameters Type of collector Flat plate

selective (FPC)

Evacuated tubular (ETC)

Compound parabolic

(CPC)

Parabolic trough (PTC)

Test intercept [-] 0.78 0.76 0.72 0.74 Test slope [W/m²K] 4.3 1.8 2.0 0.56 Collector orientation South South N-S N-S Collector slope 45° 45° 45° 45° Concentration ratio - - 1.5 60 General solar heating system parameters Storage volume to collector area ratio 70 litres/m² UA of auxiliary storage tank Negligible Pipe heat loss Negligible Collector-store heat exchanger Tank-side flow-rate/area: 0.015 kg s-1 m-2

Heat exchanger effectiveness: 0.7 Load heat exchanger effectiveness 0.7

Table 2.7.8: Values of the main parameters adopted in the calculation with PHIBAR f-Chart method.



2.7.5 Results

In the present analysis an effective floor area of 100 m² and 200 m², for residential and office building respectively, have been considered.

2.7.5.1 Residential building application

Tables from 2.7.9 to 2.7.12 show the results obtained for three reasonable fraction of consumed primary energy to be saved, applied to the reference residential building.

SITE #1 – PALERMO Application Residential – Floor area: 100 m² Annual loads Cooling: 7627 KWh - Heating: 5916 KWh PE-consumption 12557 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)

Type of collector FPC ETC CPC PTC PTC FPC ETC FPC ETC Specific collector area needed to attain the desire d primary energy saving [m²/m² floor area ]

60% 0.16 0.08 0.24 0.08 0.08 0.14 0.08 0.13 0.08 70% 0.21 0.10 0.30 0.09 0.10 0.18 0.10 0.16 0.09 % of PE-saving 80% 0.25 0.12 0.40 0.11 0.13 0.22 0.12 0.20 0.11

Table 2.7.9: Palermo (IT) - Results of the analysis.

SITE #2 - ATHENS Application Residential – Floor area: 100 m² Annual loads Cooling: 6717 KWh - Heating: 7850 KWh PE-consumption 13258 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)


60% 0.19 0.10 0.28 0.09 0.10 0.17 0.10 0.16 0.10 70% 0.24 0.12 0.37 0.11 0.12 0.21 0.12 0.19 0.12 % of PE-saving 80% 0.31 0.14 0.46 0.14 0.14 0.26 0.14 0.24 0.14

Table 2.7.10: Athens (GR) - Results of the analysis.

SITE #3 - MILAN Application Residential – Floor area: 100 m² Annual loads Cooling: 4417 KWh - Heating: 18046 KWh PE-consumption 19455 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)




60% 0.57 0.23 0.72 0.31 0.44 0.48 0.20 0.45 0.21 70% 0.80 0.35 1.08 0.49 0.78 0.63 0.26 0.72 0.32 % of PE-saving 80% 1.50 0.54 1.59 0.71 1.50 0.90 0.38 1.32 0.49

Table 2.7.11: Milan (IT) - Results of the analysis.

SITE #4 - FRANKFURT Application Residential – Floor area: 100 m² Annual loads Cooling: 1120 KWh - Heating: 22721 KWh PE-consumption 20053 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)


60% 1.25 0.52 1.11 0.49 1.12 1.16 0.48 1.18 0.49 70% 1.75 0.68 1.56 0.63 2.35 1.62 0.64 1.72 0.67 % of PE-saving 80% 2.55 0.92 2.10 0.87 - 2.35 0.85 2.41 0.90

Table 2.7.12: Frankfurt - Results of the analysis. Table 2.7.13 summarize the results obtained for 70% of primary energy saved. In the table are reported, grouped by collector technology, the values of collector areas needed for a square meter of floor area to be conditioned, together with the corresponding annual solar fractions.

Case study: RESIDENTIAL BUILDING – PE-saving = 70%

PALERMO ATHENS MILAN FRANKFURT System Configuration Ac/A floor

[m²/m²] fsol

[-] Ac/A floor [m²/m²]

fsol


fsol


fsol

[-]

H2O/LiBr–SE 0.21 0.55 0.24 0.55 0.80 0.57 1.75 0.58

ADS 0.18 0.47 0.21 0.49 0.63 0.54 1.62 0.57 FPC

DEC 0.16 0.52 0.19 0.53 0.72 0.56 1.72 0.58

H2O/LiBr–SE 0.10 0.55 0.12 0.55 0.35 0.57 0.68 0.58

ADS 0.10 0.47 0.12 0.49 0.26 0.54 0.64 0.57 ETC

DEC 0.09 0.52 0.12 0.53 0.32 0.56 0.67 0.58

H2O/LiBr–DE 0.09 0.68 0.11 0.67 0.49 0.62 0.63 0.59 PTC

NH3/H2O 0.10 0.66 0.12 0.70 0.78 0.81 2.35 0.90

CPC H2O/LiBr–DE 0.30 0.68 0.37 0.67 1.08 0.62 1.56 0.59

Table 2.7.13: Required collector area per square meter of conditioned floor area and respective annual solar fraction in residential building.

Figure 2.7.5 compares graphically the required specific collector areas obtained for the different configurations taken into account. Furthermore, in the same graph are reported the yearly values of global and 1-axis tracking solar irradiance on tilted plane.



Residential building - PE-saving = 70%

0,00

0,20

0,40

0,60

0,80

1,00

1,20

1,40

1,60

1,80

2,00

2,20

2,40

2,60

PalermoR = 1.29

AthensR = 0.86

MilanR = 0.24

FrankfurtR = 0.05

Sol

ar fi

eld

spec

ific

area

[m

²/m²

floor

are

a]

0

200

400

600

800

1000

1200

1400

1600

1800

2000

Yearly solar irradiance on tilted plane [K

Wh/m

²]

FPC + LiBr single effect

FPC + Adsorption

FPC + DEC

ETC + LiBr single effect

ETC + Adsorption

ETC + DEC

PTC + LiBr double effect

PTC + Ammonia/Water

CPC + LiBr double effect

Global irradiance

Direct irradiance (1-axis tracking)

Figure 2.7.5: Required specific collector area for different technologies and climates grouped by

collector technology. R represents the cooling/heating load ratio.

2.7.5.2 Office building application

Analogously, in this sub-paragraph, Tables from 2.7.14 to 2.7.17 show the results obtained for the same fractions of consumed primary energy to be saved, applied to the office-building case study.

SITE #1 - PALERMO Application Office – Floor area: 200 m² Annual loads Cooling: 18651 KWh - Heating: 12366 KWh PE-consumption 28956 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)


60% 0.19 0.10 0.28 0.08 0.10 0.17 0.09 0.15 0.09 70% 0.24 0.12 0.37 0.10 0.12 0.21 0.11 0.19 0.11 % of PE-saving 80% 0.29 0.14 0.50 0.12 0.15 0.25 0.13 0.23 0.13

Table 2.7.14: Palermo (IT) - Results of the analysis.



SITE #2 - ATHENS Application Office – Floor area: 200 m² Annual loads Cooling: 17868 KWh - Heating: 18722 KWh PE-consumption 33469 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)

Type of collector FPC ETC CPC PTC PTC FPC ETC FPC ETC Specific collectors area needed to attain the desir ed primary energy saving [m²/m² floor area ]

60% 0.24 0.12 0.36 0.11 0.12 0.21 0.11 0.19 0.11 70% 0.31 0.15 0.46 0.13 0.15 0.26 0.14 0.24 0.14 % of PE-saving 80% 0.39 0.18 0.59 0.17 0.18 0.32 0.17 0.30 0.17

Table 2.7.15: Athens (GR) - Results of the analysis.

SITE #3 - MILAN Application Office – Floor area: 200 m² Annual loads Cooling: 11472 KWh - Heating: 44227 KWh PE-consumption 48328 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)


60% 0.69 0.28 0.84 0.38 0.53 0.56 0.24 0.54 0.25 70% 0.95 0.42 1.23 0.60 0.91 0.73 0.31 0.84 0.38 % of PE-saving 80% 1.59 0.65 1.78 0.85 1.63 1.01 0.45 1.46 0.59

Table 2.7.16: Milan (IT) - Results of the analysis.

SITE #4 - FRANKFURT Application Office – Floor area: 200 m² Annual loads Cooling: 2662 KWh - Heating: 60861 KWh PE-consumption 53380 KWh Solar cooling system




AHE-Cycle

Adsorption chiller

(H2O/Si-gel)

DEC (Desiccant

wheel)


60% 1.53 0.65 1.38 0.63 1.43 1.43 0.60 1.53 0.64 70% 2.19 0.86 1.95 0.83 3.03 2.05 0.81 2.18 0.86 % of PE-saving 80% 3.20 1.18 2.60 1.08 - 2.95 1.10 3.20 1.18

Table 2.7.17: Frankfurt - Results of the analysis. Table 2.7.18 summarize the results obtained for 70% of PE-saved. In the table are reported, grouped by collector technology, the values of collector areas needed for a square meter of floor area to be conditioned, together with the corresponding annual solar fractions.



Case study: OFFICE BUILDING – PE-saving = 70%

PALERMO ATHENS MILAN FRANKFURT System Configuration Ac/A floor

[m²/m²] fsol


fsol


fsol


fsol

[-]

H2O/LiBr–SE 0.24 0.54 0.31 0.55 0.95 0.57 2.19 0.58

ADS 0.21 0.47 0.26 0.48 0.73 0.53 2.05 0.57 FPC

DEC 0.19 0.52 0.24 0.53 0.84 0.56 2.18 0.58

H2O/LiBr–SE 0.12 0.54 0.15 0.55 0.42 0.57 0.86 0.58

ADS 0.11 0.47 0.14 0.48 0.31 0.53 0.81 0.57 ETC

DEC 0.11 0.52 0.14 0.53 0.38 0.56 0.86 0.58

H2O/LiBr–DE 0.10 0.69 0.13 0.67 0.60 0.62 0.83 0.59 PTC

NH3/H2O 0.12 0.65 0.15 0.69 0.91 0.81 3.03 0.90

CPC H2O/LiBr–DE 0.37 0.69 0.46 0.67 1.23 0.62 1.95 0.59

Table 2.7.18: Required collector area per square meter of conditioned floor area and respective annual solar fraction in office building.

Finally, Figure 2.7.6 compares the required specific collector areas obtained for the office-building case study.

Office building - PE-saving = 70%

0,00

0,20

0,40

0,60

0,80

1,00

1,20

1,40

1,60

1,80

2,00

2,20

2,40

2,60

2,80

3,00

3,20

PalermoR = 1.51

AthensR = 0.95

MilanR = 0.26

FrankfurtR = 0.04

Sol

ar fi

eld

spec

ific

area

[m

²/m²

floor

are

a]

0

200

400

600

800

1000

1200

1400

1600

1800

2000

Yearly solar irradiance on tilted plane [K

Wh/m

²]

FPC + LiBr single effect

FPC + Adsorption

FPC + DEC

ETC + LiBr single effect

ETC + Adsorption

ETC + DEC

PTC + LiBr double effect

PTC + Ammonia/Water

CPC + LiBr double effect

Global irradiance

Direct irradiance (1-axis tracking)

Figure 2.7.6: Required specific collector area for different technologies and climates grouped by

collector technology. R represents the cooling/heating load ratio.



3. SOLAR DESALINATION

3.1 OVERVIEW OF TRADITIONAL DESALINATION TECHNOLOGI ES IN EUROPEAN COUNTRIES

According to the data reported in the most recent IDA desalting plants inventory (/Wan04/), at the end of 2003 the total seawater desalination capacity worldwide, both installed and contracted, was 37.8 million m³/d, coming from 10,350 plants. A slight weakening in the market trend has occurred during the last two-year period, in comparison with the previous one, which represented the best result as ever. In any case, the capacity has increased of 4,8 million m³/d since December 2001, corresponding to some 15% of the full amount. Despite these impressive figures, the application of desalination technologies still involves a limited area. It is to be noticed that nearly 50% of the total capacity is concentrated in 3 countries only (Saudi Arabia, USA, UAE), while considering the leading 15 countries altogether a good 80% of the worldwide installed capacity is reached. This not uniform distribution is due to the peculiarity of desalination market, which is mainly influenced by the water supplying cost. The economic barrier makes desalination technologies attractive for installations in areas, where the natural renewable water resources are not sufficient to satisfy the needs of the people and the development of the society. Really, in such conditions, drinking and process water has to be transported, in some cases over long distances through pipelines or by ship, and treated, with a marked rise in cost. As a result, the water produced in desalting plants can economically compete with that coming from other traditional water supplying methods. This situation is already being carried out in arid Middle East regions or small islands where the population lives from tourism. In effect, in countries such like UAE, Virgin islands, Antilles, Qatar, Gibraltar, Kuwait a desalination capacity of 1,000 l/d per capita or even more is installed, able to meet almost half of the minimum standards fixed by WHO, when all uses (residential, industrial and agricultural) are included.

COUNTRY Total capacity [10³ m³/d]

Per capita capacity [l/d]

Austria 26.2 3.2 Belgium 10.2 1.0 Denmark 26.1 4.9 Finland 1.5 0.3 France 219.3 3.7 Germany 289.9 3.5 Greece 76.7 7.3 Ireland 12.5 3.3 Italy 673.7 11.7 Netherlands 238.2 14.9 Portugal 13.5 1.3 Spain 2419.0 61.2 Sweden 4.6 0.5 United Kingdom 295.9 4.9 TOTAL 4307.1 11.4

Table 3.1.1: Total and per capita installed desalination capacity in EU-15 countries in December 2003.

Vice versa, in most cases, renewable water resources shortage is not a very critical issue for European countries and as a result desalination technologies are not significantly widespread.



The data reported in Table 3.1.1 show that in EU-15 countries little more than 10% of worldwide desalination capacity is installed and the daily water production is normally of few litres per capita. The main exceptions are Spain, Netherlands, Italy, and Greece, due to the presence of large coastal regions and numerous small islands. In particular it is to be remarked the expansion of these technologies in Spain: installed capacity has been more than quadrupled in six years only, growing from the 529,900 m³/d (/Wan98/) of 1997 to the 2,419,000 m³/d of 2003. Thanks to this outstanding development Spain has become the fourth country in the world for installed capacity and succeed in covering an important fraction of national fresh water requirements for residential use. In Table 3.1.2 the contribution of the major processes to total installed capacity is shown. The data are limited to desalting plants rated at 700 m³/d or more, which roughly constitutes the minimum size for applications on industrial scale, as for as thermal processes are concerned. On the other hand, it can be observed that more than 90% of installed capacity is included within this range, due to necessity of producing large quantities of fresh water, curbing the final cost as much as possible. Therefore, it appears to be convincing to assume the data in Table 3.1.2 as representative of all seawater desalination installations. In this regard, it is to be stated that, according to the classification adopted by the IDA Inventory, desalting plants based upon nanofiltration process are also included in the reverse osmosis (RO) category; their contribution to the total capacity relevant to this technology is however negligible. Similarly multi-effect evaporation (ME) and both mechanical and thermal vapour compression (VC) plants are arranged in a single group, since often such processes are applied in combination. The share of each technology is not simple to distinguish: a reasonable approximation is to divide the total capacity in half between ME e VC. This assumption has been followed for all performed evaluations.

COUNTRY # of plants

Capacity [10³ m³/d]

% total capacity

RO [%]

ME/VC [%]

MSF [%]

ED [%]

Energy consumption

[GWh/y] Austria 5 18.3 70.0 100.0 0.0 0.0 0.0 47.3 Belgium 4 6.4 62.9 84.4 15.6 0.0 0.0 20.3 Denmark 5 25.0 96.1 96.8 3.2 0.0 0.0 48.1 Finland 0 0.0 0.0 0.0 0.0 0.0 0.0 3.3 France 24 195.6 89.2 92.8 4.2 0.5 2.5 403.0 Germany 80 229.8 79.3 77.9 14.9 6.8 0.4 640.0 Greece 29 53.4 69.6 36.2 21.2 10.9 31.7 150.1 Ireland 5 7.9 63.0 77.8 0.0 0.0 22.2 18.4 Italy 132 609.5 90.5 25.9 17.2 41.5 15.4 2125.1 Netherlands 27 221.7 93.1 48.1 20.3 31.2 0.4 732.6 Portugal 6 11.0 81.6 100.0 0.0 0.0 0.0 24.4 Spain 287 2323.3 96.0 87.7 2.9 3.8 5.6 4558.6 Sweden 2 1.5 33.1 100.0 0.0 0.0 0.0 8.3 United Kingdom 77 229.0 77.4 83.3 4.0 4.0 8.7 550.0 TOTAL 683 3932.4 91.3 74.7 7.2 11.2 6.9 9329.6

Table 3.1.2: Proportion of processes for desalting plants rated at 700 m³/d or more in EU-15 countries

As for the proportion of the different processes, Table 3.1.2 shows that reverse osmosis (RO) is by far the dominant desalination technology in European countries. This is an absolutely opposite situation in comparison with Middle East, where multi-stage flash (MSF) covers nearly 80% of total installed capacity. General reasons of this countertendency in EU-15 are the lesser availability of fossil fuel, the more reduced capacity per unit which makes it costly water



produced by thermal processes, the presence of an adequate electric grid, the less critical characteristics of raw water. Finally, a rough estimation of electrical energy consumption for water desalination in EU-15 is presented. The assessment has been carried out assuming the same conversion criterion from thermal, if any, to electrical energy and the same average values of energy need per m³ of produced water for the different processes given in /Wan04/. In contrast to the remarkable consumptions described above, renewable energies are still exploited to a negligible extent to power desalting plants: actually the installed capacity worldwide is little more than 9,000 m³/d coming from 97 plants, which reach a significant size (over 100 m³/d) in 12 cases, exceeding 700 m³/d in 4 cases only. In effect these installations usually pursue demonstrative aims, while in a limited number they are utilized to satisfy the municipal drinking water needs of small communities. The solar source (photovoltaic, solar thermal collectors, solar pond, solar still) is by far the most used, driving 84 plants but normally the capacity per unit is extremely modest; as for the remaining plants, 11 of them are powered by wind and 2 of them by geothermal energy.

LOCATION COUNTRY Capacity [m³/d] Process Renewable energy

Milos Greece 1800 ME GEOTHERMAL Syros Greece 960 RO WIND Kimolos Greece 100 ME GEOTHERMAL Almeria Spain 72 ME SOLAR Cadarache France 60 RO SOLAR Gran Canaria Spain 50 MVC WIND Berken Germany 20 MSF SOLAR Patmos Greece 20 OTHER SOLAR Planier France 12 RO WIND San Nicola Italy 12 RO SOLAR Süderoog Germany 6 RO WIND Kimolos Greece 6 OTHER SOLAR Megisti Greece 6 OTHER SOLAR Symi Greece 6 OTHER SOLAR Aegina Greece 5 OTHER SOLAR Fiskardo Greece 5 OTHER SOLAR Kionion Greece 5 OTHER SOLAR Symi Greece 5 OTHER SOLAR Thirasia Greece 5 RO WIND Bari Italy 5 MSF SOLAR Marettimo Italy 5 RO SOLAR Aegina Greece 4 OTHER SOLAR Nisyros Greece 4 OTHER SOLAR Lavrio Attiki Greece 3 RO SOLAR Lavrio Attiki Greece 3 RO WIND Salamis Greece 2 OTHER SOLAR Las Marinas Spain 2 OTHER SOLAR

Table 3.1.3: Desalting plants powered by renewable energies in EU-15 countries.

A slightly improved situation occurs in EU-15, where about 35% of world total desalination capacity powered by renewables is installed. In Table 3.1.3 the complete list of European plants



reported in IDA inventory is given. As it can be observed, the largest number of applications is located in Greece, being especially suitable for fulfilling the fresh water requirements of small islands, which are plentiful in this country. In fact the main European plants powered by renewables in general and by solar energy in particular are here located, respectively in Milos e Patmos Island. It is to be noted that the last one is a solar still operating since late ’60. Vice versa the absolutely largest capacity solar desalting plant has been erected at the Plataforma Solar de Almeria (PSA), southern Spain, in the framework of the activity of research, development and testing on concentrating solar technologies. The state of application of traditional desalination technologies in European countries can be summarized by the following focal points:

� a remarkable desalination capacity is installed in Spain only, where the market shows an impressive growth;

� in few other countries, such as Netherlands, Italy, and Greece, there’s a significant utilization of desalination processes, though to a more limited extent;

� the prevailing process is by far RO, which electrical energy consumption can be assumed in 4-7 KWh/m³, if seawater is the raw water, lower if brackish water is the raw water;

� almost the total of operating desalting plants has a capacity of 700 m³/d or more, which is able to meet the residential fresh water needs of a community of 4,000 people at least with a minimum installed power around 150 KW;

� the equivalent electrical energy consumption in EU-15 for desalting water is of the order of magnitude of 104 GWh per year.

At the present time renewable energies contribution in the considered countries is really modest, since it does not even reach 0,1% of total installed desalination capacity. Specifically solar energy, considering all applicable technologies, delivers the insignificant figure of 247 m³/d, coming from demonstrative plants or designed to fulfil the municipal drinking water requirements of villages, mostly located in small Greek islands. _________________________________ Finally, some interesting remarks about the current trends in market are reported below for some countries. COUNTRY Some information about market trend

GREECE In Greece, there is not an organized market for desalination systems, as denoted by the common use of the term. Nevertheless, there is an increasing number of installations, referring mainly to RO technology. Most of them are located to the islands, servicing small communities or tourism industry. Since most of these installations are EC funded projects, there are not typical representative market data. However, as regards the dominant technology of RO, the referred market and technology data are the typical worldwide data. A special case, but having no connection to the NEGST project, is the wide use of desalination systems in the marine sector. Due to the international character of this sector though, it is not possible to extract market data for Greece. During last years, there is an R&D effort to develop thermal driven systems, referring mainly to the existence of a strong local solar thermal industry.

ITALY Italy has been interested in seawater desalination since ’60, when the first applications on industrial scale took place. Thanks to this long tradition, above all in the field of thermal processes, an Italian company, Fisia Italimpianti, is still the world leading manufacturer of MSF plants and in the list of top manufacturers several further national companies (Snam Progetti, Ansaldo, Fantuzzi Reggiane, etc.) hold a prominent position. Despite the level of excellence achieved in desalination industry, the application of these technologies is not so extensive on the national territory. According to the data reported in the most recent IDA inventory, in December 2003 a total capacity of about 674,000 m³/d was installed, placing

th



Italy in the 11th ranking worldwide, but corresponding to only 12 l/d per capita. Moreover the increase in installed capacity during the last six years period has been less than 30%, quite modest in comparison with other countries affected by fresh water shortage problems. Obviously desalting plants are mostly located in Southern regions, with a peak in concentration in Sicily and especially in the numerous small islands near Sicily. This limitation to specific areas of the country is due mainly to the economic factor: in general the production cost is very high compared to the typical supply charge, being water highly subsidized by the government. Therefore the development of the desalination market would require mid-long term supporting policies, not sporadic actions in consequence of drought periods. A further barrier is the elevated energy consumption per m³ of desalted water, which is particularly penalizing in a scenario of national energy generation inadequate to fulfil the internal demand. This could be the stimulus for a wider exploitation of renewable energies, which currently give a negligible contribution. Finally the impact of emissions from desalination processes on the environment could be a critical aspect to make the erection of large capacity plants tolerable by local communities. With regard to the proportion of the different desalination processes, the peculiarity of Italy is the absence of a really dominant technology. Actually, due to the presence of a solid national industry, MSF is still the most utilized process with more than 40% of total capacity, but the rising attention towards energy efficiency is leading to a significant expansion in application of RO. The share relevant to this process has grown from little more than 20% in 1997 to nearly 26% in 2003. Also ME/VC and ED give an important contribution of around 17% and 15% respectively, as a result of the testing of different solutions according to the variability along the national territory in raw water characteristics, access to energy, qualification and quantification of the water demand, etc.

SPAIN In Spain there are a lot of desalination plants due to some reasons; could be appointed out the irregular rainfall of our geography, the high water consumption in zones with irrigated agriculture and the tourist infrastructure which increases the consumption in drier seasons. So desalination of seawater (42%) and brackish water (58%) is one of the ways of meeting water demand. Most of desalination plants in Spain are situated in Levante, Murcia, Andalucia, Mallorca, and Canary Islands and in Spanish cities of the North of Africa. In 2000, the installed capacity was 700.000 m³/d in Spain (around 800 plants), while in Canary Island was 350.000 m³/d (around 300 plants) and in Murcia was 149.000 m³/d (around 90 plants). Regarding the used technology, it could be come out the use of membrane process as reverse osmosis (RO) (87%) and electro-dialysis (ED) (9%). Both of them, RO and ED, are used for brackish water desalination, but only RO compete with distillation processes in seawater desalination.

Table 3.1.4: Market trends in some EU countries (Source: country data delivered by NEGST WP5 participants).

3.2 MULTIPLE EFFECT EVAPORATION


In order to introduce the working principle of Multiple Effect Evaporation (MEE), we will introduce first the principles of single effect evaporation (/Des02/) and also the single effect direct solar evaporation (/INE99/). In Figure 3.2.1 a Single Effect Evaporation desalination process is represented. In a single effect evaporator there are two main parts – the evaporator and the condenser. In the evaporator, the seawater is sprayed in a space heated by steam produced by a conventional boiler. The sprayed water evaporates and there will be a separation between the salts and the water in the form of vapour. The vapour is conducted out of the evaporator and will go through the condenser in order to obtain the distilled water. In the condenser, the vapour is condensed by the effect of cooling, using the seawater before being distillate. Both the Evaporator and Condenser are heat exchangers. In the first one, steam is used to heat the seawater and induce evaporation. In the second, as already explained, the cold seawater is used to condense the vapour that was previously separated from the salts.



This system has very little industrial applications since its thermal performance ratio is less then one, i.e., the amount of water produced is less than the amount of heating steam used to operate the system (/Des02/). Single effect solar evaporation is also not very efficient but it is interesting to understand its working principal. In /INE99/ a summary of such systems is presented. /INE99/, is a compilation of the results of a project developed in the framework of the CYTED Programme – Sub-programme VI:

SOLCYTED. This Programme involved Portugal and Spain and countries from Latin-America. The project title is “Produccion de agua potable para pequenos consumos humanos” and in its framework different aspects of production of drinkable water for human consumption were studied. Several prototypes were developed (see /INE99/). Some of these will be described in section 3.2.2.1 of this document and are all considered low capacity production systems. In Figure 3.2.2 a schematic representation of a single effect direct solar evaporation desalination process is represented.

Figure 3.2.2: Single effect solar evaporation desalination process (Source: /INE99/)

The single effect solar still is formed by a shallow container with insulation in its bottom part and with a transparent cover with some tilt on top. In the shallow container the seawater is introduced. By effect of the solar radiation incident on the transparent cover the water is heated and it evaporates. Water vapour in contact with the cover (colder) will condense and will form a thin film of water that will slide to the lateral channels and will be collected in the form of distilled water. In these systems the energy that results from the condensation of water in the transparent cover is lost by the system. Multi effect evaporation is conceived to avoid the losses of the single effect evaporation. In Figure 3.2.3 a Multi effect evaporation system is schematically represented.

Figure 3.2.1: Single effect evaporation desalination process

(Source : /Des02/)



Figure 3.2.3: Schematic representation of a Multi effect evaporation system (Source: /Des02/- page 150)

The conventional multi effect evaporator is formed by n simple effect evaporators. In the first element, the hot steam produced by a boiler is used in a heat exchanger to evaporate de seawater and produce a first separation between water (in the form of vapour) and brine (seawater with a higher salt concentration). The vapour produced is transferred to the second element transporting the heat that will be used to produce again evaporation of the brine, which is delivered in this second element by spray nozzles as in the first case. In the last element the vapour that is not transported and used in another element, goes through the condenser joining the distillated water produced in the other elements and the condensation heat is used to preheat the seawater that goes to the first element. All this process is described in a detailed way in /Des02/. In the case of multiple effect solar evaporation, there is also the possibility to re-use the heat resulting from the condensation of vapour that, in the simple effect solar evaporation is lost through the transparent cover.

A schematic representation of this multi effect can be seen in Figure 3.2.4. The single effect or multi-effect solar evaporation described can be called direct solar evaporation, i.e., the solar distillation system is an integrated system (solar collection is performed in the same device that performs the distillation), but solar evaporation can also be considered with systems where the solar collectors are separated from the distillation unit – indirect solar evaporation . Indirect solar evaporation systems were also the object of the studies performed under the framework of CYTED project (/INE99/). These systems were always considered for remote application considering the need for a small production of water (low capacity production systems).

Figure 3.2.4: Schematic representation of multi effect

solar evaporation (Source: /INE99/)



Figure 3.2.5 is a schematic representation of a multi-effect evaporation system for small productions. Several prototypes were constructed and studied in CYTED project. In this type of multi-effect distillation unit the heat necessary to the evaporation process could be delivered by solar collectors of different types or by burning biomass (/INE99/). In section 3.2.2., examples of solar evaporation, either with single effect or multi-effect, are given.

3.2.2 Research and demonstration systems

3.2.2.1 Low capacity production systems

According to /Des02/, the single effect evaporation plants, which we can call conventional plants since they will not use Solar Energy in a direct or indirect form, are seldom used in Industry. They are mainly used in marine vessels. Single effect solar evaporation plants, although not very efficient, are used where fresh water demand is low and land is inexpensive (/Bla03/). In /INE99/, several of these single effect evaporators were studied and different prototypes are presented (see chapters III.3 and VII of /INE99/). The fact that they can be constructed in a not very expensive way and that there operation does not require very special technical skills makes them an interesting solution in some developing areas.

Figure 3.2.6 shows one of the prototypes tested in the frame of CYTED sub-programme VI. Indirect solar evaporation systems were also the object of the studies performed under the framework of CYTED project. These systems were always considered for remote application considering a small production of water. Two examples are shown in Figure 3.2.6 and Figure 3.2.7.

Figure 3.2.5: Multi-stage indirect solar evaporation

(legends in English are missing)

Figure 3.2.6: Modular solar still – El Chaco Salteño, Argentina. Produces

water for a group of families living in a rural remote area with plenty of salt water (Source: /INE99/).



In Figure 3.2.7 it is possible to see a picture of a MED-system with 6 stages, integrated heat exchanger and flat plate collector the was installed as a pilot plant at the Brasil coast by the Solar Institut Jülich (SIJ) and was designed by MÜLLER and SCHWARZER (/Mül03/). The heat transferred between the flat plate collectors and the distillation unit work by thermosyphon effect, heating the first stage of the MED with a heat exchanger. The daily production is of 50.2 l, i.e., a specific productivity of 25.1 l/m²/d.

Figure 3.2.7: MED-pilot plant of by the SIJ in Brazil (Source: /Mül03/)

3.2.2.2 Medium and large capacity production systems

Multi effect evaporation plants have been used mainly in the Chemical Industry and their first industrial use was in the sugar industry in the 19th century (/Ahk86/). The first desalination plants of the multi effect evaporation type suffered of problems like fouling, scaling and corrosion and the development of multistage flash processes stopped the development of Multi effect plants (/Des02/). Other facts that gave a larger development to MSF (multi-stage flash) evaporation are lower cost and apparently high efficiency, specially when large size plants are considered (/Bla03/). In the second half of the XX century, only small scale

Figure 3.2.6: Asymmetric CPC coupled to a multi-effect

distillation unit (Source: /Joy93/)



MED plants were built, but in the last decade, a new interest in MED plants was found and they are now technically and economically competing with MSF (/Bla03/). In medium and large MED plants it is also possible to use the heat produced by a solar system to assist the heat supply system. In this case it is also and indirect solar evaporation with the difference that the energy needed for the distillation process is not solar-only. These last systems are much more common when associating solar with Multi-stage evaporation. Several examples of these systems are referred in /Bla03/. The examples given use different collector types - flat plate, evacuated tubes and parabolic-trough concentrators. In a recent study (/Bla03/) performed in the frame work on European project – DGXII-FPV AQUASOL Project 2002-2006, a technical and economical comparison of Solar-assisted heat supply systems for a multi effect seawater distillation unit was performed considering different collector types – CPC, Parabolic trough and evacuated tube CPC. The specific results of this study are referred in section 3.2.2.5. Examples of operating plants are more easily found outside Europe, mainly in Africa and Middle-East. Some examples of operating plants are given below.

3.2.2.3 Abu Dhabi

Since 1985, in Abu Dhabi, UAE, a solar MED plant with an average capacity of 80 m³/d is in operation successfully. The 18-stage MED-plant was manufactured by Sasakura and Sanyo of Japan. (Sasakura usually builds large-scale desalination plants.) The plant consists of three subsystems: the solar collector field, the heat accumulator and the sea water evaporator. A simplified schematic of the plant is shown in Figure 3.2.8 and a pictorial view in Figure 3.2.9.

Figure 3.2.8: Plan of the solar thermal driven desalination plant at Abu Dhabi (Source: /Eln01/).

The plant is equipped with vacuum technique and chemical water pre-treatment. It is powered by 1,860 m² of evacuated tube, flat plate collectors. In order to ensure that the evaporator can run 24 hours per day during sunny days, a thermally stratified heat accumulator with a capacity of 300 m³ is incorporated in the design to provide the thermal energy required during the night-time. The electrical energy required by the different pumps is provided from the main grid. The 18 effects are arranged above each other. The specific heat consummation is 49 kWh/m³ distillate at a sea water temperature of 35 °C and a hot water temperature of 99 °C (continuous operation). The plant is in an excellent condition and maintained by two fulltime technicians (/Eln01/).



Figure 3.2.9: Picture of the desalination plant at Abu Dhabi.(Source: /Eln01/)

Report of a new plant with a much higher capacity is referred in /Bla03/ to be reported by Vermey, J.W. (2003).

3.2.2.4 Plataforma Solar de Almeria (PSA)

Projects in the field of Solar thermal desalination system have been developed In Plataforma Solar de Almeria (PSA), in Spain, since the 80’s. These projects have been develloped by the Spanish research institute CIEMAT - Centro de Investigaciones Energeticas, Medioambientales y Technologicas and the DLR - Deutsche Forschungsanstalt für Luft und Raumfahrt e.V. In /Joy93/ are reported the results of the project developed between 1987 and 1993. The project had two phases:

� Phase I – to study the reliability and technical feasibility of solar thermal energy in seawater desalination

� Phase II – to develop an optimised solar desalination system by implementing in the system initially installed at the PSA those improvements that could make it more competitive with conventional desalination systems”

In this project a Multi-effect Distillation plant was installed in 1988. This plant was formed by a 14-effect MED plant, a parabolic trough solar collector field and a thermocline thermal energy storage tank (115 m3). The collector field as an area of 2672 m2 and is formed by parabolic trough collectors manufactured by ACUREX (USA), model 3001 (see Figure 3.2.10).



Figure 3.2.10: Parabolic Trough Collectors. PSA, Spain (Source: /Zar94/)

The collector field was an already existing field in PSA and was not specially installed for the MED system. The energy delivered was much higher (6.5 MWht/day) then what was needed – 3 MWht/8 hours daily operation. The nominal output of this installation was of 3 m3/h. The avaluation of this plant showed a performance ratio higher then 9 (i.e. number of kgs of distillated produced by 2300 kJ heat input). From the first phase of the project it was possible to conclude that two important improvements would be:

a) the plant electrical demand could be reduced by replacing the initial hydroejector-based vacuum system with a steam ejector system

b) the plant thermal demand could be 50% reduced, by incorporating s double-effect absorption heat pump to the MED plant.”

A conventional MED plant needs to cool the final condenser with seawater in order to condense the steam produced in the last effect. The amount of cooling water required by the condenser depends on the seawater temperature. 2/3 of this cooling water is rejected back into the sea, thus wasting an important amount thermal energy (~110 kWh at 35 °C). To eliminate this waste of energy, a double-effect absorption heat pump was coupled to the final condenser. The heat pump delivers 200 kW of thermal energy at 65 °C to the MED plant. The desalination process in the plant evaporator body uses only 90 of the 200 kW, while the remaining 110 kW are recovered by the heat pump evaporator at 35 °C and pumped to usable temperature of 65 °C. For this, the heat pump needs 90 kW thermal power at 180 °C. The energy consumption of the desalination system is thus reduced from 200kW to 90 kW. Therefore, the thermal energy consumption of the desalination system was reduced by 44% from 63 to 36 kWh/m³ and the electric consumption by 12% from 3.3 to 2.9 kWh/m³. Thus, the price could be reduced from 3.3 to 2.36 €/m³. The production capacity was about 3 m³/h at 12 h/d in a yearly average. The implementation of these improvements was carried out and concluded in 1993. With the second phase of the project, it was possible to detect several technical problems associated to the improvements introduced and to find ways to overcome these problems (see /Zar94/ for details). It was also possible, with theoretical studies performed during the project, to do a cost analysis of this technology and compare it with others as can be seen in Fig. 11. This graph can only give qualitative information since the costs are not converted to €. Taking this into account, it is



possible to conclude that Solar MED technology shows slightly higher costs then Conventional MED technology.

Figure 3.2.11: Water production cost – Comparison of MED technology (Coventional and Solar) with

Reverse Osmosis (Conventional and with PV).

3.2.2.5 AQUASOL Project

The AQUASOL project - “Enhanced Zero Discharge Seawater Desalination using Hybrid Solar Technology” - is a on-going project financed in the 5th Frame Work Programme. It is based on the experience gained in the previous projects of PSA (/Zar94/). The objective of this project is to improve the present techno-economic efficiency of solar MED systems and reduce the cost of water production. The project will develop three main technological aspects:

a) development of a stationary solar collector of the CPC type to supply heat at medium temperature (70-100ºC). The CPC collector field is coupled to a thermal storage and sill have as backup a gas boiler, allowing a 24 hour operation of the MED plant.

b) Development of a Double Effect Absorption Heat Pump (DEAHP) optimised and fully integrated within the MED process to reduce the energy input needed and to improve the overall energy efficiency of the process

c) Reduce to zero any discharge from the process by recuperating the salt from the brine. An advanced solar dryer system will be designed and developed in the project.

The choice of the CPC collector is justified in /Bla03/. The advanced solar dryer is schematically represented in Figure 3.2.12 and described in /Col03/.



A partial prototype of this dried has been tested at INETI and a final version will be constructed in Greece for production of salt from see water. Figure 3.2.13 shows a schematic representation of the set-up of the new Solar assisted distillation system to be installed at PSA, Almeria, which constitutes a demonstration installation. Taking into account the technological improvements to be obtained in the Aquasol project and considering that the MED plant will have a solar contribution of 50% to the total energy required for operation of the plant it is possible to determine the cost per square meter of solar collectors needed to obtain the same cost of production of water as a conventional MED plant (/Bla02/). The graph below shows that for a cost of Natural gas of 4.5 Euro/GJ the cost of the solar assisted MED plant should be of 125 Euro/m2.

Figure 3.2.13: Schematic representation of the set-up of the solar assisted distillation plant - AQUASOL

Project (Source: /Bla02/).

Figure 3.2.12: Advanced solar drier design (Source:

/Col03/).



Figure 3.2.14: Estimated equivalent solar hardware cost versus fossil fuel cost to obtain the same cost

than a conventional MED plant with a 50% solar contribution to the total plant energy requirement (Source: /Bla02/).

3.3 MULTI-STAGE FLASH EVAPORATION


Conventional Multiple Stage Flash (MSF) sea water desalination systems work at temperatures of 90-120°C and consist of up to 40 stages with suc cessively decreasing pressure. The capacity of installed plants ranges from 4,000 to 500,000 m³/d (/Buc04/).

Figure 3.3.1: Diagram of a multiple stage flash plant (Source: /Bur00/).

In the MSF process, seen in Figure 3.3.1, sea water is heated in a vessel called the brine heater. This is generally done by condensing steam of conventional power plants on a bank of tubes that carry seawater which passes through the vessel. This heated sea water then flows into another vessel, called a stage, where the ambient pressure is lower (operated under vacuum conditions), causing the water to boil immediately. The sudden introduction of the heated water into the chamber causes it to boil rapidly, almost exploding or flashing into steam. Generally, only a small percentage of this water is converted to steam (water vapour), depending on the pressure maintained in this stage, since boiling will continue only until the water cools (furnishing the heat of vaporization) to the boiling point.



The steam generated by flashing is converted to fresh water by being condensed on tubes of heat exchangers that run through each stage. The incoming feed water going to the brine heater cools the tubes. This, in turn, heats the feed water so that the amount of thermal energy needed in the brine heater to raise the temperature of the seawater is reduced (/Buc04/). The main advantage of the MSF process is the ease and reliability of the process. Heat exchange with the saline water does not occur through heat transfer surfaces, there is therefore no risk of reduced heat transfer by scaling. Precipitation of inorganic material may occur within the chambers, and can be reduced by applying acid or antiscalants. The top brine temperature is limited to about 110°C by the risk of scaling. Biocides may be added as well to reduce growth of bacteria; these products will not end up in the product water because of the concept of the process. MSF is also insensitive to the initial feed concentrations and to the presence of suspended particles. The product water contains about 50 ppm of total dissolved salts. Corrosion is easier to control with MSF compared to multiple effect distillation (MED), because the design is less complex.

3.3.2 Energy Consumption and Costs

The most important disadvantage of MSF is the higher energy consumption, which makes MSF a more expensive technology than MED and only economically competitive when energy costs are very low. However, MSF is still an important process for seawater desalination, although there is a clear tendency towards MED and reverse osmosis (RO) (/VdB02/). A MSF system normally consists of 15 to 25 stages. The energy efficiency increases with the amount of stages. Typical values range from 20 to 60 kWhthermal/m³. However, the enlarged exchange area and a more complex operation method with higher demand of monitoring result in higher investment costs. An enlarged temperature difference between the first and the last stage increases the energy efficiency, but an increased inlet temperature leads to growing problems with scaling, fouling and corrosion. The technology is mainly applied in big systems with a daily capacity of 4000 to 57000 m³. The MSF technology is for smaller systems not economically feasible (/Mül02/). Most of the systems for desalination used in the Arab region are parts of multipurpose systems, which work with combined heat and power plants. A steam or gas turbine generates electricity and the desalination plant (MSF/MED) uses the waste heat for the required temperature. This way less energy is required than if the heat and power were generated separately. It is however a disadvantage that the operation of the turbine (especially for counter pressure plants) and the heat generation is directed by the heat demand of the desalination plant, i.e. its water demand. Furthermore, thermal desalination plants cannot use the by-product electricity as an efficient energy source. It is possible to operate the combined heat and power plant with a MSF/MED plant (heat) in combination with a RO plant (electricity), which decreases the dependency on a specific desalination technology. This way, the more capital intensive MSF/MED plant can be operated at full load and the RO plant can react on water demand fluctuations. Further can the electricity surplus be sold or delivered to the grid or other consumers.

3.3.3 Integration of Solar Heat in Multiple Stage Flash Evaporation Plants

There are two different system concepts for integration of solar heat in multiple stage flash evaporation plants. Conventional MSF systems are operated at temperatures between 90 and 120°C, as mentioned above. However, it is also poss ible to operate at lower temperatures by directly preheating the brine. The preheated brine, leaving the heat recovery section, is heated by an external energy supply until it reaches its top temperature. The first option is to replace this external energy supply (usually steam that comes from a conventional power plant) by solar generated steam. Then the multiple stage flash process begins. The following technologies are available to produce solar generated steam:

� Direct steam generation (DSG) with medium-temperature collectors such as parabolic trough collectors (PTC)



� Pressurized collector loops using water or thermal oil as heat transfer medium. The heated heat transfer medium is then sent to a boiler which would generate the steam required by the MSF plant. Evacuated tube or parabolic trough collectors can be used for this purpose.

These solar desalination systems have been thermo-economically evaluated by Garcia-Rodriguez and Gomez-Camacho (/Gar99/). The other option is to use the solar field directly as brine heater. A small hot brine storage can be used to avoid the effect of solar irradiance transients. This thermal storage ensures that the top brine temperature is constant to avoid unstable operation of the desalination plant. Solar collectors heat the brine during daylight and a conventional energy supply drives the MSF plant during night time. The advantage of this second concept is that the operating temperatures of the collector field can be much lower and very simple collector technologies can be used such as the concrete tube collectors that were used in the MSF system in India described below.


3.3.4.1 Sun Utility Network

A study carried out by H.P. Garga in 2002 shows an example of a mobile MSF plant by Sun Utility Network (/SUN03/). This example shows that the MSF process can be downsized, with high technical expenses, and then driven by solar energy. The system uses evacuated tube collectors and operates at a process temperature of 115°C. The maximum production capacity of the system is 200 m³/d.

3.3.4.2 Thar Desert in India

A MSF plant with a fresh water capacity of 52.5 x 106 m³/a and a collector area of 11.52 km² was taken into operation in 1980 in the Thar Desert of India. The solar collectors are rectangular concrete tubes, half buried in the ground, through which seawater flows and is heated by solar energy. The heated seawater with a temperature of about 60°C is distilled in a MSF unit. Pumping of the sea water to the site and through the MSF unit is powered by 415 wind turbines each with a capacity of 200 kW. Economic analysis of the scheme shows that the MSF system is compared in favour with the existing fossil fuel fired desalination plants of the equivalent capacity (/Raj80/).

Figure 3.3.2 shows the general flow diagram on the multistage flash evaporation process and the “solar field”. As the heated seawater leaves the desalination plant, it is pumped through the “solar field” which heats it to 60°C. The “solar field” consists of tubular concrete collectors. There are 9,600 of these collectors on each side of the desalination unit each with the length of 609.6 m. Water at 60°C is flash evaporated in the desalination unit to yield 1.36 x 104 m3/hr of fresh water for a typical day in June at about 2 p.m.

Figure 3.3.2: Flow diagram of the MSF process and the “solar field”.



The solar field is the most important component of the system since it provides the energy to effect desalination and the right choice of solar collector is important. For the present scheme the solar collectors should meet the following criteria: (1) Ease of fabrication and maintenance; (2) Cost effectiveness; (3) minimum corrosion with seawater; (4) Durability and ruggedness to last the life of the plant. Several collectors and materials, which would meet the above criteria, were looked into, and finally concrete was chosen as the collector material. Apart from meeting all of the above criteria, there is an added advantage for picking it as the collector material. The clay content (a necessary ingredient of concrete) of the That soil is high (15%), and thus local raw materials can be used effectively in making these collectors, thereby reducing the cost. Figure 3.3.3 shows the cross section of the collector used in the present scheme. The present configuration was decided from strength considerations of concrete together with the need to maximize the surface area exposed to solar radiation.

Figure 3.3.3: Cross section of the solar collector.

The heated seawater from the “solar field” goes to the desalination unit for production of fresh water. The desalination unit is the regular MSF type. The MSF unit for the present scheme is a 20-stage evaporator with terminal temperature difference (TTD) of 1.67°C. The present design is simply an extrapolation on the design by Brice et al. (/Bri63/) for the production of about 1.36 x 104 m³/hr of fresh water. As can be expected the output from the desalination unit depends upon the brine temperature entering it. For different months the output of the unit has been calculated. The method for calculating the output is very well known (/Sil66/) and those methods have been used in the present case. The distillate output starts as soon as the outlet temperature from the “solar field” reaches 32.2°C. Thus for the month of June the scheme operates for 24 hr while for the month of March it works for 19 hr and for the month of January only 9 hours.

Figure 3.3.4 shows the output for various months. There is no output during the months of July, August and half of September since in these months the sky is mostly cloudy [12]. More over these months can be used for yearly maintenance, if any. The integration of figure 4 than gives us the yearly output of 5.25 x 107 m3 of fresh water. The cost of water from the present scheme compares favourably with that from a fuel fired MSF plant, the crossover taking place at about 8 years (/Raj80/).

Figure 3.3.4: Distillate outputs for different months



3.4 APPLICATIONS FOR SMALL USERS

During the last years, a significant development of small scale applications takes place, focusing mainly at RO technology. Nevertheless, thermal desalination remains an option, especially for countries with high solar potential and water shortage problems, such us the Mediterranean countries. In general, thermal desalination technology is considered simple and easy to implement, with possibility to support and maintain the system locally, even in countries with low-level technical infrastructure. Additional advantages are the existence of a strong local solar thermal industry in most Mediterranean countries and the possibility to operate the system in a multi-source (solar, waste heat, etc) and multi-use context (heat or fresh water production depending to user’s needs). When talking about solar thermal desalination technology for small users, one has to note that the main characteristics of this kind of technology are:

� the small size as regards production of water, as we do not talk for production plants � high degree of autonomous operation, with reference to a conventional source

On an application basis, and with reference to the above mentioned characteristics, interest focuses at decentralized cases, not connected to the grid, such as small communities, isolated areas, remote islands. Even though that, on a theoretical basis, there is a large number of possible combinations of thermal solar energy and desalination processes (Figure 3.4.1), in practice many of these are unlikely ever to be used, especially in small scale applications.

Figure 3.4.1: Desalination Processes coupled to solar thermal energy

More specifically, systems that operate with electricity/mechanical power, such us Thermal/Mechanical Vapour Compression (TVC/MVC), Reverse osmosis (RO), Electrodialysis (ED) Systems, are out of discussion, as small scale electricity production by solar thermal energy has been proven to be unfavourable to a small scale, though technical feasible. On the other hand, distillation systems powered by thermal energy appear to be a potential solution. More specifically, the following systems are referred:

[A] Multi-Effect Evaporation (MED) [B] Multi-Stage Flash Evaporation (MSF) [C] Solar Still [D] Humidification – dehumidification

There exist some other thermal solar desalination concepts as well, e.g. membrane distillation. These systems though are in the phase of experimental investigation, and they do not seem to be mature enough in terms of entering the solar thermal market in the near future. With reference to the distillation technologies mentioned, MED and MSF technology present operating conditions that impose use of vacuum tubes or concentration principle for the solar technology to be selected. Nevertheless, major constraints are imposed by the unfavourable



development of the multiple-stage distillation technology in small-scale, and these solutions will not be examined. The interest and, consequently, current analysis focus at the technologies of Solar Stills and Humidification-dehumidification, which appear to be proper for small scale applications, presenting the potential to be coupled with flat-plate collectors. These technologies can be examined as a promising alternative decentralized solution, demonstrating ease of construction using locally available materials, minimum operation and maintenance requirements and friendliness to the environment, with regard to the desalination, as well as the solar collector's technology. Through the concept of thermal desalination technologies, and especially those coupled to flat-plate solar collectors, collected thermal solar energy is used for the evaporation of saline or brackish water, while the fresh water recovery is performed at the condenser. Technological solutions present alternatives on the basis of:

� The way collected heat is exploited. More specifically many researchers propose the direct heating of saline water, while others propose the use of a heat exchanger. Trade-off between the above alternatives, refer, on one hand, to the significant restrictions on the materials of the collector, and, on the other hand, to the temperature reduction on the heat exchanger.

� Recovery or not of condensation heat that can be used for the pre-heating of saline water, considering the significant increase to the system’s efficiency.

As regards the solar part of the system, interest focuses at the referred issue of materials choice, while another issue imposed is the achievement of optimal efficiency by solar collectors in high temperatures of operation. This parameter is considered crucial for the evaporation-condensation desalination processes, as the provision of heat in temperatures of 85-95oC, acts quite positively to the requested evaporation process of saline water. In fact, this need appears to be more emerge in the case of direct evaporation (humidification-dehumidification process), as in the case of solar stills operation temperatures are lower due to the high thermal losses imposed by the configuration.

3.4.1 Solar still

Solar stills (conventional greenhouse type) appear to be simple and environmental friendly installations, thus presenting significant advantages for their use in decentralized areas. Their main disadvantage is the low output in distilled water in comparison with other desalination systems. However, it has been proven that an increase of saline water temperature leads to significantly higher outputs. This can happen by the coupling of a conventional solar still with a solar collector field and hot water storage tank (active solar stills), thus making a hybrid solar desalination and water heating system which can provide simultaneously distilled water and hot water.

3.4.1.1 Conventional Greenhouse type Solar Stills

- Principle of operation Conventional solar stills, in their simplest form, constitute of a basin, containing the water to be evaporated, as well as a transparent cover, letting penetration of solar radiation (Figure 3.4.2). The basin is covered with a thin material film, of black colour or simply painted black, in order to maximize absorptance of solar radiation.



The principle of operation is very simple. In the free space between the cover and the basin, due to the temperature difference between the water (high temp.) and the cover (low temp.) upward streams of vapour and air are developed. In fact, the temperature of the air-vapour mixture becomes higher as distance to the basin decreases. During this motion of upward streams, and especially through the contact to the cover, condensation of part of the vapour occurs. The remaining part of the mixture gets colder and moves to the basin surface. Condensed water is

formed as a film in the cover, flowing to the collection channels aside of the still. It is evident that as temperature difference between the cover and the basin-water surface increases, the creation of upward streams will be more intense, thus the production of water higher. Nevertheless, higher temperature imposes higher thermal losses of various parts of the installation. Indicative values for water temperature between 50-70°C are referred, reaching 80°C at the daytime of high radiation, while the latter is dependent to the water layer thickness (thinner layer leads to higher temperatures). Reduction to the thickness/depth of the water layer in the basin is expected to improve the productivity of the still, mainly due to the higher basin temperature. One has to note though that under this situation, the remaining water film would chill very quickly by the night-time, thus the evaporation would stop. In case the conditions let evaporation during night-time, as a consequence of the stored heat into the water-layer during the day, production would continue during the night as well. In this case, the installation might operate as a heat storage unit, thus the adjustion of the water layer thickness would permit heat storage through the system. - Construction Materials With reference to construction, and the materials used, the following can be pointed out:

� the basin is usually covered either with black paint, elastic carpets or plastic material, while one has to note the use of linoleum carpet ingrained with asphalt

� the cover is made of glass or special treated plastic (for maximizing the flow of condensate on the cover), in order to ensure high transmittance, and low absorptance of solar radiation, joint to satisfactory mechanical properties

� special care has to be taken for the insulation, especially beneath the seawater evaporator basins in order to reduce ground heat losses

- Alternative configurations The slope of the cover of greenhouse solar stills might be single or double. What has to be noted is that on the basis of motion of the sun, in different seasons and site locations the maximum radiation may be higher for a double slope still and the performance may also be better. On the other hand, a single slope still has less convection and radiation losses and the shaded region may be utilized as an additional condensation surface. A practice that increases performance of the system refers to the increase of the basin temperature or decrease of the cover temperature, or both. The proposed arrangements refer to a double glass cover, on a feed back flow and counter flow. Under the context of increasing performance for a solar still, other simple practices are referred as well, such as increasing the wettability of the glass cover inside surface (special treatment), adding a secondary condenser, putting black dye to the water, etc.

Figure 3.4.2: Greenhouse type Solar Still



The fact that a reduction in the depth of brine in the basin of the single effect type still improves the productivity of the still, mainly due to the higher basin temperature, led to the construction of a wick-type collector-evaporator. The advantage of the wick is to keep the brine as shallow as possible while avoiding dry spots. The above configurations refer mainly to single-type stills. The alternative is to increase the number of stages, thus utilizing the latent heat of condensation. The re-utilization of latent heat of condensation, for further distillation, can be carried out as double-effect

distillation. When more than two stages are involved, this is generally known as a multi-effect distillation system. The additional production resulting from the multi-effect still compared with that from the simple solar still should be justified, however, with the additional cost incurred in the more complicated multi-effect still. In this configurations, and for example for a double basin system (Figure 3.4.3), it has been proven that the process of flowing water over the glass cover has a good effect on the upper basin distillate output, and does not affect the lower basin.

3.4.1.2 Active solar stills – Coupling to flat-plate solar collectors

As stated above, increasing of the temperature difference between basin-water and the cover is expected to increase productivity of the plant. This can happen if the plant is connected to a heat source. In this report interest focuses at the solar energy coupling. The temperature level of the heat source enables the use of flat-plate collectors (or of solar ponds). The heat supply by the solar field to the solar still might be direct or indirect. In the first case, the saline water is circulated between the still and the collector with a small pump. The heat derived from the collector is directly supplied to the still and increases the rate of evaporation of the still. Alternatively the heat driven from the collector could be indirectly supplied to the still through a heat exchanger and a storage tank placed just below the basin (Figure 3.4.4). The overall system efficiency will be reduced due to the chances for lower temperature heat supply to the still and more energy losses. However, such an indirect still-collector system will protect the collector from corrosion and scale deposits caused by saline water. The indirect configuration leads to an increasing performance, as stored solar heat in the storage tank might be used for production at the night time, also.

Figure 3.4.4: Active solar still – Indirect coupling to solar collectors

Figure 3.4.3: Double basin solar still



In general the combination of solar thermal heat source with a solar still leads to higher performance of the system. Under proper planning of the system a 100% increase of the produced water quantity / day might be achieved (see paragraph 3.4.1.3). An interesting fact for active solar still systems is their capability to provide simultaneously distilled water and hot water (hybrid systems from the users point of view). Thus, they present a high level of utilization, especially if one considers the seasonal character of water demand, being higher at the period around summer, while in the winter hot water is also a concern. The case of active multiple stage solar stills, presents interest on a theoretical basis. Nevertheless, experiments have shown that for multiple stage, there is a demand for high temperature operating conditions in order to present efficient evaporation and condensation. Thus, the use of flat-plate collectors is not suggested, as the efficiency of the solar field will be poor at the operation temperatures required. The solution of vacuum tubes, joint to the eventual technological complexity of multiple stage solution, is considered rather marginal, with reference to the decentralized character of these solutions. In addition, economical terms become unfavourable, given the scale of application.

3.4.1.3 Performance characteristics of solar stills

Solar stills are applications addressed to small users, as their water production is low. In terms of efficiency, for conventional solar stills, an average value of 25-40% is referred for the cold months, while in the summer the respective value is 30-60%. This figure denotes the thermal energy used for evaporation (and finally used for the production of m kg of condensed water) divided to the incident to the system solar energy. On an approximate method, based upon the general principle of solar still operation, the performance of a conventional solar still might be calculated by the following simplified relation:

evapQ

AHnP

⋅⋅=

where :

n : overall efficiency of the still H : solar energy incident to the greenhouse area (MJ/m²) A : aperture area of the greenhouse (m²) Qevap : latent heat of vaporisation of water (2.26 MJ/kg)

Thus, with the assumption of typical average values for solar radiation and solar still efficiency, calculated daily production of water is 2.5 lt/m², referred as a general rule of thumb value for conventional solar stills. Respectively, for the case of active solar stills, a value of 6 lt/m²/d is referred. This value is verified through relevant research work. More specifically, experimental and theoretical results predict a 100% output increase in the case of active solar stills, comparing to passive systems, under the proper design. Of course this output is achieved under the assumption that there is not draw-off for hot water demand.

3.4.1.4 Techno-economical characterization of solar stills – Case studies

As it has been pointed out through the analysis, solar stills are a promising decentralized solution, as they present ease of construction, using locally available materials, and minimum operation and maintenance requirements, joint though to a low output in distilled water in comparison with other desalination systems, and considerable space requirements. Under this context, during 60’s and 70’s a lot of installations have been developed through out the world, referring to capacities less than 20m³/d mainly. Characteristic is the case of single



basin solar stills, which were constructed at that period on four Greek Islands to provide small communities with fresh water. Their capacity ranged from 12-40 m³/d of distilled water with a black surface area 2008-8600 m². One of the main problems that operation of various installations has demonstrated is the presence of deposits of scale and corrosion which deteriorate the unit’s performance, caused by the direct contact of the heating element (solar absorber) and saline water. Coupling to a solar field of flat–plate collectors increases significantly the productivity of such systems (6 lt/m²/d). Nevertheless, in the case of direct heat supply systems, flow of saline water through the collectors leads to significant restrictions on the materials of the collector, thus imposing a topic for research. On a level of techno-economical analysis, in the case of active solar stills, the installation cost increases as well, burdened by the cost of the solar-field. Thus, one has to study the trade-off between cost and performance increase. This analysis should take into account the potential of active solar stills to produce hot water as well. This benefit might lead to a high degree of system utilization, as in many areas, the demand for water is high in the summer months, while in the winter this demand decreases, and the potential of hot water availability might be more important.

3.4.2 Humidification-dehumidification

While conventional solar stills have inherently a major problem of energy loss in the form of latent heat of condensation of water, solar desalination processes based on the humidification - dehumidification principle, lead to a significant improvement in the efficiency of solar desalination units. In addition, the use of a working fluid (air), separating heating element (absorber) and saline water limits to a great extent problems by the presence of scale deposits and corrosion. The humidification-dehumidification technique is especially suited for seawater desalination when the demand for water is decentralized. Several advantages of this technique are referred, such us flexibility in capacity, moderate installation and operating costs, simplicity and the possibility of using low temperature energy (solar, geothermal, recovered energy or cogeneration). In this report, interest focuses at systems coupled to solar energy, and more specifically to flat-plate collectors.

3.4.2.1 Principle of operation

The basic principle of humidification – dehumidification desalination systems is the humidification of ambient air through heated seawater (seawater evaporation) and condensation of water vapour from humid air by contact with a cooling surface. The vapour carrying capability of air increases with temperature, i.e.1 kg of dry air can carry 0.5 kg of vapour and about 670 kcal when its temperature increases from 30°C to 80 °C. Generally, the condensation occurs in a heat exchanger in which salt water is preheated by the latent heat of recovery. The process takes place at ambient pressure and at temperatures usually between 50oC and 85°C, thus the use of flat-plate collectors is feasible, ensuring high levels of performance. What has to be noted is that the availability of heat in temperatures 85-95°C would act quite positively to the evaporation process, thus an issue is imposed, regarding research towards the development of high-performance flat-plate collectors in the specific operating conditions. As in the case of solar-stills, the heat supply by flat-plate collectors might be direct (direct pass of saline water through the collectors) or indirect (heat exchanger presence). In the case of indirect systems, the use of a storage tank is also possible. Potentially, heat storage could ensure the operation of the system for a period exceeding the day-time. For the process of humidification-dehumidification, the term multi-effect is used. This term does not impose multiple stage operation, although this is feasible, but denotes the ratio of heat input to heat utilized for distillate production (GOR>1, gained output ratio).



3.4.2.2 Types of systems- Performance Characteristics

Two types of MEH (Multi Effect Humidification) are referred: open-air/closed-water cycle and open-water/closed-air cycle. - MEH – open- water/closed-air In these systems (Figure 3.4.5), heat is recovered by air circulation between a humidifier and a condenser using natural or forced draft circulation. The saline water feed to the condenser is preheated by the evolved latent heat of condensation of water, which is usually lost in the single-basin still. The saline water leaving the condenser is further heated in a flat-plate solar collector and then sprayed over the packing in the humidifier. Resulting brine water is rejected. There have been developed several pilot MEH units – open-water/close-air. With reference to the results of the respective studies/experiments, one may point-out the following :

� The condensation heat recovery might reach 70% of the evaporation heat (GOR:3-4.5), while a production of 12 L/m2/d without thermal storage can be achieved

� Increasing of the water temperature at the inlet to the humidifier of the MEH unit, and air circulation is essential for raising the performance of the system

� Natural as well as forced air circulation systems are referred. For the latter, although the higher circulation of air (at least up to a limit) seems to increase the performance, one has to consider the energy consumption of the fans

Conve

ctio

n

Condense

r

Brine Reflux

natu

ral

Eva

pora

tor

PreheatedSea Water

Heat Source ( 75...85°C)

Hot Sea Water

Distillate

Cold Raw Water

Figure 3.4.5: MEH open-water / closed-air cycle system diagram

- MEH – closed-water/open-air In these systems (Figure 3.4.6), the closed water circulation is in contact with a continuous flow of cold air in the evaporation chamber. The air is heated and loaded with moisture as it passes upwards through the falling hot water in the evaporation chamber.



After passing through a condenser cooled with cold seawater, the partially dehumidified air leaves the unit while the condensate (distillate) is collected. Water is recycled or re-circulated. Incoming cold air provides a cooling source for the circulating water before it re-enters the condenser. Systems with closed saline water cycle ensure a high utilisation of the salt water for fresh water production, although the production of fresh water, as resulting by the experimental installations, has not been proven to be higher than in closed air systems. Referred study of the effect of air flow-rate on the production efficiency, presented a maximum value. The reason for this is related to fact that the increase of the airflow rate, leads to an increase to the heat and mass transfer coefficients in the humidifier and condenser but eventually decreases the operating temperature of the system.

3.4.2.3 Components materials - Characteristics

The most important component for a humidification-dehumidification system is the heat exchanger / humidifier component. In general, the use of plastic materials is suggested, in order to prevent corrosion phenomena. On the other hand, the main problem with plastic material is the questionable resistance in the presented operating temperatures. Polypropylene has been demonstrated as a favourable material. The direct/indirect heat supply configuration has a strong influence to the characteristics of the heat exchanger/humidifier component. More specifically, for a direct system, heated saline water from the collector flows directly through the humidifier, while for an indirect system, the configuration has to ensure the heat transfer to the saline water or the humidifier. Typical heat exchanger geometry consists of tubes placed on perforated plates and supplied by special joints. This makes the eventual dismantling and replacement easier. In the case of direct systems, a well-performing geometry for the humidifier is vertically suspended tissues or fleece, made of polypropylene and over which, the hot seawater is normally distributed. The condenser is a polypropylene bridged double plate heat exchanger through which the cool brine is pumped upwards. The condensate runs down the plates and trickles into a collecting basin. Referred configuration can be implemented to indirect systems also. More specifically, in the case of a heat exchanger presence between the collector and the saline water, the system configuration can be the same. In the case that the hot water of the collector flows through the humidifier, pipes of hot (fresh) water have to be adjusted to the tissues. For the humidifier material, the solution of textile tissues is also a cheap and efficient solution. The geometry of packed tower may also be used for the humidifier, depending on results to be achieved and design conditions. The humidifier material (e.g honeycomb paper, etc) should generally be of such a size and shape as to provide a high contact surface and a low pressure drop, as it has an effect on the thermal efficiency and productivity of the unit. The operating conditions of the condenser refer to lower temperatures (potentially reaching 50°C), thus the use of plastic material is more fea sible. An horizontal tube bundle through which

Heat Input

Eva

pora

tor

Con

dens

er

Warm Air

Distillate

ColdMixed Air

Seawater

Water Recycle

Flat Plate Collector

Air Inflow

Air Outflow

Figure 3.4.6: MEH closed-water/open-air cycle

system diagram



the brine coolant passes in counter-current flow to the fresh water stream surrounding the tube bundle is the most used configuration. As regards the solar part, in terms of materials, and especially for direct heated systems the restrictions due to the contact of saline water with the collector remain, as in the case of active solar stills, directing research to this topic.

3.4.2.4 Techno-economical characterization of solar humidification-dehumidification systems

Concluding the above analysis, solar humidification-dehumidification systems, present a rather improved efficiency when compared to active solar stills. Construction is simple and can be based in locally available materials for most components. Nevertheless, a level of complexity appears as regards mainly the humidifier, but the condenser as well as. For these components, the use of polypropylene might solve corrosion problems arising from the contact of saline water to the humidifier/condenser surface, while a simple and cheap solution for the contact surfaces might be textile pads. The coupling to solar energy and more specifically to flat-plate collectors is a proven solution. Achievement of high-performance in operation conditions of 85-95°C, would act quite positively to the evaporation process, and consequently to the efficiency of the system. In addition, direct heated systems are favourable in terms of energy exploitation (as there is no temperature reduction in the solar part – desalination system heat exchanger), but there is an emerging issue referring to the restrictions to the materials of the collector. Most humidification-dehumidification systems referred are experimental/pilot plants, presenting various configurations. Although a fair amount of simulation studies and experiments have been conducted, further design simulation is required to fully understand the complicated effects of air and water flow rates, the optimum size of individual components or modules of the unit and to generate a comprehensive model for the system – both technical and economical. An arising speculation refers to the burdening of the installation cost with the collectors cost, but there one has to see the potential of these systems to operate in a hybrid manner (fresh water/hot water production) also, as in the case of active solar stills. Concluding, research, so far conducted, has demonstrated the potential of solar humidification-dehumidification technology as an alternative for small capacity desalination plants, up to 10 m³/d.

3.5 COMPARISON OF THE DIFFERENT TECHNOLOGY UNDER INVESTIGATION

In this paragraph, a comparative analysis between different seawater desalination systems driven by low to medium solar thermal collectors will be performed. Concerning the desalination processes, a preliminary screening has been carried out, based upon the information reported in the previous sections. Really the interest will be focused on multi-effect evaporation (ME) only, evaluating different plant configurations. The reasons of the exclusion of other possible systems are summarized as it follows:

� Applications for small users, such like solar stills and humidification-dehumidification, are characterized by a modest performance ratio, thus an enormous collector area per m³/d of installed capacity is required, limiting the interest in these technologies to regions where land availability is not critical;

� Multi-stage flash evaporation (MSF) requires a higher operating temperature than ME to achieve an equal performance ratio, is less flexible with respect to the load variations, has a higher power consumption, and is more sensitive to plant size both in terms of efficiency and cost;

� Cogeneration is valuable for large scale applications and normally requires the use of high temperature concentrating solar technologies, therefore, owing to both these aspects, an investigation of the potential for the combined power and fresh water production by solar energy would go beyond the aims of this work.



3.5.1 Selected solar desalination systems

As previously stated, the basic ME process has been assumed as the reference system. The related values assigned to the parameters needed for the current estimation are listed in Table 3.5.1.

Parameter Value

Number of effects 12

Driving heat source temperature 70 °C

Boiling temperature in the last effect 35 °C

Performance Ratio 9

Table 3.5.1: Values adopted for the assessment of the basic ME (Source: /Des99/).

The most useful solar technology under this driving heat source temperature is constituted by evacuated tubular collectors (ETC). It is to be noted that a reduction in the number of effects would allow to operate the system at a lower temperature: as a rule of thumb 4 effects are equivalent to 10°C. Clearly, to cut down the collec tors working temperature is attractive, since it leads to an increase in their efficiency and makes it possible to use less advanced products, but being the PR of the process roughly in direct proportion to the number of effects, the total system efficiency would be depressed. On the other hand, since the upper limit for the top brine temperature is 63°C, because of scale formation pro blems, an excessive increase in the number of effects would produce an intolerable growth of the heat transfer area. In addition, the use of thermal vapour compression (TVC) has been considered, including in the process a steam jet ejector which entrains the vapour from the last effect and compresses it up to the pressure required by the first one. The operating principle of the ejector is equal to that described in Section 2.6.2, except for the different temperature limits. The increase in PR has been calculated, drawing the ejector entrainment ratio from the diagram reported in Figure 2.6.3, with an inclusive efficiency of nozzle, diffuser and mixing chamber equal to 0.7. Two different operational conditions have been considered, according to the motive steam pressure, which has been fixed to 3 and 10 bar respectively. The first value roughly matches with the lower limit of the range of application of the ejector technology. Regarding the second one, due to the modest growth of the entrainment ratio with the motive steam pressure, the benefits coming from raising this parameter beyond 10 bar are nullified by the increased value of the steam, and so it is not justifiable in practice to exceed this limit, although ME-TVC plants working with pressures up to 45 bar do exist (/Ala05/). The reasons to restrict the range under investigation to the aforesaid value are even more urgent for desalting plants driven by solar thermal collectors, since the overcoming of this limit impose the application of advanced concentrating solar technologies. On the contrary, both evacuated tubular and medium temperature parabolic trough collectors (PTC) can be used to work with a pressure reduced to 3 bar, while PTC only are suitable to raise the pressure up to 10 bar. Finally the potential of boosting the basic process performance via a double-effect absorption heat pump (DEAHP), using H2O/LiBr as working pair has been included in the analysis. The detailed description of the absorption cycle, relevant to the functioning as chiller, is reported in section 2.2.2. Again the only difference of the reverse cycle is related to the increase in the evaporator and condenser temperature, which allows to reach a COP of about 2.2 (/Mil97/). Being the driving temperature still around 180°C, P TC must be used to drive this process. It is to be noted that single-effect absorption heat pump has not been considered, due to the working temperature of at least 140°C required to d eliver vapour at about 65 °C. This relatively high value combined with a COP of 1.6 makes it preferable to utilize, under the same motive steam conditions, an ejector, which allows roughly an equal gain in efficiency mostly preserving plant simplicity.



System Performance boosting device Type of

collector Temperature

[°C] PR

ME-BASIC - ETC 70 9

ME-TVC-3A Ejector working at around 3 bar ETC 130 14

ME-TVC-3B Ejector working at around 3 bar PTC 130 14

ME-TVC-10 Ejector working at around 10 bar PTC 180 15

ME-DEAHP Double-effect absorption heat pump: H2O/LiBr as working pair PTC 180 20

Table 3.5.2: Summary of the investigated systems.

In Table 3.5.2 the five solar desalination systems selected for the analysis are summarized. The values adopted for the two main parameters required for the energy comparison, such like driving heat source temperature and desalting process performance ratio, are also indicated.

3.5.2 Calculation methodology

The main hypotheses and the methodology adopted for the calculation are here briefly discussed:

� The system is supposed to be located at Almeria. This is not a restrictive approach, since, as shown in section 3.1, desalination mostly concerns South Europe countries and Spain in particular.

� The plant capacity has been fixed at 10 m³/d, able to meet the domestic fresh water requirements of more than 50 people. This value has been chosen being the current investigation focused on small users application. On the other hand the scaling up of the obtained results affects within limits the economic performance of the system only, which has been omitted at this stage of the study.

� The analysis has been performed using the PHIBAR f-Chart method, relevant to a general solar heating system, in the same way as the comparison of solar assisted air-conditioning systems presented in section 2.7.

� The load to be fulfilled has been calculated, assuming a thermal energy consumption of 2300 KJ, which is the rounded off value of the heat of vaporization of water at 70°C, for the production of a quantity of distillate, expressed in Kg, equal to the system PR.

� The collector and the general solar heating system parameters adopted in the calculation are again those listed in Table 2.7.3.

� The collector area for each system has been determined by iteration, until the solar fraction is equal to 1 only in the month with the best weather conditions.

3.5.3 Results

In Table 3.5.3, the obtained results for the 5 systems under investigation are reported. The daily load has been calculated as stated before.

System Load [GJ/d]

Solar Fraction

Collector Area [m²]

Specific water production [(m³/year)/m²]

ME-BASIC 2,56 0,85 224 13,9

ME-TVC-3A 1,64 0,83 202 15,0

ME-TVC-3B 1,64 0,77 102 27,6

ME-TVC-10 1,53 0,76 100 27,7

ME-DEAHP 1,15 0,75 75 36,5

Table 3.5.3: Collector area required by the investigated systems and annual specific water production.



The focal point is represented by the collector area: the specific fresh water production on annual basis can be easily drawn from its value and the related solar fraction. It is to be noticed that, even in the month with the worst climate, the solar fraction is nearly 0.6 for all systems. Therefore, in principle the plants can be operated without the contribution of no backup source, by adopting an appropriate load modulation and a relatively short term storage. Finally, in Table 3.5.4 the main elements for the comparison between systems are summarized. On one hand the gain in terms of annual water production per m² of collector area with respect to the basic ME process is considered, on the other the additional devices required and the related extra costs, power consumption and plant complexity are listed. System Water production increase Additional devices

ME-TVC-3A 8% Ejector using motive steam at 3 bar

ME-TVC-3B 99% Parabolic trough collectors Ejector using motive steam at 3 bar

ME-TVC-10 100% Parabolic trough collectors Ejector using motive steam at 10 bar

ME-DEAHP 164% Parabolic trough collectors Double-effect absorption heat pump with H2O/LiBr as working pair

Table 3.5.4: Key elements for comparing systems using thermo-compression to the basic ME process.



4. CONCLUSIONS

A notable increase in power demand in EU-15 area during the hot season has been observed in the last years, as a result of the wide spreading of air-conditioners. According to the most recent forecasts, the present total energy consumptions related to the cooling of buildings should come to around 105 GWh per year. Really, basing upon the present trend, it appears to be plausible that this figure shall be correct to a significantly higher value. Moreover, it is expected a rate of growth of the air-conditioned demand at least similar in the near to medium term, with a consequent further dramatic increase in the summer power requirements. This phenomenon is especially marked in South European countries: in effect over 60% of the total EU-15 energy consumptions related to air-conditioning is concentrated in Spain and Italy and these countries, together with Greece, are also distinguished by the higher values per capita. In this scenario, the exploitation of solar energy, which abounds in these regions during the hot season, seems to be a valuable option to mitigate the consumption of conventional fuels due to the cooling requirements. Nevertheless, in Europe solar driven air-conditioning systems are still in a development stage, as demonstrated by the limited number of plants currently in operation. On the other hand several cooling technologies, which can be in principle powered by low to medium temperature solar thermal collectors, are currently available on the market. Considering also all the possible different configurations for each technology and the various types of usable collectors, a wide range of solar cooling systems have to be evaluated in a complete investigation aimed at the selection of the most promising ones. To limit the number of the systems to be compared, a first screening criterion is the COP value: in fact the modest energy performance of a cooling system, when driven by low to medium solar thermal collectors, leads to a dramatic increase in the required area to meet a given cooling load. This is the main reason of the exclusion from the energy analysis of technologies, such like Organic Rankine Cycles or ejector air-conditioning systems. Similarly Thermo-Chemical Accumulators (TCA), despite these systems represent an interesting technology but still at development stage, have not been included, due to the correspondence of their efficiency and working temperature to those of a single effect water/LiBr absorption chiller, which is on the other hand commercially available. For the purposes of this work, the choice of the most appropriate type of solar collector to be coupled with the remaining cooling technologies and the comparison between the resulting systems has been limited to the global energy performance. Clearly, this evaluation may deliver different indications, according to the type of user (only cooling or both heating and cooling), the ratio between cooling and heating loads, the climate conditions, and the amount of primary energy to be saved. From the related results, presented in section 2.7, it can be drawn that evacuated tubular collectors seem to be the most suitable solar technology to be coupled with low temperature cooling systems (single effect water/LiBr absorption chillers, DEC systems or adsorption chillers). For example, the specific collector area requested to save the 70% of primary energy is about one half, or even less according to the location, with respect to the area relevant to flat-plate collectors, for residential and office buildings. However, it is to be stressed that the estimation has been carried out with reference to high performance ETC collectors, including future improvements in the collector technology also related to the development of dedicated products. On the other hand, for ETC characterized by an efficiency curve closer to the average of the products presently marketed, the specific collector area significantly increases (up to 50%) and the energy benefit compared to flat-plate collectors is strongly reduced. Furthermore, the aforesaid cooling technologies driven by ETC seem to be absolutely the most energy effective solar cooling systems for installations located in hot climate regions. Quite the opposite, advanced solar technologies, such as parabolic trough collectors having a high performance figure, could become competitive in sites characterized by heating loads comparable or greater than cooling loads and an adequate level of direct solar radiation. In this case, cooling machines based on a reversible cycle, such like absorption chillers using



ammonia/water as working pairs, could have real perspectives of development, in particular if a more cost-effective concentrating solar technology will be available on the market. A similar assessment of the equivalent electrical energy consumed for brackish and seawater desalination in EU-15 countries has been carried out for desalination, based upon the data reported in the IDA most recent inventory. The estimated figure is around 104 GWh per year, thus one order of magnitude smaller if compared to the consumptions associated to air-conditioning in the same area. Furthermore the growth of desalination market has been less rapid in latest years and it is expected to keep this trend. On the other hand, as for cooling, also desalination demand is concentrated in the most populated countries of South Europe (over 70% of the total desalination capacity is installed in Spain and Italy). Besides the quite different impact on the EU energy present and future resources, the key distinctive factor of solar desalination is the fact that the application of conventional desalination technologies is on its turn limited to particularly favourable conditions. In other terms, solar desalination must compete not only with fossil fuel based desalination processes, but above all with other traditional water supplying methods. In addition the opportunity of distributed production, which could support the dissemination of solar desalination, appears to be more limited compared to thermally driven cooling systems. Actually, the multi-effect evaporation seems to be the most qualified candidate for the role of reference desalination process to be coupled with solar thermal collectors, due to several favourable characteristics, such like the relatively high performance ratio, the reduced working temperature and the adequate flexibility to the load variation. On the other hand, ME plants are currently erected starting from a capacity around 500 m³/d in order to achieve reasonable economies of scale. This figure corresponds to the fulfilment of the fresh water requirements of nearly 3000 people, if the use is limited to residential civil needs. Clearly smaller capacity systems aimed at specific applications are in principle possible, but the capital cost, which is normally extremely sensitive to the plant scale, would rise dramatically. Furthermore to reduce the required collector area, an adequate performance ratio must be reached via an increase in the number of effects and, as a result, in the plant complexity and dimensions. Therefore a centralized production in a desalting plant of significant capacity is certainly more convenient, considering also that fresh water can be easily stored and transported over a relatively long distance. Quite the opposite, for cooling a localized production is needed and both absorption chillers and desiccant cooling systems for small-scale applications are by now commercially available. This represents a notable discriminating factor, since the further penetration to the market of solar thermal collectors, which is the main target of the NEGST Project, would significantly benefit from the commercialisation of pre-assembled products, simple to install and manage. In conclusion the work relevant to the development of desalination systems driven by solar thermal collectors must concern principally the improvement and cost reduction of small capacity ME plants. Other aspects to be considered to assess the solar desalination potential are the significant power consumption of around 2 KWh per m³ of produced water and the contribution to the initial investment coming from the cost of the required solar collectors area and the related problems of land availability. Concerning the last aspects, a collector area of around 4 m² per person can be assumed for the basic ME, which is comparable to the value pertinent on average to solar cooling systems. Finally, some indications for a pre-selection of the most promising plant configurations can be drawn from the results reported in section 3.5. Besides the basic ME process, the inclusion of a double-effect absorption heat pump, using H2O/LiBr as working pair, appears to be interesting, causing a reduction of about two-thirds of the required collectors area. On the other hand the resulting system is by far more costly and complex to be designed and operated; moreover it must be driven by an advanced concentrating solar technology, such as parabolic trough collectors.



On the contrary thermo-compression via a steam jet ejector does not appear to have a great potential. In fact, when the motive steam pressure is not sufficiently high, the decrease in the required collector area is not significant, unless advanced solar technologies are used. Otherwise, for higher motive pressures, the decrease in the area is very reduced with respect to the DEAHP, which can be equally adopted in this case. In conclusion the results presented in this study constitute a preliminary step towards the assessment of those technologies, whether for cooling or desalination, that are most suitable for the coupling with low to medium solar thermal collectors. For an extensive analysis, other aspects such as level of commercial maturity, economic potential, presence of technological barriers and so on, must be deeply investigated. A definitive evaluation of the most promising solar cooling and desalination systems and their level of interest as a possible incentive to the market of solar thermal collectors will be analysed in detail in the next steps of the NEGST Project Work Package 5.



ANNEX A – Country data on conventional cooling syst ems

The following fact-sheets summarize information about the national market, such as the typical application, size, efficiency, cost, and the market trend. Available national data concern:

� AUSTRIA � GERMANY � GREECE � ITALY � SPAIN � SWEDEN

Country AUSTRIA

Indicator Remarks

Market sharing by technology

Percentage of several cooling technologies (2002) sold by the Austrian AC- and cooling manufacturer: Extra Cooling Technology 2% (Ab- and Adsorption) Components 3% Large Cooling Plants (industry) 5% Others 5% Cooling Equipment (Grocery, market place) 6% Cooling + Freezer chamber 6% Refrigerant Cooling (commerce and industry) 12% Commercial Cooling + Freezer chests 13% Small Cooling/AC-machines (commerce) 23% Large Cooling/AC-machines (commerce) 25% 100%

Client groups - break down: Gastronomy 2.0% Free time business (Hotel ...) 3.8% Private clients 3.9% Automotive 4.3% Mechanical engineering + steel manufacture 5.3% Other 7.5% Chemical industry 7.5% Energy contracting 9.1% Food trade 13.9% Construction industry 14.0% Grocery production 28.9%

100%

The data are based on the manufactured AC- and cooling equipment in Austria. In 2002 the turnover of this market was 1300 Million (incl. business segment of cooling, air-conditioning and heat pump technology). This amount covers the inland and external trade. 38% of the turnover is gained by external trade activities.

The break down for the clients includes only data of inland trade.



Typical size Cooling technologies on the Austrian AC- and cooling market separated by the ranges of cooling capacity: Extra Cooling Technology > 150kW 1% (Ab- and Adsorption) Large cooling plants (industry) > 150kW 3% Other < 10kW 3% Refrigerant Cooling < 150kW 4% (commerce and industry) Refrigerant Cooling > 150kW 5% (commerce and industry) Large Cooling/AC-machines > 150kW 6% (commerce) Large Cooling/AC-machines < 150kW 9% (commerce) Other < 150kW 11% Small Cooling/AC-machines < 10kW 58% (commerce) 100%

Efficiency N.A. Depending on the operating conditions and the technology

Costs (plant cost, running expenses, maintenance, etc.)

N.A.

Market trend: In general architecture should avoid high cooling loads for residential and tertiary building sector in Austria. Consequently the non-active cooling (night ventilation, earth ground coupled heat exchanger, external shading devices etc.) becomes more and more relevant. For such cooling or air-conditioning appliances, where active cooling systems are absolutely needed, experts see basic trends for next 20 years. Trends can be summarised as follows:

� Products should be more environmental friendly � More energy efficient � Reliable and robust

Improvement of existing technologies (mainly compression chiller):

� Reduction of the energy consumption in a range of 30 – 50% � Reduction of refrigerant leakages � Reduction of refrigerant volume per cooling capacity � Improvement of LCCP (Life Cycle Climate Performance)

R&D for technology beside compression chiller:

� District heating providers are interested in thermal driven cooling machines. Research and pilot project are ongoing to provide cold water by ab/adsorption cooling machines driven by district heating network.

There are three running pilot projects for solar cooling systems in Austria. This thermal driven cooling technology is highly wanted by both the solar collector industry and some governmental institution. More demo plants and research activities on solar cooling technology both on the overall systems performance and the component level are expected.

References: � Simader, G. R.: Klimatisierung, Kühlung und Klimaschutz: Technologien, Wirtschaftlichkeit und

CO2-Reduktionspotenziale, Editor - Austrian Energy Agency, Wien, 2005 � Rausch M.: Analyse der österreichischen Kältewirtschaft, Diplom Thesis at FH Pinkafeld

(Studiengang BTM), August 2002



Country GERMANY

Indicator Remarks

Market sharing by technology

Mainly climatisation of public and industrial buildings, the bigger part of systems are supply air systems for heating and cooling (compression cooling).

Market share of RAC (compression cooling) [1]

33,4 % split systems 10,0 % multi split systems

46,3 % single duct systems

10,3 % single packaged systems Absorption chillers [5]

Absorption chiller units for cooling and refrigerating are not of significance in comparison with compression chillers. Installation per year: ca. 70 systems, increasing demand expected (10-30% two-stage systems installed, all other systems are single-stage)

RAC dominant system type in households, but domestic air conditioning less common, rate in non-residential sector is significantly higher

Typical size Small RAC systems < 10 kW cooling capacity

Efficiency Small units: approx. COP = 2-3


Small systems: approx. 1000 Euro/kW

Split systems slightly cheaper (mass production)

Market trend: In recent years growing interest for thermal driven chillers. Up to now only central systems installed. Decentral application in individual rooms not available on the market [3].

Global trend for Germany: huge increase of air conditioned areas, market growth until 2020 10 % per year, in 2020 energy demand for air conditioning will probably be 40 times higher than in 1990 [1].

Construction trend: increase of split systems to more than 50 % at the expense of single duct systems.

References: 1. Energy Efficiency of Room Air-Conditioners (EERAC) – Study for the Directorate-General for

Energy (DGXVII) of the Commission of the EU – Final Report – May 1999 2. Energy Efficiency and Certification of Central Air Conditioners (EECCAC) – Study for the D.G.

Transportation-Energy (DGTREN) of the Commission of the EU – Final Report – Volumes I-II-III – April 2003

3. BINE themeninfo I/04 – Klimatisieren mit Sonne und Wärme – 2004 4. FGK, Fachinstitut Gebäude-Klima e.V., Bietigheim, Germany – 2005 – personal communication 5. Wo steht die Kältetechnik in Deutschland und weltwe it? – Prof. Dr.-Ing. Johannes Reichelt,

Fachhochschule Karlsruhe - 2000

Comment: Barely no market data for German traditional cooling market available.

[4]: Approx. 1,5 million AC systems installed; yearly growth around 150.000 systems; approx. 40 million m² conditioned area; required cooling power per m²: 100-120 W/m².

Production volume [5]: Approx. 2200 companies dealing with cooling / conditioning systems, production volume: approx. 2,5 billion Euro (with inclusion of air conditioning for cars and small cooling and refrigerating units: approx. 8,5 billion Euro); 40 % is produced for export.



Country GREECE

Indicator Remarks

Market share by technology

Vapour compression chillers (split units) (99%)

Thermal driven chillers for industrial applications (<1%)

For thermal driven chillers, this portion is expected to increase significantly during the next years, due to the increasing penetration of natural gas

Typical size 3.5 KW for split unit -

Efficiency COP around 3 Depending on the operating conditions


For a typical split unit of 3.5 KW the cost lies within the range of 250-500 €.

For a larger system of 5.3 KW, the respective range is 600-1000 €.

Running expenses refer to the electricity consumption: 0.09 €/kWh. Maintenance costs are estimated about 50 €/year.

In Greece there is no production of such units, except assembling of central air-conditioning units

Market trend: From a technological point of view, there are no significant changes expected, apart from the increasing penetration of thermal driven systems (due to natural gas penetration), and the domination of inverter type systems, as regards vapour compression systems.

As regards the market, what has to be pointed out is a gradual but constant increase of mostly Chinese products penetration, characterized by lower prices.

References: � Air-conditioning, Sectorial Study, ICAP SA, 1998

� Recent articles mainly from specialized magazines in the field of energy



Country ITALY

Indicator Remarks


Vapour compression chillers represent 99% of the market. Among these, RAC units correspond to about 77%.

In particular, the market share of RAC:

� split and multi-split systems 92% � single duct and single packaged systems 8%

The remaining market share (∼1%) is held by thermal driven chillers for large-scale applications.

Italy is the largest RAC consumer and the largest manufacturer in the EU.

Typical size Up to 7 KW for split units.

From 17 to 350 KW for large compression chillers (water or air cooling).

Up to 2.5 MW for thermally driven chillers (absorption water/LiBr for air conditioning).

Efficiency COP lies within the range 2.5 ÷ 5 from small to large systems.

Depending on the type of system and the operating conditions.


For a typical split unit the average cost is around 450 €. For multi-split units is around 1500 €.

Market trend:

� In the last years a notable market growth has been observed for small compression chillers (up to 40% for room air conditioners and in particular for split and multi-split units)

� A small reduction of sales (around 5-6%) has been observed instead for large scale systems

� Lately, there is a significant increase in the interest towards thermally driven chillers

References: � CO.AER - Associazione dei costruttori di apparecchiature e impianti aeraulici (2004): Indagine

statistica sul mercato dei componenti ed impianti di condizionamento dell’aria. (http://www.coaer.it)



Country SPAIN

Indicator Remarks


Vapour compression chillers represent 88% of the market. Among these, 64.3% corresponds to domestic systems and 23.7% corresponds to commercial-residential installations. Thermal driven chillers for industrial applications (12%)

During 2003 the market of air conditioning increased 20.6% due to the increasing of domestic systems because of high temperatures in the last years.

Typical size < 7 KW for domestic use (for split units)

7-70 KW for commercial-residential use

> 70 KW for industrial applications

Efficiency COP fluctuates within the range 2.5 ÷ 5

Depending on the type of system and the operating conditions.


For a typical split unit the average cost is within the range 400-800 Euros. The inverter type systems are 180 Euros more expensive.

Market trend: � The market of inverter type systems is increasing in spite of these systems are more expensive, but

the energy consumption is lesser. � Due to high water consumption and to “legionella bacteria” in the cooling towers, water condensed

chillers is below 30% of the total, while that air condensed chillers increases its market till 70-80%. � Regarding the market, what has to be pointed out is 61.5% products come from Oriental Countries.

References: � AFEC (Asociación de Fabricantes de Equipos de Climatización) (www.afec.es) � CIATESA Aire Acondicionado. (www.ciatesa.com)



Country SWEDEN

Indicator Remarks


Compressor driven chillers in single buildings, cooling being supplied either to the ventilation air or to ceiling cooling panels and also room fan coils.

Direct and indirect evaporative cooling. District cooling, where the cooling is supplied from:

1. Free cooling, often from lakes where the temperature at the bottom is 4°C year round;

2. Absorption chillers running directly from the power or district heating plant;

3. Compressor chillers; Absorption chillers running from district heating, although this market segment is still small

Underground storage using aquifers or boreholes, often in combination with heat pump system for winter heating.

Concerning underground storage, these are used both for individual buildings or in combination with district cooling. Generally the cooling is supplied as free cooling. In 2002 there were 40 large scale aquifer stores and 200 stores with more than 10 boreholes. Not all of the borehole stores are used for cooling however.

Typical size N.A.

Efficiency N.A.


N.A. Statistics are in the process of being compiled

Market trend: The trends that can be found, but not quantified are: � Increased use of air-air heat pumps in domestic and small office locations. Sales have increased

significantly over the last 5 years but it is impossible to judge how much these are used for cooling. � Increased installation of comfort cooling in commercial and public buildings due to perceived warmer

summers and increased internal loads. � District cooling has increased from 180 to 600 GWhcooling in the period 1998-2002. In 2002 there were 800

customers with an installed capacity of 500 MW making it, together with France, the largest market in Europe. 40% of this capacity is for year round base load. The estimated market for district cooling is 2 TWhcooling by 2010.

� Increased interest in providing local cooling using thermally driven cooling processes, with heat supplied from district heating.

� There is one system using snow storage from winter to summer, and several feasibility studies have been recently made for similar new projects. These projects are for relatively large cooling loads.

References: � Frohm, H. (2004): District Cooling in Europe - Potential, Technology, Benefits, Success factors. In

Presented at IEA Seminar Cooling buildings in Warming Climate, June 2004. � STEM (2004): Energiläget 2003, Energimyndigheten, Eskilstuna, Sweden. www.stem.se. � Andersson O., G. Hellström and B. Nordel (2003): Heating and Cooling with UTES in Sweden - Current

Situation and Potential Market Development. In Proceedings FUTURESTOCK 2003, Warsaw Poland. � Feldhusen H. & Martì i Ruiz, F. (2001): District Cooling – Present Market Assessment. Stockholm: Kungl

Tekniska Högskolan, Institutionen för Energiteknik, avdelningen för Tillämpad termodynamik och kylteknik. Examensarbete nr. E:277/2001.

� Westin P. (2002): Fjärrkylans framtid och potential. Fjärrvärmeföreningens Temadagar Fjärrkyla.



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