Applicability of the Solar Thermal Ultra High Vacuum Collector for Heating,Cooling and Power Generation

Applicability of the solar thermal Ultra High Vacuum collector for heating, cooling and power generation

Konstantinos Chatzichristos May, 2014

A catalogue record is available from the Eindhoven University of Technology Library

ISBN: 978-90-444-1280-2 (Eindverslagen Stan Ackermans Instituut ; 2014/022)

EINDHOVEN UNIVERSITY OF TECHNOLOGY

Stan Ackermans Institute

SMART ENERGY BUILDINGS & CITIES

APPLICABILITY OF THE SOLAR THERMAL ULTRA HIGH VACUUM COLLECTOR FOR HEATING, COOLING AND POWER GENERATION

Konstantinos Chatzichristos

A thesis submitted in partial fulfillment of the requirements for the degree of Professional Doctorate of Engineering

Eindhoven, the Netherlands

May, 2014

This thesis has been established in collaboration with

ABSTRACT

In the present work the applicability of the solar thermal Ultra High Vcacuum collector for heating, cooling and power generation is examined. In Chapter 1 the research question and the framework of the current project is presented. In Chapter 2 the unique features of the solar thermal Ultra High Vacuum (UHV) collector are presented with the aim of gaining deep understanding of the involved technology and the extraordinary performance of the collector. The solar thermal Ultra High Vacuum (UHV) collector is an evacuated flat plate collector that presents performance at high temperatures without solar tracking. The use of Ultra High Vacuum (UHV) allows the collector to “trap” the absorbed solar irradiation, under harsh weather conditions and in locations with high percentage of diffuse light. The result is that the solar thermal UHV collector is able to deliver heat in a wide range applications and covering the energy demand for heating, cooling and power generation. Chapter 3 deals with the design considerations of solar thermal systems that the solar thermal UHV collector is integrated to. In addition, guidelines for designing a generic solar system with the solar collectors are presented, while emphasis is given to the solar field of the installation. After analyzing the solar thermal UHV collector as a component and its integration to a general solar system, the gained knowledge is applied in specific studies based on a reference building in Chapter 4 and Chapter 5. In more detail, solar heating and solar cooling according to the existing building infrastructure is examined in Chapter 4, with target to investigate the applicability of the solar collector in the mentioned applications. In Chapter 5 the solar thermal UHV collector is utilized as a heat source for power generation via a small scale Rankine cycle. Different system configurations are studied and the comparison of their productivity in different locations takes place. Finally, taking into account the performed investigations useful conclusions are presented in Chapter 6 for the studied heating, cooling and power generation applications. Furthermore, recommendations that could improve the studied solar systems are presented in the same chapter.

Table of Contents

Table of Figures ...... 5 Table of Tables ...... 7 Chapter 1. Problem definition ...... 9 1.1 Introduction ...... 9 1.2 SolCalor ...... 9 1.3 Applicability of the solar thermal Ultra High Vacuum collector for heating, cooling and power generation ...... 10 1.4 Outline of the report ...... 13 Chapter 2. The Solar thermal UHV collector ...... 15 2.1 The solar thermal Ultra High Vacuum collector ...... 15 2.2 Characteristics of the collector ...... 16 2.2.1 Ultra High Vacuum ...... 17 2.2.2 Getter pump ...... 18 2.2.3 Metal -to- glass welding ...... 18 2.2.4 Selective surface coating ...... 18 2.2.5 Available configurations ...... 19 2.2.6 Performance of the solar thermal UHV collector ...... 19 2.2.7 High stagnation temperature ...... 21 2.2.8 Operation temperature and pressure ...... 22 2.3 Comparison to other solar thermal collectors ...... 22 2.4 Applications ...... 23 Chapter 3. Solar thermal systems with solar thermal UHV collectors ...... 25 3.1 Introduction ...... 25 3.2 Design considerations and system requirements ...... 26 3.2.1 Concept of solar thermal systems...... 26 3.2.2 Consumption - Thermal load ...... 26 3.2.3 Generation - Solar field ...... 27 3.2.4 Collection of the necessary information ...... 29 3.3 Model for the solar thermal system with the UHV collectors ...... 29 3.3.1 Design of the solar field ...... 29

3.3.2 Model of the solar system ...... 33 3.3.3 Validation of the applied model ...... 36 Chapter 4. Solar heating and cooling: Integration of the solar thermal UHV collector to a building .. 39 4.1 Case description ...... 39 4.1.1 The institute ...... 39 4.1.2 Cooling system of the institute ...... 40 4.1.3 Heating system of the new building ...... 40 4.1.4 Proposal for using the UHV collectors ...... 42 4.2 Solar heating ...... 42 4.2.1 Heating profile of the new building ...... 42 4.2.2 Solar thermal system ...... 45 4.2.3 Results ...... 47 4.3 Solar cooling ...... 51 4.3.1 Heating demand profile of the applied chillers ...... 51 4.3.2 Solar cooling with absorption chillers ...... 53 4.3.3 Results ...... 55 Chapter 5. Small scale steam cycle powered by solar thermal UHV collectors ...... 59 5.1 Power generation with a small scale steam cycle ...... 59 5.1.1 Description of a Rankine cycle ...... 59 5.1.2 Heating demand of a Rankine cyle ...... 60 5.1.3 Variations of the heating process in a Rankine cycle ...... 61 5.1.4 Heat demand of a Rankine cycle with an unfired boiler ...... 62 5.2 Powering a Rankine cycle with solar thermal UHV collectors ...... 63 5.2.1 Connection of the solar system to the natural gas burner ...... 63 5.2.2 Direct coupling of the solar field ...... 64 5.2.3 Connection through a buffer tank ...... 68 5.3 Comparison to the solar systems’ productivity in Spain ...... 73 5.4 Combined heat and power ...... 76 Chapter 6. Conclusions and Recommendations ...... 79 6.1 Solar heating and cooling ...... 79 6.1.1 Conclusions ...... 79 6.1.2 Recommendations ...... 80

6.2 Power generation ...... 81 6.2.1 Conclusions ...... 81 6.2.2 Recommendations ...... 82 References ...... 85 APPENDIX A: Project parameters questionnaire ...... 87 APPENDIX B: Solar and geometric equations ...... 89 APEENDIX C: Solar Irradiation data ...... 91 APPENDIX D: Solar heating and solar cooling results tables ...... 95 APPENDIX E: Power generation results tables ...... 99

Table of Figures FIGURE 1.1 THE HEATING SYSTEM OF THE NEW BUILDING ...... 11 FIGURE 2.1 THE ULTRA HIGH VACUUM SOLAR THERMAL COLLECTOR ...... 15 FIGURE 2.2 STAGNATION TEMPERATURE UNDER HARSH WEATHER CONDITIONS ...... 16 FIGURE 2.3 CROSS SECTION OF THE SOLAR THERMAL UHV COLLECTOR (REPRESENTATION) ...... 16 FIGURE 2.4 RELATIONSHIP BETWEEN THE QUALITY OF VACUUM AND THE MEAN FREE PATH LENGTH (SOURCE: HTTP://WWW.UIC- GMBH.DE/EN/BASICS/VACUUM-TECHNOLOGY.HTML) ...... 17

FIGURE 2.5 COLLECTOR CONFIGURATIONS WITH A CONCENTRATION FACTOR (LEFT TO RIGHT) 1 (C0), 2 (C1) AND 2.68 (C2) ...... 19

FIGURE 2.6 OPTICAL EFFICIENCY OF THE SOLAR THERMAL UHV COLLECTOR (C0, C1 AND C2) IN FUNCTION OF THE TEMPERATURE DIFFERENCE ...... 20 2 FIGURE 2.7 PERFORMANCE CURVES OF THE C1 CONFIGURATION IN REGARDS TO THE ABSORBER, APERTURE, AND GROSS AREA [M ] OF THE SOLAR THERMAL UHV COLLECTOR ...... 21

FIGURE 2.8 VARIATION OF THE PERFORMANCE OF THE SOLAR THERMAL UHV COLLECTOR (C1) AT DIFFERENT IRRADIANCE LEVELS (100- 2 1,000 W/M ) ...... 21 2 2 FIGURE 2.9 VARIATION OF THE OPTICAL EFFICIENCY AT 300 W/M (SOLID LINE CURVES) AND 1,000 W/M (DOT STYLE CURVES) ...... 22 FIGURE 2.10 PART OF THE INSTALLATION OF THE SOLAR THERMAL UHV COLLECTORS AT THE GENEVA AIRPORT ...... 24 FIGURE 3.1 TYPICAL SCHEME OF A HYDRAULIC INSTALLATION WITH UHV SOLAR THERMAL COLLECTORS ...... 25 FIGURE 3.2 THE CONFIGURATION OF A SOLAR FIELD IS DETERMINED BY THE DESIRED MASS FLOW (GREEN ARROW) AND THE TEMPERATURE INCREASE OF THE HEAT CARRIER (RED ARROW) ...... 27 FIGURE 3.3 DESIGN OF THE SOLAR FIELD, BASED ON THE TARGETED THERMAL LOAD ...... 30 FIGURE 3.4 ILLUSTRATION OF THE MODEL USED FOR THE SOLAR THERMAL SYSTEM WITH THE UHV COLLECTORS WITH AN INTEGRATED BUFFER TANK ...... 34 FIGURE 3.5 THE MODEL USED FOR THE BUFFER TANK OF THE SOLAR SYSTEM ...... 35 FIGURE 4.1 THE BUILDINGS OF THE INSTITUTE ...... 39 FIGURE 4.2 HEATING INSTALLATION OF THE BUILDING ...... 41 2 FIGURE 4.3 AVAILABLE ROOF AREA FOR THE SOLAR FIELDS FOR SPACE HEATING OF THE NEW BUILDING (IN BLUE, 240 M ) AND FOR SOLAR 2 COOLING OF THE INSTITUTE (IN ORANGE, 1,400 M ) ...... 42 FIGURE 4.4 DISTRIBUTION OF THE DEMAND FOR SPACE HEATING OF THE NEW BUILDING THROUGHOUT THE YEAR ...... 44 FIGURE 4.5 THE LOAD DURATION CURVE OF THE HEATING DEMAND ...... 44 FIGURE 4.6 DESIGN OF THE SOLAR SYSTEM ACCORDING TO THE MINIMAL MONTHLY ENERGY DEMAND ...... 46 FIGURE 4.7 DESIGN OF THE SOLAR SYSTEM ACCORDING TO MAXIMAL ROOF SPACE FOR THE SOLAR FIELD ...... 46 FIGURE 4.8 THE PRODUCTIVITY OF THE SOLAR FIELD CONSISTING OF 4 AND 24 COLLECTORS COMPARED TO THE ENERGY DEMAND FOR SPACE HEATING ...... 47 FIGURE 4.9 ENERGY DELIVERED BY THE INVESTIGATED SOLAR THERMAL SYSTEMS ...... 49 FIGURE 4.10 TEMPERATURE VARIATION OF THE BUFFER TANK FOR THE SYSTEM CONFIGURATIONS A, B AND C ...... 51 FIGURE 4.11 THE HEATING DEMAND OF AN ABSORPTION CHILLER, SIZED ACCORDING TO THE APPLIED SOLAR FIELD AND THE FREE COOLING TECHNIQUE ...... 53 FIGURE 4.12 SOLAR COOLING WITH A 1-STEP ABSORPTION CHILLER...... 54 FIGURE 4.13 SOLAR COOLING WITH A 2-STEP ABSORPTION CHILLER...... 55 FIGURE 4.14 SOLAR ENERGY DELIVERED TO THE 1-STEP ABSORPTION CHILLER THROUGHOUT THE YEAR ...... 55 FIGURE 4.15 OPERATION HOURS OF THE SOLAR THERMAL INSTALLATION CONNECTED TO A 1-STEP ABSORPTION CHILLER ...... 56 FIGURE 4.16 SOLAR ENERGY DELIVERED TO THE 2-STEP ABSORPTION CHILLER THROUGHOUT THE YEAR ...... 57 FIGURE 4.17 OPERATION HOURS OF THE SOLAR THERMAL INSTALLATION CONNECTED TO A 2-STEP ABSORPTION CHILLER ...... 58 FIGURE 5.1 A RANKINE CYCLE ...... 60 FIGURE 5.2 POSSIBLE WAYS FOR HEAT ADDITION IN A STEAM CYCLE ...... 61

FIGURE 5.3 HEAT ADDITION TO THE CYCLE THROUGH A HEAT EXCHANGER (UNFIRED BOILER) ...... 62 FIGURE 5.4 DIRECT COUPLING THE SOLAR FIELD WITH THE AUXILIARY BURNER OF THE CYCLE THROUGH A 4-WAY VALVE ...... 64 FIGURE 5.5 THE SOLAR POTENTIAL IN HOURS OF IRRADIANCE THROUGHOUT A CALENDAR YEAR IN ERLANGEN, GERMANY ...... 65 FIGURE 5.6 GENERATED ENERGY BY THE SOLAR FIELD COUPLED TO THE AUXILIARY BURNER OF THE HEATING PROCESS ...... 67 FIGURE 5.7 SOLAR SYSTEM WITH AN INTEGRATED BUFFER TANK CONNECTED IN SERIES TO THE BURNER OF THE CYCLE’S HEATING PROCESS ...... 69 FIGURE 5.8 DELIVERED ENERGY BY THE SOLAR SYSTEM WITH AN INTEGRATED BUFFER TANK, CONNECTED IN SERIES TO THE AUXILIARY BURNER OF THE HEATING PROCESS...... 70 FIGURE 5.9 OPERATION HOURS OF THE SOLAR THERMAL INSTALLATION FOR THE DESIGNED THERMAL LOAD ...... 72 FIGURE 5.10 SOLAR POTENTIAL IN ERLANGEN (GERMANY) AND VALENCIA (SPAIN), RESPECTIVELY ...... 74 FIGURE 5.11 COMPARISON OF THE DIRECT COUPLED SOLAR FIELD (7 COLLECTORS) IN ERLANGEN AND VALENCIA ...... 75 FIGURE 5.12 COMPARISON OF THE SOLAR SYSTEMS WITH AN INTEGRATED BUFFER TANK (24 COLLECTORS) IN ERLANGEN AND VALENCIA 76 FIGURE 5.13 ENERGY DEMAND FOR SPACE HEATING WITH A HIGHER AND LOWER CAPACITY OF 11.7 KW ...... 77 FIGURE 5.14 HEAT REJECTED BY THE CONDENSER OF THE CYCLE THROUGHOUT THE YEAR ...... 78 FIGURE 6.1 SOLAR FIELD THAT BYPASSES THE BUFFER TANK DURING WINTER MONTHS ...... 82 FIGURE 6.2 HEAT ADDITION TO THE WORKING FLUID (WATER) OF THE STEAM CYCLE ...... 83 FIGURE 6.3 SOLAR ENERGY EJECTED TO THE EVAPORATOR OF THE HEATING PROCESS OF THE RANKINE CYCLE ...... 84 FIGURE 6.4 EXPLOITATION OF A FLASH VESSEL FOR STEAM PRODUCTION OF THE RANKINE CYCLE ...... 84

Table of Tables

TABLE 2.1 THE OPTICAL PARAMETERS OF C0, C1 AND C2 ...... 20 TABLE 4.1 THE HEATING LOADS OF THE BUILDING ...... 41 TABLE 4.2 SOLAR FACTOR OF THE COMPARED SOLAR FIELDS THROUGHOUT THE YEAR ...... 48 TABLE 4.3 SOLAR FRACTIONS OF THE COMPARED SOLAR SYSTEMS FOR EACH MONTH ...... 50 TABLE 4.4 MAIN PARAMETERS OF THE SYSTEM CONFIGURATIONS A, B AND C ...... 50 TABLE 4.5 SOLAR FACTOR OF THE SOLAR COOLING SYSTEM (1 STEP ABSORPTION CHILLER) ...... 56 TABLE 4.6 SOLAR FACTOR OF THE SOLAR COOLING SYSTEM (2 STEP ABSORPTION CHILLER) ...... 57

TABLE 5.1 THE POWER (QSF,OUT) AND THE TEMPERATURE (TSF,OUT) OUTPUT OF THE SOLAR FIELD AT THE AVERAGE VALUES OF THE ELEVEN IRRADIANCE RANGES ...... 66 TABLE 5.2 SOLAR FRACTION OF THE SYSTEM CONSISTING OF 7 UHV COLLECTORS ...... 68 3 TABLE 5.3 SOLAR FRACTIONS OF THE SYSTEM CONSISTING OF 24 UHV COLLECTORS AND A 15 M BUFFER TANK CHANGE AS WELL ...... 71 TABLE 5.4 INFLUENCE OF THE SOLAR FRACTIONS OF THE SOLAR SYSTEM WITH THE INTEGRATED BUFFER TANK ...... 73 TABLE 5.5 SOLAR FRACTIONS OF THE DIRECT COUPLED SOLAR FIELD (7 COLLECTORS) IN ERLANGEN AND VALENCIA ...... 75 TABLE 5.6 SOLAR FRACTIONS OF THE SOLAR SYSTEMS WITH AN INTEGRATED BUFFER TANK (24 COLLECTORS) IN ERLANGEN AND VALENCIA ...... 76

Chapter 1. Problem definition

1.1 Introduction Buildings are responsible for 40% of the annual energy consumption in Europe. Due to this fact and in combination with the climate and energy targets of 2020, the European Commission adopted a further directive regarding the reduction of the energy consumption in buildings. According to the Energy Performance of Buildings Directive (2010), after 2020 only nearly-zero energy buildings will be constructed in the country-members of EU. In order to achieve this target engineers and building designers have to focus on the reduction of energy losses and the integration of renewable energy systems in buildings.

Towards the implementation of the energy targets of 2020, ENIAC JU “Direct Components and Grids” (DCC+G) is a European project that aims to develop an energy-efficient, integrated DC energy distribution system, based on innovative semiconductor power technologies for commercial buildings. The DCC+G project consortium consists of 14 European companies and research institutes from 5 countries and is being supported by ENIAC JU and by national funding authorities. The DCC+G project mainly focuses on the reduction of energy consumption, on the energy distribution and management (smart grid) and on the sustainable and efficient energy generation. Integration of these modules will be demonstrated in office and retail test-bed locations. Among others, the energy demand of the test-beds will be covered by solar thermal installations.

1.2 SolCalor SolCalor is a business-to-business small medium enterprise (SME) that is active in the solar thermal energy sector. SolCalor aims to play a significant role in the economic use of solar energy by substantially increasing the implementation of the solar thermal Ultra High Vacuum collector, a unique solar thermal collector manufactured by SRB Energy. SolCalor is the exclusive distributor of the solar thermal UHV collector in the area of BeNeLux and posseses the technical knowhow for the design of solar thermal installations with the UHV solar thermal collectors. The added value of the company to the society is the introduction of the innovative solar thermal UHV collector to the market of BeNeLux, and the ability of the firm to provide solutions-oriented-services using solar thermal technology. The core activity of SolCalor is the creation of a sales network and marketing of the UHV collectors in the BeNeLux region. The firm is business-to-business oriented; therefore the focus is on conducting strategic partnerships with stakeholders in the renewable energy sector.

SRB Energy is the first company in the world to have developed, industrialized and distributed a solar thermal collector featuring Ultra High Vacuum technology. SRB Energy is the outcome of research activities held at the European Organization of Nuclear Research (CERN), where its research and development (R&D) laboratories are still located. In the laboratories, research activities are carried out on technologies and materials adopted for the UHV collector and its applications, in order to improve its design and reliability. The R&D activities focus on the continuous improvement of the solar thermal UHV collector and its manufacturing process.

1.3 Applicability of the solar thermal Ultra High Vacuum collector for heating, cooling and power generation SolCalor as one of the involved partners of the DCC+G project is going to integrate the solar thermal Ultra High Vacuum (UHV) collector in a pre-selected demonstration office building, in order to explore the benefits and the limitations of the applicable systems with the solar thermal UHV collector. This is going to be realized by theoretical studies for various solar thermal systems covering heating, cooling and power generation of the reference office building and in line to the existing infrastructure.

The office building is part of an institute located in Erlangen, South Germany. Due to the functionality of the institute (laboratories, clean rooms and offices) there is a high electricity consumption through the whole year. In 2011 the whole institute consumed around 4 GWh of electricity, from which more than 25% was used to cover the cooling demand of the institute. In 2011 the consumption of thermal energy supplied by the local district heating network was 2 GWh. The institute consists of 4 buildings: a so- called old building, 2 clean rooms with a total surface area of more than 1,500 m2 (a small clean room of 500 m2 and a big clean room of 1,000 m2) and a new building with a total surface area of 1,600 m2.

The cooling system of the institute consists of a centralized installation that distributes cooling all around the institute. Cooling is generated by three Trane compressor chillers of two types, one RTHA

450 (Pcool,max=814 kW) that is used for air conditioning in summer and two RTHA 215 (Pcool,max=469 kW) that are used in turns over the whole year for industrial cooling of the clean rooms. The RTHA 450 chiller feeds a low temperature circuit (Tfeed=6˚C , Tret=12˚C), while the two RTHA 250 chillers provide chilled water on the circuit that operates in temperatures Tfeed=12˚C and Tret=17˚C.

The new building of the campus (reference office building) was built in 2012 to support the activities of the institute with an additional 1,600 m2 of useful surface area and its heating demand is covered by the local district heating network. The connection of the heating installation to the district heating system is taking place through a 100 kW heat exchanger. The installation delivers thermal energy to the four thermal loads of the building air conditioning, hot tap water, radiators and concrete core activation. The central heating circuit is separated in a low and a high temperature branch. The first one feeds with energy to the radiators and the core activation component, while the high temperature branch supplies energy to the air conditioning system and the hot water loads (Figure 1.1).

Figure 1.1 The heating system of the new building

The management of the campus faces two challenges related to the energy consumption of the institute:

1. Due to the functionality of the buildings of the institute (mainly consisting of clean rooms), there is an enormous demand for cooling that has to be supplied in a constant base (24/7). Possible integration of solar cooling at the roof of the buildings could decrease the peaks of electricity, or even replace a part of the capacity of the compressor chillers by using absorption chillers.

2. Furthermore, an extension of the campus has been scheduled with the construction of an extra office building identical to the so called new building. The main goal is to become as much as possible energy independent/efficient, since the local district heating system network operates already on its full capacity and it is not possible to provide the newly scheduled building with thermal energy. The solar thermal collectors could provide the building with heat, in order to cover its heating demand.

Objective The objective of the present work is to investigate the applicability of the solar thermal UHV collector to the selected test-bed in Erlangen for heating and cooling applications and for power generation. Moreover, the generated report will be used as a “reference manual” with concrete technical information on how to apply solar thermal technology with the solar thermal Ultra High Vacuum collector.

Project cases The project cases that will be covered in the present report are the following:

1. Solar Heating For the design of the solar system oriented for solar heating, there are two main constraints that will be taken into account. The first is the fact that the demand for space heating in summer months is low

11 compared to winter months, thus over-production of energy will occur when heat is not required for space heating. The second one is related to the available roof space of the building for the installation of the solar thermal UHV collector. Under these constraints the cases that will be examined are as follows:

a. The design of a solar thermal system with a limited generation based on the heating demand during summer months.

b. The design of a solar thermal system that exploits all the available roof space of the building. In this case, the total available surface is utilized in order to identify the maximal amount of energy that can be extracted for the given building. By increasing the size of the solar field, a higher heating demand can be covered in winter, although overproduction of heat will occur in the summer season. The latter scenario will not be an issue, if the excess generated energy of the solar system can be consumed either by a heating or a cooling load.

In practice, the first case can be considered as a pilot installation of an upcoming larger solar system (second case) that may be installed later in the future. Through the pilot installation, the developer and the end user may identify the practical challenges and the improvements required for the optimum utilization of the UHV collectors to the building.

2. Solar Cooling For the investigation of solar cooling with the UHV collectors heat driven chillers (absorption machines) will be used. During the last decade solar cooling with absorption machines is getting more and more popular. An absorption chiller is thermodynamically similar to a conventional vapor compression chiller, with the difference that the refrigeration cycle instead of using a compressor, utilizes a complex group of devices, in order the working fluid of the cycle to upgrade from a low pressure to a high pressure state. The driving force of an absorption chiller is thermal energy, thus it is a heat driven machine, rather than an electrically-driven compressor chiller. Depending on the type of the absorption chiller used, the coefficient of performance varies in a range of 0.6 to 1.2, while the conventional compressor chillers operate with an average COP=3.0. The first ones are less efficient compared to the compressor chillers, however absorption machines rely upon the energy input that is relatively cheap, like solar energy or waste heat.

Due to the limitation of the existing solar collectors (flat plate and evacuated tubes on low vacuum) to operate efficiently for temperatures higher than 90˚C, the majority of the applications is constrained to one step absorption chillers, operating on a temperature level of 70-90˚C, with a COP 0.6-0.7. On the contrary, the ability of the solar thermal UHV collector to operate on relatively medium to high operation temperatures compared to the existing solar collectors, allows the use of double step absorption chillers (150-180˚C) with a COP 1.1-1.2.

3. Power generation (steam turbine) The solar systems that are going to be investigated for power generation will be connected to a small scale power cycle. In more detail, power generation will be realized by integrating the solar thermal UHV collector to a steam cycle (Rankine cycle) equipped with a turbine having a power output of 1.2 kWe. Suitable layouts of the solar system will be investigated and system simulations will be performed, in order to explore the potentials of generating electricity via the solar thermal UHV collector. The results will be extracted under the local solar potential (Erlangen, South Germany), and will be compared to the performance of the same solar systems if they would be installed in a more mild climate (Valencia, Spain).

Space availability One of the most important parameters for the realization of a solar thermal project is the availability of surface in order the solar field to be installed. The available roof area that can be used for the collectors installation is limited to two surface areas consisting of 240 m2 (roof of the office building) and 1,400 m2 (roof of the old building). Considering that each solar thermal collector needs around 5 m2 of gross surface area for installation, plus approximately another 5 m2 to overcome the over shading among the collectors, the big solar field can facilitate around 140 collectors, while the small solar field around 24 collectors.

1.4 Outline of the report Firstly an introduction chapter (Chapter 2) regarding the applied technology in the solar thermal Ultra High Vacuum collector is given, focusing on the exploitation of the Ultra High Vacuum (UHV), the getter pump technology, the metal -to- glass welding and the applied selective surface coating. Later on, the available configurations of the product and their performance are presented. Moreover, emphasis is given on the achievable high stagnation temperature, the operation temperature and pressure, while comparison to other solar thermal collectors is taking place, as well. Chapter 2 closes with some reference projects that the UHV collector has been already integrated to. In Chapter 3 solar thermal systems with the solar thermal UHV collector are analyzed by setting the design considerations and the system requirements. Starting point is the concept of solar thermal systems and more detailed discussion regarding the energy consumption (thermal load), energy generation (solar field). In addition a model applied for the solar thermal system with the UHV collectors is presented with focus on the design of the solar field and the model of the applied solar system. Solar heating and cooling integration to a building by the solar thermal UHV collector is investigated in Chapter 4. The applicable limitations and constraints are presented (case description) on the level of the institute, its cooling system and the heating system of the new building (reference building). The applied solar systems and the results of each of the two aspects (heating and cooling) are treated separately with target the deep understanding of the existing challenges. In Chapter 5 a small scale steam cycle powered by solar thermal UHV collectors is studied. A short introduction with the description of a Rankine cycle and its heating demand sets the framework of the investigation, while the available variations of the heating process tailor the applied solar systems. The studied solar system layouts include the connection of the solar system to the natural gas burner through direct coupling of the solar field, and via a connection through a buffer tank. The applied systems are compared to the productivity of the same systems under a more favorable solar

13 potential location and finally the aspect of combined heat and power in the new office is studied. The report closes with a chapter dedicated to conclusions and recommendations (Chapter 6).

Chapter 2. The Solar thermal UHV collector

2.1 The solar thermal Ultra High Vacuum collector The solar thermal Ultra High Vacuum (UHV) collector is an evacuated flat plate collector with high performance that uses innovative technology transferred from CERN (Figure 2.1). The use of the Ultra High Vacuum (UHV), the integrated non-evaporable getter pump and the absorber selective surface coating are some of the most important technologies covered by a CERN patent in 2003 and published in the framework of CERN’s Technology Transfer policy.

Details about the design and the production of the solar thermal collector are given in [1]. One of the greatest competitive advantages of the solar thermal UHV collector is its ability to minimize its thermal losses to the ambient surrounding. This fact allows the solar thermal system that is being integrated with the UHV collector to exploit even the lowest levels of irradiance (diffuse light) for heat production. The applicability of the SRB Energy collector in climates with high solar irradiation, but also in areas with high portion of diffuse daylight - like central and northern Europe - turns the solar thermal UHV collector into a reliable solution when high performance is needed.

Figure 2.1 The Ultra High Vacuum solar thermal collector

The uniqueness of the solar thermal UHV collector is shown in Figure 2.2. As can be seen the UHV collector is covered by a pile of snow, but a portion of sunlight still penetrates though this thick layer of snow and falls on the absorbers of the collector. The Ultra High Vacuum acts as a perfect insulation material, maintaining the generated heat inside the collector (indication of stagnation temperature 80˚C at the thermometer), while the snow remains in a solid state.

Figure 2.2 Stagnation temperature under harsh weather conditions

The Ultra High Vacuum collector could be characterized as a hybrid solar thermal collector that combines features of a flat plate, a vacuum tube and a compound parabolic collector. The UHV collector maintains the basic characteristics of a flat plate collector by using: i) absorbers and welded tubes and ii) the flat geometry of the panel. Although the concept of using vacuum as thermal insulation and getter materials as a mean to maintain the vacuum inside a solar collector is being used already, the solar thermal UHV collector upgrades this concept further. The UHV collector employs vacuum of 10-8 Torr (UHV) and non-evaporable getter materials. Finally, the integration of reflection mirrors under the UHV collector increases the intensity of sunlight that falls on the absorbers, which is a feature of the compound parabolic concentrating (CPC) collectors.

2.2. Characteristics of the collector A representation of the cross section of the solar thermal UHV collector is given in Figure 2.3.

Figure 2.3 Cross section of the solar thermal UHV collector (representation)

The most important technologies incorporated into the solar thermal UHV collector are listed below:

2.2.1 Ultra High Vacuum Vacuum as an insulation is being in use for a long time in vacuum tube solar thermal collectors, in commercial vacuum flasks and recently even as an insulation measure integrated to building materials [2]. As shown in Figure 2.4, the quality of vacuum can be categorized according to the level of “emptiness” from Ultra High Vacuum to atmospheric pressure, covering a range from 10-11 mbar to 103 mbar. In the same figure, the variation of the mean free path length in accordance to the level of vacuum is presented. From the kinetic gas theory [3], it is well known that when the density of a gas increases, the gas molecules become closer to each other. Therefore, it is easier for the molecules to collide with each other and with the walls of the chamber, causing variation at the energy content of the molecules. This variation can be equivalent to heat transfer either from one side of the chamber to the other, or from the inner bulk of the chamber towards the external walls. On the other hand, by decreasing the density of a gas inside a finite chamber, the collisions among molecules decrease (high value of free mean path), and therefore there is no exchange of energy among the molecules and the chamber walls.

Figure 2.4 Relationship between the quality of vacuum and the mean free path length (source: http://www.uic-gmbh.de/en/basics/vacuum-technology.html)

During the production process of the solar thermal UHV collector, the collector is made vacuum tight to the level of 1.33x10-8 mbar (10-8 Torr), that equals to a perfect thermal insulation. By creating UHV conditions inside the panel, the concentration of the air molecules decreases radically -over 10 orders of magnitude- compared to atmospheric air (760 Torr). The low concentration of molecules prevents the heat transfer via the conduction and convection mechanisms, from the absorber to the glass and to the metallic frame of the collector. Therefore, the heat losses of the solar thermal UHV collector to the surrounding through the conduction and convection are eliminated. In this way the captured heat from the solar beams is being “trapped” inside the collector, and especially at the absorber and thereafter at the cooling tubes of the collector. Due to the presence of Ultra High Vacuum inside the collector, limitations to the manufacture materials are applied according to the principles of the vacuum technology [4].

2.2.2 Getter pump The use of getter pumps in order to maintain the vacuum inside sealed chambers has taken place for decades not only in industry, but also in research activities. Some examples are the thermionic valves (electron tubes) and the applicability of the getter materials to particle accelerators. Getters are materials with the ability to react with free gases and form chemical compounds at their surface. In this sense a getter, acts as a gas “pump” capturing the surrounded gas molecules. It is worth mentioning that the ability of the getter pumps is finite and it stops functioning, when the material becomes saturated by the absorbed gas molecules. There are two types of getters, the evaporable and the non- evaporable (NEG). Although, in the first category chemical compounds are formed and stabilized at the surface of the getter, the Non-Evaporable Getters are able to restore their saturated surface by diffusing the absorbed gasses at their bulk [5]. This property of the NEG is feasible by providing heat to the NEG pumps.

The Non Evaporable Getter (NEG) technology is one of the most important examples of maintaining UHV conditions in a sealed chamber. Due to the nature of the NEG pumps their ability to absorb gasses can be diminished. When declining, the pumping speed of NEG can be easily restored, by reactivating the getter material with a thermal process called reactivation. Non-evaporable getter (NEG) film coatings were developed at CERN to provide linear pumping for vacuum chambers and to realize ultra-high vacuum at low temperatures [6].

The same technology has been applied to the solar thermal UHV collectors with a target to ensure the UHV inside the panel [7]. Moreover, due to the high temperatures that may be experienced by the absorbers of the collector (e.g. stagnation temperature of 400˚C), molecules from the coating of the absorbers may be gasified and diffused internally to the panel, fact that decreases the level of vacuum inside the sealed collector. In order to maintain the UHV inside the panel for more than 25 years, each collector contains a NEG pump which absorbs the gasses inside the collector. A novel feature of this technology is the ability of the surface of the NEG pump to be regenerated using solar irradiation.

2.2.3 Metal -to- glass welding The creation of a sealed chamber is of primary importance for the operation of the collector, since the level of the Ultra High Vacuum inside the collector has to be maintained during the life span of the collector. Furthermore, the collector should use as much glazing surface as possible, so the maximal amount of available sunlight can reach the absorbers. In order to accomplish the features mentioned above, a special welding technique between the metal and glass surfaces takes place during the manufacturing process. In this way, the conservation of the vacuum inside the panel and the normal operation of the getter pump are ensured during the 20 years life span of the solar thermal UHV collector.

2.2.4 Selective surface coating A special treatment with black-Cr selective coating at the surface of the absorbers of the collector is being realized. The selective coating allows the maximal possible sunlight to be absorbed, but also prevents the emittance of the absorbed energy from the absorbers to the ambient, which is translated

18 to thermal losses through radiation. The absorbers of the solar thermal UHV collector are characterized by high absorbance (92%) and low emissivity (3.5%) of visible and infrared radiation, respectively.

2.2.5 Available configurations In order to increase the cost effectiveness of the collector and to achieve a high efficiency at high operation temperature, it is possible to integrate reflection mirrors to the collector. In this way, solar beams can be absorbed also from the back side of the collector resulting in an increase of the absorbed energy per m2 of collector surface. There are several possible reflector configurations that can be integrated to the collector, depending on the desirable increased concentration of sunlight. The collector configurations that are being used are: a) cylindrical mirrors which are able to reach a concentration factor of 2 (referred as a C1 configuration) and b) a combination of cylindrical and lateral mirrors with a concentration factor of 2.68, known as C2 configuration (Figure 2.5).

Figure 2.5 Collector configurations with a concentration factor (left to right) 1 (C0), 2 (C1) and 2.68 (C2)

2.2.6 Performance of the solar thermal UHV collector It is well known that the optical efficiency of a solar thermal collector is influenced by several variables. The most important of these are the physical constraints of the collectors that arise during their production (optical parameters), because they determine the applicability of the produced solar thermal collector in terms of performance, energy yield, generated power and operation temperatures. These parameters are known as the optical efficiency of the collector, η0, the first-order coefficient of the 2 2 collector efficiency, α1 [W/m ˚C], and the second-order coefficient of the collector efficiency, α2 [W/m ˚C2]. Moreover, the operation conditions and of course the local environment that the collector operate in plays important role, as well.

The optical parameters of a solar thermal collector (η0, α1 and α2) are determined by performance tests (e.g. according to the standard EN12975-2), that are conducted by certified test laboratories [8]. The trainee applied the parameters of the solar thermal UHV collector in the performance model of the solar thermal UHV collector, in order the performance curves to be extracted (Figures 2.6-2.9). The optical efficiency of a solar thermal collector is expressed as:

Τ− 2 Τ−avT amb ( avT amb ) ηη=01 − α− α 2 (2.1) IIref ref

2 with Τ−avT amb , the temperature difference under a reference irradiance e.g. Iref =1,000 W/m . The TT+ average temperature of the collector, T is defined as T = in out , with T and T , the inlet and av av 2 in out outlet temperature of the collector, respectively. The optical parameters of the collector configurations are determined according to the standard EN12975-2 and are presented in Table 2.1.

2 2 2 collector configuration η0 α1[W/m ˚C] α2 [W/m ˚C ]

C0 0.811 0.53 0.0095

C1 0.618 0.101 0.0048

C2 0.53 0.21 0.0042

Table 2.1 The optical parameters of C0, C1 and C2

The optical parameters extracted by the standard EN12975-2 (Table 2.1), have been analyzed with target the solar system designer to obtain a deeper understanding of the described technology. The 2 curves in the following figures (2.6, 2.7 and 2.8) are illustrated under an irradiance Iref=1,000 W/m , ambient temperature Tamb=10˚C, and with a temperature difference between the inlet and outlet temperature of the collector ΔΤ=20˚C. Figure 2.6 presents the characteristic efficiency curves of the solar thermal UHV collector for the three available configurations (C0, C1 and C2).

1.00 0.90 C0 0.80 C1 0.70 C2

η 0.60 0.50 0.40 efficiency, 0.30 0.20 0.10 0.00 10 30 50 70 90 110 130 150 170 190 210 230 250 270 290 310 330 350 370 390 Τ -T [˚C] av amb

Figure 2.6 Optical efficiency of the solar thermal UHV collector (C0, C1 and C2) in function of the temperature difference

Due to the peculiar geometry of the configurations C1 and C2, it is useful to determine the performance difference of each configuration, based on the definitions of the absorber, aperture and gross surface area (m2) that are widely used in solar thermal technology. In Figure 2.7, the optical efficiency of 2 configuration C1 is presented as function of the different terms used for the collector surface area (m ).

1.00 absorber 0.90 aperture 0.80 gross 0.70

η 0.60 0.50 0.40 efficiency, 0.30 0.20 0.10 0.00 10 30 50 70 90 110 130 150 170 190 210 230 250 270 290 310 330 T -T [˚C] av amb

2 Figure 2.7 Performance curves of the C1 configuration in regards to the absorber, aperture, and gross area [m ] of the solar thermal UHV collector

Since the operation of a solar thermal installation depends upon the presence of sunlight, the variation of the collector efficiency due to the intensity of sunlight is crucial for the performance of the overall installation. Figure 2.8 illustrates the efficiency of the C1 configuration for a range of solar irradiance levels (100-1,000 W/m2).

1.00 100 200 300 400 500 0.90 600 700 800 900 1000 0.80 0.70

η 0.60 0.50 0.40 efficiency, 0.30 0.20 0.10 0.00 10 30 50 70 90 110 130 150 170 190 210 230 250 270 290 310 330 350 370 390 Tav-Tamb[˚C]

Figure 2.8 Variation of the performance of the solar thermal UHV collector (C1) at different irradiance levels (100-1,000 W/m2)

2.2.7 High stagnation temperature Stagnation temperature is the temperature at which the heat generation of the collector is in equilibrium with the heat losses of the collector to the ambient, when heat is not delivered from the solar collector to the system (absence of thermal loads). The stagnation temperature is a characteristic of each collector and is function of the local climate conditions (irradiance and ambient temperature). At common solar thermal collectors, stagnation temperature can be up to 200˚C, while the solar thermal

UHV collector depending on the applied mirror configuration can reach a stagnation temperature up to 400˚C.

2.2.8 Operation temperature and pressure The main temperature range for which the UHV collector can provide heat at high performance is from 30˚C to 200˚C, depending on the location of the installation, the system application and the applied configuration of the collectors. The collector operates in flow rates from 150 l/h to 3,000 l/h, while the maximum fluid pressure that can be applied inside the cooling tubes of the collector is 10 bar. Due to the ability of the UHV collector to reach operation temperature up to 200˚C with a relatively high efficiency and stagnation temperatures up to 400˚C, the collector is suitable for industrial and heavy duty energy demanded applications.

2.3 Comparison to other solar thermal collectors The vast majority (around 90%) of solar collectors utilized worldwide at solar thermal installations employs either flat plate, or vacuum tubes collectors [9]. The comparison of the solar thermal UHV collector to the two categories mentioned above is of high importance, in order to highlight the competitive advantages of the former.

Figure 2.9 presents in dotted lines the characteristic curves of a C1 UHV collector, a vacuum tube and a 2 flat plate collector, under irradiance of 1,000 W/m and ambient temperature Tamb=10˚C. The temperature difference between the inlet and outlet temperature of the collectors has been established as ΔΤ=20˚C. Assuming a fluid inlet temperature Tin =50˚C (Tav=60˚C and Tav-Tamb=50˚C), the optical efficiency of the three solar collectors ranges at the level of 60%. Assuming a higher Tin=90˚C (Tav=100˚C and Tav-Tamb=90˚C), the optical efficiency of the UHV collector remains over 55%, while the values of the vacuum tube and flat plate have been decreased to 45% and 35%, respectively. In a similar way, if higher operation temperatures are needed, for instance Tin=150˚C (Tav=160˚C and Tav-Tamb=150˚C), the efficiency of the UHV collector is 50%, the performance of the vacuum tube has been decreased to 20% and the flat plate collector is experiencing a stagnation mode.

Figure 2.9 Variation of the optical efficiency at 300 W/m2 (solid line curves) and 1,000 W/m2 (dot style curves)

The above comparison took place under an irradiance of 1,000 W/m2. As mentioned before, the unique features of the solar thermal UHV collector, makes the collector suitable for climates where high percentage of diffuse light is available (Central and Northern Europe). At these high latitudes, irradiance at the level of 1,000 W/m2 is not experienced very often through the passage of a calendar year, fact that affects the performance of a solar thermal collector. Figure 2.9 demonstrates also the performance variation of the examined solar thermal collectors, under irradiance of 300 W/m2 (solid lines). Using once again the example with the reduced temperature Tav-Tamb=70˚C, as expected, the decrease of the available solar irradiance results in a clear decline of the optical efficiency of the collectors. Under 300 W/m2 (solid line curves), the optical efficiency of the solar thermal UHV collector decreased to 55%, compared to the value of 60% under 1,000 W/m2 (dot style curve). In sharp contrast, the optical efficiency of the vacuum tube collector has been decreased to less than 15% and the flat plate collector is on stagnation.

In addition to the comparison of the performance curves of the solar thermal collectors, Figure 2.9 illustrates the distribution of the solar irradiance in hours through a calendar year (8,760 hours) in Eindhoven, The Netherlands (central Europe). During the year, there are only 82 hours that the solar irradiance is ranges to 900-1,000 W/m2, while the hours that the irradiance is “around” 300 W/m2 exceed the 800 hours annually (200-300 W/m2 and 300-400 W/m2 ranges). This fact, affects the overall performance of the installation that the solar collectors are integrated into, since – in principle – the installation should operate year around. Also, it should be mentioned that due to the distribution of the solar irradiance during the day and seasonal cycle, the performance of a solar thermal collector should be examined in function of the irradiance variation, instead of taking only into account its performance at 1,000 W/m2.

2.4 Applications The UHV solar thermal collector is able to deliver heat in a wide range of industrial and building applications covering the energy demand for heating and cooling. The scale of the applicable solar thermal systems can be ranged from domestic and district installations to heavy duty industrial processes. Some of the possible applications are: i) Heating processes from 30˚C to 200˚C can be covered by the UHV collectors maintaining a high level of performance and providing the end user with the needed heat. ii) Solar cooling installations with single or double effect absorption chillers that operate on a temperature level of 70-90˚C, with a COP 0.6-0.7 and 150-180˚C with a COP 1.1-1.2, respectively.

A number of projects for which the solar thermal UHV collectors have been utilized as a driving heat source, has already been in operation:

Solar cooling and heating at HEFAME group, Spain Due to its high energy consumption, the pharmaceutical company HEFAME (Hermandad Farmacéutica del Mediterráneo) has adopted a hybrid solar installation, using the solar thermal UHV collectors and PV panels. This innovative project, under the name of MEDICOOL is the biggest of this type in the world, generating thermal power of 2.2 MWth and cooling power of 1.5 MWth. The solar thermal UHV

23 collectors are used as heat source to provide an absorption chiller with thermal power. The solar field consists of more than 3,200 m2a of UHV collectors that heats up thermal oil to 110˚C .The annual solar yield of the solar field is 3 GWh and the annual production of cooling is 2.1 GWh.

Industrial cooling process at RNB Cosmetics, Spain The company RNB Cosmetics exploits cold water to control the cooling demand of the chemical reactors in its production line. A solar field consisting of 130 UHV collectors, produces hot water at 100°C that is delivered to an absorption chiller. The chiller exploits the solar thermal energy to generate chilled water at 7°C. The chilled water is then stored in a tank that feeds the cooling processes, minimizing the use of conventional chillers that will work only when the cooling demand is higher than the capacity of the solar system. The annual solar energy yield is 590 MWh and the energy delivered by the absorption chiller is 413 MWh/year.

Industrial heating process at Colas Swiss Holding, Switzerland The asphalt manufacturer Colas has integrated a solar field with the UHV collectors, in order the latter to heat the bitumen storage tanks of an asphalt production plant in Geneva. In this type of industry, the bitumen has to be maintained in a liquid state; therefore, it is stored in well-insulated vessels that facilitate the liquid bitumen at a temperature of 180˚C. The installed solar field consists of 20 solar 2 thermal UHV (80 m aperture) delivering thermal power of 38 kWpeak. The annual generation of heat by the solar thermal installation is 26 MWh.

District heating at the Geneva Airport, Switzerland Located on the roof of the main terminal of the Geneva Airport, 282 (aperture area of 1,139 m2) solar thermal UHV collectors are installed, in order to provide the west terminal of the airport with thermal energy for heating and air conditioning (Figure 2.10). The solar field is heating up thermal oil at 130˚C for the distribution system of the International Airport of Geneva and supplies the airport with hot water and space heating. In summer, the solar heat will run an absorption chiller to cool the buildings. The annual production of energy delivered for heating and cooling is 70 MWh and 300 MWh, respectively. The annual solar yield of the solar field is 590 MWh.

Figure 2.10 Part of the installation of the solar thermal UHV collectors at the Geneva airport

Chapter 3. Solar thermal systems with solar thermal UHV collectors

3.1 Introduction Solar thermal technology is characterized by a wide variety of system configurations, a fact that makes it difficult to formulate a general method that fits all possible solar thermal applications. The reason for the various system combinations, stems from the quite wide range of applicable processes that the solar thermal technology can be integrated with. Each solar thermal installation is customized for the specific needs of a project; therefore, solutions that fit in one situation, may not be suitable for another installation. In this chapter guidelines for a generic system using the solar thermal UHV collectors will be presented, while emphasis will given to the solar field of the installation.

The solar thermal systems that the UHV collectors are being integrated to, follow the same basic layout of a typical solar thermal system. Figure 3.1 shows a typical simplified scheme of a hydraulic installation with the UHV collectors.

Figure 3.1 Typical scheme of a hydraulic installation with UHV solar thermal collectors

In the primary loop (or solar loop), the solar thermal UHV collectors absorb solar energy and deliver it to the transport fluid, which circulates in the overall installation. The transport fluid accumulates the generated heat in the storage tank and feeds the energy to the heat exchanger. In the secondary loop, another transport fluid takes the heat from the exchanger and transfers it to the thermal loads. The circulation of the transport fluid in both loops is accomplished by circulation pumps; thus appropriate system control is needed for normal operation of the installation. The heat exchanger, depending on the operation temperatures, can be installed either as an external heat exchanger (e.g. plate or double-pipe heat exchanger), or as a serpentine integrated in the buffer tank. Depending on the nature of the process, separation of the primary loop from the rest of the installation may be needed (open or closed- circulation loop). Moreover, the operation pressure and temperature determines the selection of the working fluid.

3.2 Design considerations and system requirements

3.2.1 Concept of solar thermal systems In a solar thermal system three main functions can be identified: energy generation (transformation actually), storage and consumption. The first one is realized when solar radiation is “captured” by the solar field. Naturally, this function can be considered the heart of each solar thermal installation and is discussed in detail in sub-chapter 3.2.3. On the other hand, the generated energy is oriented to cover the demand side of the end user; therefore the former has to “serve” the demand profile of the energy consumer. Discussion about the thermal loads is covered in sub-chapter 3.2.2.

In most of the cases, storage of the generated heat is a necessity, in order the system to be less sensitive to the solar irradiance variation. In order to achieve the coupling between the energy supply and demand during the operation cycles of the installation, storage tanks can be utilized. By buffering supply and demand, energy is added to the storage tank whenever solar irradiance is available and removed when user demand increases. Crucial parameters for sizing a storage tank are its maximum temperature, the total volume and the insulation of the tank. The tanks should be carefully sized, in order to avoid low temperatures inside the tank and to minimize the heat losses to the ambient.

3.2.2 Consumption - Thermal load For every supported process in a solar thermal installation, one of the most important parameters for designing the integrated solar thermal system is the thermal load that the system will be able to handle. The fluctuation of the thermal load (base and peak load) and its dependence on time (load profile) may vary significantly, depending on the nature and the targets of the process. In order to optimize the design of a solar thermal system it is important to understand the loads handled by the system and the available resources to be met. The thermal load and load profile of the solar thermal system should be available on a daily, weekly and annual basis. Variation of the heat demand throughout the year or the existence of a demand during only specific periods of the year affects the design of the solar system thus making each solution unique (optimum solution). These variations can be caused by scheduled maintenance, vacation periods, or short fluctuations on the consumption pattern such as during a weekend.

Consideration of the operation temperature at which the thermal load is coupled to the solar installation is one of the major aspects for the design of a solar thermal installation and it should be determined as accurate as possible. The operation temperature of an examined thermal load depends on the nature of the latter and it may be either time dependent, or simply a fixed temperature difference (ΔΤ) between the inlet and the outlet of the thermal load.

Furthermore, two other major aspects for determining the application of the installed solar thermal system are: a) the user requirements and b) the targets set by the specific request. Special conditions to be fulfilled in terms of safety or legislations may largely shape the final installation. The set targets of the solar thermal system that is under investigation, create the constraints that should be followed during the design and establish the boundaries either from a technical, or financial point of view. Among others, the targets of a solar thermal installation may determined by the desirable solar fraction of the

26 overall installation or coverage of the base load. In addition, attention should be given to the match between the generated energy and the energy demand in the event there is lower need for storage; thus, thermal losses by storage can be minimized. Last, but not least, if the solar thermal installation is combined to an existing infrastructure, the coherence and the normal operation with the rest of the installation are of high importance.

3.2.3 Generation - Solar field Solar field As it becomes clear, the heart of a solar thermal system is the solar field. Designing the solar field is a function of the desired mass flow, pressure drop, applied temperatures at the solar field (ΔΤ) and the configuration of the collector loops (series and parallel connections), with the target to maximize the energy productivity of the solar field. For the installations with the solar thermal UHV collectors the appropriate collector configurations among C0, C1 or C2 (as discussed in Chapter 2) should be chosen in order to optimize the performance of the solar thermal system depending on the target temperature and the solar irradiation potential. In most cases the availability of surface for placing the solar collectors is the main factor if a solar thermal project will proceed or not. Shading from obstacles (local constraints) may increase the necessary surface of the solar field and may determine in high portion the choice of the collectors’ tilt and orientation.

The solar thermal UHV collectors may be connected together in series, parallel or a combination of series and parallel arrangements as shown in Figure 3.2. The number of the strings (parallel connection of the strings) is determined from the desired mass flow of the solar field which is coupled to the desired power of the solar field (green arrow), while the number of the collectors connected at each string (connection in series) is determined by the desired temperature increase of the heat carrier (red arrow). Moreover, the final configuration depends on the geometry of the available area defined for the solar field mounting.

Figure 3.2 The configuration of a solar field is determined by the desired mass flow (green arrow) and the temperature increase of the heat carrier (red arrow)

Another variable that affects the configuration of a solar field are the friction losses realized from the circulation of the working fluid (heat carrier) that is achieved by a circulation pump. The pressure drop

27 inside the solar field should be maintained on the minimal possible value, since electricity consumption used for pumping, known as the system’s parasitic energy, influences the overall performance of the solar thermal installation.

The normal operation of the solar field is ensured by the control system of the installation. An example is the solar irradiance sensor that switches on the circulation pump under a determined irradiance level. Safety equipment such as an expansion tank is used, in order the heat carrier to overcome the density decrease that is experienced inside the solar field during the heat addition. Similarly, pressure relief valves are mandatory part of the solar field to avoid the existence of a pressure higher than the expansion tank’s ability to handle.

Heat Transport Fluid Since the UHV solar thermal collector is able to cover a wide range of applications (30˚C to 200˚C), the appropriate heat transport fluid (HTF) should be selected based on the working temperature of each installation. The most commonly working fluid used in processes where heat transfer is needed is water, because of its availability, the relevant low cost and its thermodynamic properties. Furthermore, in solar thermal systems the HTF should remain in a liquid phase, in order to increase the heat transfer between the absorber and the cooling pipes of the collector. In order to maintain the water in a liquid state at temperatures above 100˚C, additional pressure has to be applied to the hydraulic system of the installation. By using thermal oils as an HTF, high temperature can be achieved, by maintaining the operation pressure of the circuit much lower compared to water. Installations with thermal oil can operate even under non- pressurized conditions (1 bar abs). Also the ability of thermal oils to reach high temperatures without vaporizing gives the advantage to the solar field designer to utilize the bulk of the thermal oil as temporary energy storage medium (thermal mass of the HTF). Additionally, the choice of the HTF is also affected by the local climate conditions. In order to prevent the formation of ice at the solar field during the winter months, water mixed with glycol is utilized to lower the freezing point of the working fluid under the desired conditions.

Overproduction of generated heat In a solar thermal installation with UHV collectors, precautions concerning the overproduction of heat have to be taken into account. The overproduction can be the result of the energy demand pattern of the installation (e.g. absence of thermal loads at weekends), or due to unexpected events. Such reasons can be a failure of the circulation pump of the secondary loop, or the buffer tank being fully loaded, causing an excess of generated heat, or the existence of a temporary disconnection between the solar field and the thermal loads due to maintenance. In general, different measures can be taken to avoid the excess of generated heat from the solar field. Depending on the best available solution, the alternatives can be: 1. Switching off the circulation pumps under a controlled manner to avoid simultaneous stagnation of the solar field. 2. Draining the working fluid out from the primary loop (solar field), with a risk that stagnation may occur. In this case, an extra tank for storing temporarily the working fluid may be needed. 3. Covering the absorber areas of the collectors, in order to prevent solar irradiation to be absorbed.

4. Utilizing of an additional buffer tank to store the extra heat, in case overproduction occurs quite often and other thermal loads are available. 5. Utilizing a heat rejection device for extracting the excess heat from the solar system. In this case the solar field, instead of feeding energy to the process, delivers the generated energy to a liquid/air exchanger. The cooler should be designed based on the peak power of the solar field, while its operation adds parasitic electric power consumption to the total installation.

3.2.4 Collection of the necessary information Following the discussion from the previous sub-chapters, in order to design a solar thermal system, a number of variables and information about the targets and the functionality of the solar installation have to be set. Three key factors that heavily influence the design of a solar thermal system are: a) the available space for the solar field (land or roof availability), b) the energy demand that has to be covered, and c) the available capital investment for the solar installation. Information that determines the overall picture of the current project under investigation is the description of the thermal loads/process and available drawings or information that could affect the existing installation. Figures concerning the existing equipment and the use of conventional energy would be helpful to determine the added value of the energy savings of the solar thermal system, or even to upgrade the overall efficiency of the total installation by choosing the suitable integration point of the solar installation.

The project parameters discussed above can be extracted from a questionnaire (APPENDIX A), which has been developed during the traineeship period for commercial activities. The questionnaire allows the designer of the solar thermal installation to gain deep knowledge of the under investigation project and to suggest the optimal solution that fits to the project.

3.3 Model for the solar thermal system with the UHV collectors In the present sub-chapter, the model that is going to be used for the simulation of the solar thermal system integrated with the UHV collectors is presented. The system follows the layout described above and consists of a solar field formulated by the solar thermal UHV collectors, a buffer tank at which the generated heat is temporarily stored in, a heat exchanger between the primary and secondary loop, and finally the thermal load that consumes the generated energy by the solar field. In order to design the solar thermal system, firstly the solar field is designed according to the requirements of the thermal load under the local reference irradiance and then the buffer tank is being integrated in the system layout. Description for each of the four components of the simulated solar system and the guidelines followed are given below.

3.3.1 Design of the solar field Thermal load The starting point for the design of the solar system is to determine the characteristics of the thermal load. The thermal load is expressed in terms of power (e.g. kW) or mass flow (e.g. kg/s), accompanied by a temperature difference Tin,load and Tout, load (Figure 3.3). The load is considered to remain constant. The provided temperature difference and mass flow of the thermal load, will be used to determine the temperature difference (Tin,field and Tout,field) and the mass flow of the solar field, respectively.

Heat exchanger In the simulated system a liquid to liquid heat exchanger is employed to connect the primary to the secondary loop of the solar system. As shown is Figure 3.3, the heat exchanger is described by setting the four temperatures Tout,field , Tin,field and Tin,load, Tout, load, of the primary and the secondary loop, respectively. The two loops can be considered as the hot and the cold side of the heat exchanger, respectively. From the description of the problem the temperatures of the cold side (Tin,load and Tout,load) are considered to be known.

Figure 3.3 Design of the solar field, based on the targeted thermal load

A rule of thumb used for the design of heat exchangers suggests that the inlet temperature of the hot side should be 10oC higher than the outlet temperature of the cold side of the heat exchanger. In this way, the heat transfer from the hot side to the cold side is ensured even at the final stage of the heat exchanger device. Similarly, in the present model the inlet temperature of the hot side of the heat exchanger ( Τout, field ) is determined as:

 Τout,, field =TC out load +10 ( 3.1)

The only unknown temperature to be determined is Τin, field . This is realized by the effectiveness-NTU method, assuming the effectiveness of the heat exchangerε = 0.9 . The definition of the effectiveness, ε, relates the four temperatures of the heat exchanger as:

Τout,, field −Τ in field εε= ⇒ Τin, field = Τ out , field −( Τ out ,, field −T in load ) (3.2) Τ−out,, fieldT in load

After calculating the temperature Τin, field , the operation temperatures of the heat exchanger have been determined and the device is able to deliver the chosen thermal power to the thermal load. It is worth to highlight that the equation (3.2) is valid only when both sides of the heat exchanger are under the same mass flow. In the given model, this assumption is taken into account in order to determine the operation temperatures of the heat exchanger.

Solar field The solar field is designed with target the fluid that circulates inside the collectors to reach the desired temperature determined by the parameters of the thermal load. The overall solar field is a product of a designed string of collectors multiplied by the number of strings needed, with target to cover the demanded mass flow. Since the solar field is actually a number of strings connected in parallel, the inlet and outlet temperature of each string is the same with the inlet and outlet temperature of the solar field, respectively:

Tin,, str= T in field and Tout , str = T out , field (3.3)

Now the challenge of designing the solar field becomes a matter of designing one string of the solar field 2 This is realized under a reference irradiance level, Iref (W/m ), which is an empirical value that varies for each location and is selected based on the local solar potential (e.g. for Eindhoven, Iref =600 W/m2). It should be mentioned that the reference irradiance is being used only to size one string of the solar field. The number of collectors that formulate the string is a result of the desired temperature

Tout, str . The collectors are connected in series, thus the outlet temperature of one collector n is the inlet temperature of the collector n +1; thus solar energy is obtained from the fluid in steps. The energy  step, Qcoll , gained from each collector is given by:

 Qcoll=η coll,, ref IQ coll − los coll (3.4)

The term ηcoll provides the efficiency of the collector considering the inlet temperature of the collector

Τin , the local ambient temperature Tamb , the reference irradiance Iref and the optical efficiency of the collector, ηcoll, ref , under the reference irradiance and the optical parameters η0 , α1 , α2 of the collector configuration is applied, as provided in Chapter 1:

Τ− 2 Τ−inT amb ( inT amb ) ηηcoll,0 ref = − α1− α2 (3.5) IIref ref

Icoll (W/collector) in expression (3.4) stands for the total irradiance that is falling on the solar collector, 2 taking into account its aperture surface area Αap, coll and the reference irradiance Iref (W/m a of collector):

IIcoll= Α ap, coll ref (3.6)

 Qlos, coll in equation (3.7) represents the heat losses in the connection piping between the collectors. The value 4, stands for the four cooling pipes that each collector includes, while Rbtw is the thermal resistance of the piping materials:

−  TTin amb Qlos, coll =4 (3.7) Rbtw

Rbtw represents the conduction and convection losses of the connection piping between the n and n +1 collector, determined according to the model of the cylindrical layer:

D + sins ln 2 D 2 1 Rbtw = + (3.8) 2π Lkins hair A surf with s is the insulation thickness of the pipe, L the length and D the diameter of the cylindrical pipe, kins the thermal conductivity of the insulation material and h the convection heat transfer coefficient to air.

Substituting equations (3.5), (3.6) and (3.7) in (3.4), the energy step of each collector is given:

2 Τ− (Τ−T ) −  inT amb in amb TTin amb QIcoll =ηα0 − 1− α2,Αap coll ref −4 (3.9) IIref ref Rbtw

The number of collectors, λ , utilized in a string is determined by the requested temperature at the exit of the string, Tout, str , and can be expressed as:

λ :TTin,1λ+ ≥ out , str (3.10)

The total energy extracted by a string of λ collectors is:

λ:TTin,1λ+ = out , str  QQstr = ∑ coll,λ (3.11) λ

Besides the equation (3.10), other conditions that should be fulfilled for the design of the string are the overall pressure drop of the fluid not to exceed 1 bar, and also turbulent flow to be ensured inside the cooling pipes of the collectors with Reynolds numbers Re=8.000 and Re=10.000 for water and thermal oil, respectively.

After designing one of the strings of the solar field, the thermal power generated by the overall solar  field is the product of the power output of one string Qstr and the number µ of the strings needed, in order the solar field to operate with the desired mass flow:

 QQsf, i=µ str (3.12)

The number of the stings, µ , in a solar field are determined by:

 Qth, load µ : msf ,1µ+ ≥ (3.13) Cp(Τout,, field −Τ in field ) Since the solar field configuration and the total number of collectors have been determined, the total aperture area of the solar field, Aap, tot , can be derived, as well.

3.3.2 Model of the solar system After designing the appropriate solar field, the energy yield for the whole year is calculated in hourly values, i , based on the total aperture area of the collectors. The heat generated by the solar field, Qsf, i , in a given houri , is:  Qsf,, i= Qt sf i ∆ (3.14)

The energy generated hourly by the solar field, Qsf, i , can be expressed as:

QAsf, i= ap , totηη opt ,, i I glob i sf (3.15)

2 with I glob, i the hourly global irradiance per m that falls on the collectors and ηsf = 0.8 is assumed an overall efficiency factor of the solar field. The hourly efficiency of the solar field is approximated by considering the optical efficiency of one collector, ηopt, i . The hourly optical efficiency ηopt, i is calculated according to the geometric characteristics of the solar radiation, provided in APPENDIX B:

 Gb Gb Τ−av TT amb  Τ−av amb ηηopt,0 i = IAM + 0.95 +α1+ α2 (3.16) GGbd+ GGbd +  GGbd ++  GG bd

It is highlighted that the hourly efficiency of the overall solar field, is calculated by equation (3.16), and not by equation (3.9), since the inlet and outlet temperatures of the solar field are not available on a hourly basis, Tin,, field i and Tout,, field i , respectively. These temperatures are not available, since as presented above the design of the solar field takes place based the design of a string under the reference irradiance.

Furthermore, the hourly generation of energy Qsf, i can be aggregated on monthly values, Qsf, k , with k= [ Jan, Dec] :

QQsf,, k =∑ sf i (3.17) i with ξ the number of hours of each month, 744, 720 and 672 for the months with 31 days (e.g. January), the months with 30 days (e.g. March) and February (28 days), respectively.

Solar system A schematic overview of the solar system by integrating a buffer tank to the designed solar field is given in Figure 3.4.

Figure 3.4 Illustration of the model used for the solar thermal system with the UHV collectors with an integrated buffer tank

The solar factor, SFsf, k , that expresses the amount of energy generated by the solar field during a month k , Qsf, k , in regards to the heat demand of the thermal load, Qdem, k , for that month is given by:

Qsf, k SFsf, k = (3.18) Qdem, k

Moreover, the solar factor of a solar system, SFsys, k , is defined as the energy delivered by the solar system in a month k , Qdel, k , divided by the heat demand of the thermal load, Qdem, k , for that month:

Qdel, k SFsys, k = (3.19) Qdem, k

Buffer tank The solar energy captured by the solar field is accumulated to a cylindrical fully mixed buffer tank (Figure 3.5). The equilibrium of energy for the buffer tank is:

Qstored,,, i= QQ sf i − los i − Q del , i (3.20)

Figure 3.5 The model used for the buffer tank of the solar system

The solar system model has been designed in order the buffer tank to deliver to the thermal load, the capacity that has been considered by the problem description ( Qdel, i ). Solar energy is considered to be added to the tank on an hourly step i ( Qsf, i ), when the optical efficiency of the collectors is ηopt, i > 0.

Condition for the latter is the availability of the global solar irradiance, I glob, i > 0 :

then then if( I glob, i >0) →ηopt ,, i >0 →Qsf i is added to the buffer tank

The tank delivers energy to the heat exchanger, also on an hourly step i ( Qdel, i ), when the temperature of the buffer tank is higher than the inlet temperature of the heat exchanger’s hot side.

The buffer tank experiences thermal losses ( Qlos, i ) that are given by:

TTbuffer,, i− amb i Qlos, i = (3.21) Rbuf

Tbuffer, i is the temperature of the buffer tank andTamb, i the temperature of the ambient varying on an hourly base, i .

The thermal resistance, Rbuf , of the cylindrical layer including the conduction and convection heat losses is given by:

D + sins ln 2 D 2 1 Rbuf = + (3.22) 2π Lkins hair A surf

35 s is the insulation thickness of the tank, L the length and D the diameter of the cylindrical tank, kins the thermal conductivity of the insulation material and h the convection heat transfer coefficient of air.

In order to determine the hourly variation of the temperature of the buffer tank Tbuffer, i , the energy balance of the buffer tank (3.16) will be used:

dT mCpbuffer, i = Q − Q − Q (3.23) dt sf,,, i los i del i with m andCp , the mass and the heat capacity of the storage medium inside the buffer tank, dT respectively. Using Euler integration the temperature derivative buffer , becomes dt dTbuffer T buffer, i− T buffer ,1 i− = with a time interval ∆=t1 hour , so (3.19) becomes: dt ∆t

mCp() Tbuffer, i− T buffer ,1 i− = Q sf , i − Q los , i − Q del , i ⇒ QQ− − Q TT= + sf,,, i los i del i (3.24) buffer, i buffer ,1 i− mCp

A design guideline used to select the volume of the buffer tank is that the temperature Τbuffer, i should not be more than 25 to 30˚C, higher than the outlet temperature of the cold side of the heat exchanger,

Tout, load :

TTbuffer,max= out , load +30 ° C ( 3.25)

The volume of the buffer tank is given by:

π D2 VL= (3.26) buffer 4

The design of the solar field and the model of the solar system described above have been implemented with the Microsoft Excel tool of the Microsoft Office suite. The model has been developed by SRB Energy and has been provided to the trainee by SolCalor, as the man tool for designing solar thermal systems with the solar thermal UHV collector. The model has been applied in the Microsoft Excel tool with target to ensure the fast implementation of the model and the adaptation of the performed calculations for different solar system designs.

3.3.3 Validation of the applied model In order to validate the design of the solar field described in paragraph 3.3.1, the performance curves presented in paragraph 2.2.6 will be applied with target to extract the same energy output by a reference solar field. For the validation, a solar thermal UHV collector with configuration C1 will be

2 applied under a fixed irradiance of 1,000 W/m and under a constant ambient temperature Tamb=10˚C, throughout the year. The inlet and outlet temperatures of the solar field are considered Tin,field=70˚C and

Tout,field=90˚C. According to the performance curve of Figure 2.6 the efficiency of the C1 configuration is 0.59. Considering the irradiance remains constant throughout the year at 1,000 W/m2 day and night, the hypothetical thermal energy generation of the C1 collector would be 20,673 kWh. According to the model that describes the design of the solar field, the annual thermal energy generated by the solar thermal collector under the unrealistic assumption that the irradiance remains constant at 1,000 W/m2 throughout the year, is 20,880 kWh. The error of the model used for the design of the solar field compared to the analog calculated before is almost 1%, a percentage that is acceptable in order to use the model described at the present subchapter.

Chapter 4. Solar heating and cooling: Integration of the solar thermal UHV collector to a building

4.1 Case description

4.1.1 The institute A solar thermal UHV collector system will be designed for a reference bloc of buildings that belongs to an institute in Erlangen, South Germany. Due to the functionality of the institute (laboratories, clean rooms and offices) there is a high electricity consumption through the whole year. In 2011 the whole institute consumed around 4 GWh of electricity, from which more than 25% was used to cover the cooling demand of the institute. In 2011 the consumption of thermal energy supplied by the local district heating network was 2 GWh. The institute consists of 4 buildings: a so-called old building, 2 clean rooms with a total surface area of more than 1,500 m2 (a small clean room of 500 m2 and a big clean room of 1,000 m2) and a new building with a total surface area of 1,600 m2 (Figure 4.1).

Figure 4.1 The buildings of the institute

The management of the campus faces two challenges related to the energy consumption of the institute:

already on its full capacity and it is not possible to provide the newly scheduled building with thermal energy. The solar thermal collectors could provide the building with heat, in order to cover its heating demand.

4.1.2 Cooling system of the institute The cooling system of the institute consists of a centralized installation that distributes cooling all around the institute. Cooling is generated by three Trane compressor chillers of two types, one RTHA

450 (Pcool,max= 814 kW, Pel,max= 234 kW) that is used for air conditioning in summer and two RTHA 215

(Pcool,max= 469 kW, Pel,max= 119 kW) that are used in turns over the whole year for industrial cooling of the clean rooms. The RTHA 450 chiller feeds a low temperature circuit (Tfeed=6˚C , Tret=12˚C), while the two RTHA 250 chillers provide chilled water on the circuit that operates in temperatures Tfeed=12˚C and

Tret=17˚C.

RTHA 450 The RTHA 450 chiller is used for air conditioning in summer, while in winter the cold air for air conditioning is provided by the outside environment (free cooling). The maximum cooling capacity of the chiller is 814 kW and it provides the terminal systems of the buildings with chilled water at 6˚C

(Tret=12˚C) and covers cooling loads of 795 kW. The two cooling towers of the installation have the ability to dissipate thermal power of 1 MW (2x500 kWth) to the environment, while the electrical consumption of them is 234 kWe.

RTHA 215 The two RTHA 215 cooling machines cover the cooling demand of the production processes in the clean rooms and the laboratories. The two machines never run at the same time and they take turns in service. Each RTHA 215 delivers chilled water at 12˚C (Tret =17˚C) and covers cooling loads of 513 kW. The cooling capacity of each chiller is 440 kW and the ability of the cooling towers to dissipate heat to the environment is 1 MW, while the electrical consumption of the cooling tower is 238 kWe. Part of the rejected heat of this cooling cycle is recovered to a storage water tank of the so called new building in order to be used at the heating circuit of the latter. This is happening only if there is any capacity left in the storage tank to be used for heat radiators of the building, while the rest is fed into the cooling tower of 1 MW.

4.1.3 Heating system of the new building The new building of the campus was built in 2012 to support the activities of the institute with an additional 1,600 m2 of useful surface area and its heating demand is covered by the local district heating network. The connection of the heating installation to the district heating system is taking place through a 100 kW heat exchanger (Figure 4.2). From the side of the district heating the water circulates at

Tfeed=80˚C and Tret= 60˚C, while from the side of the building at Tfeed=50 and Tret=70˚C. The installation delivers thermal energy to the four thermal loads of the building air conditioning, tap water, radiators and concrete core activation. The central heating circuit is separated in a low and a high temperature branch. The first one feeds with energy to the radiators and the core activation component, while the high temperature branch supplies energy to the air conditioning system and the hot water loads.

Space heating is provided to the building by radiators that operate at low temperature (Tfeed=45°C,

Tret=40°C) and by an air duct installation that operates at the higher temperature (Tfeed=70°C, Tret=50°C). The capacity of the loads is 25 kW and 56 kW, respectively. The demand of the building in hot tap water is covered by a 500 L tank that is connected to the high temperature branch of the installation

(Tfeed=70°C, Tret=50°C) with a capacity of 30 kW. At the thermal loads of the building is included a concrete activation component (Tfeed=27°C, Tret=20°C) with a nominal load of 25 kW. The thermal loads of the installation are presented in Table 4.1.

The installation of the building includes also a hot water storage tank with a capacity of 1,000 L. This tanks stands as a backup of the district heating network and recovers the heat rejected by the cooling circuit of the RTHA 215 chillers (Tout=45˚C, Tin=40˚C). When the water temperature of the tank reaches 45°C, the tank supports the space heating system of the new building (radiators); thus saving energy from the district network. Since the maximal temperature of the buffer tank is 45˚C, the loads that can be covered by the heat recovery system are limited to the ones connected to the low temperature branch (space heating and core activation component).

Figure 4.2 Heating installation of the building

type ΔΤ(˚C) power demand (kW) comment space heating 70/50 56 air ducts (high temperature branch) tap water 70/50 30 tank of 500 L (high temperature branch) space heating 45/40 25 Radiators (low temperature branch) activation component 27/20 25 heating at winter (low temperature branch) (16/20) (50) (cooling in summer) Table 4.1 The heating loads of the building

4.1.4 Proposal for using the UHV collectors The available roof area that can be used for the solar fields is shown in Figure 4.3. The area in blue is approximately 240 m2, while the area in orange is around 1,400 m2. Considering that each solar thermal collector needs around 5 m2 of gross surface area for installation, plus approximately another 5 m2 to overcome the over shading among the collectors each other, the big solar field can facilitate around 140 collectors (in orange color), while the small solar field around 24 collectors (in blue color).

Figure 4.3 Available roof area for the solar fields for space heating of the new building (in blue, 240 m2) and for solar cooling of the institute (in orange, 1,400 m2)

The purpose of the following investigation is the solar thermal UHV collectors to operate as a heat source for solar heating and solar cooling, with target to evaluate the applicability of the solar UHV collectors for the above applications in South Germany. Due to the absence of the energy consumption patterns of the institute for heating and cooling, the following designs of the solar systems will be based on consumption estimates that have been taken into account for each case. The coverage of the heating demand of the new building is going to be investigated in sub-chapter 4.2, while solar cooling for the centralized cooling installation of the institute will be covered in sub-chapter 4.3. In more detail, in the first case the solar thermal UHV collectors will substitute partially the district heating network by covering the demand of the new building for space heating, while in the second case a solar cooling installation will be exploited to reduce the capacity of the existing compressor chillers.

4.2 Solar heating

4.2.1 Heating profile of the new building As discussed earlier in Chapter 3, starting point of the design of a solar thermal system is the good understanding of the design requirements that are related to the applicable thermal load. In the given project and due to the absence of the energy consumption pattern - of the so called new building - for heating, a simplified prognosis of the energy demand will be realized, based on consumption estimates and according to the available information. Although the above fact influences the overall performance of the solar thermal system compared to the energy demand, a well known method will be applied to determine the heating demand profile of the building. The heating demand profile of the new building

42 will be estimated by the Heating Degree Hour (HDH) method. In this method the heating demand profile is derived from the difference between a base temperature and the local average ambient air varied on an hourly basis. The base temperature, Tbase, is a reference outdoor temperature above which the building does not require heating; thus the Heating Degree Hour value is a positive quantity. On the contrary, when the ambient air temperature, Tamb , is greater than the Tbase, heating is not required and the HDH quantity is zero. Degree-hour-based calculations can be greatly affected by the base temperature of the degree days used. A base temperature used widely for calculations is Tbase =15.5°C.

The hourly ambient air temperature, Tamb,i , is provided by the local climate data (Erlangen, South Germany). The Heating Degree Hours (HDH) is determined as:

Tbase− T amb,, i , for Tbase > T amb i HDHi =  (4.1)  0 ,for Tbase ≤ T amb, i

 The power demand of the heating system, Qheat, i , is considered:

 Qheat, i= U overall ⋅ A ⋅ HDHi (4.2)

2 i the time level (hourly step), Uoverall =0.5 W / ( mK) is the overall heat transfer coefficient of the

2 building, Am=1,474 is the external building surface, TCbase =15.5 ° the ambient temperature at which heating is not needed inside the building and Tamb, i the hourly ambient temperature derived by the local weather data.

The annual heating demand of the building, Qheat, an , is the summation of the hourly values, Qheat, i is:

8,760

QQheat,, an =∑ heat i (4.3) i

 with Qheat,, i= Qt heat i ∆ and ∆=t1 hour .

The annual heating demand of the new building is estimated to 50,859 kWh and its distribution throughout the year is given in Figure 4.4 (the monthly figures are provided in APPENDIX D1). As expected, the heating demand increases during the winter months with a maximum of almost 9 MWh in January, while the least demand for heating is experienced during July with approximately 0.5 MWh.

10000 Heat Demand [kWh] 9000

8000

7000

6000

5000 kWh 4000

3000

2000

1000

Figure 4.4 Distribution of the demand for space heating of the new building throughout the year

In order to choose the designed capacity of the solar thermal system a better understanding of the distribution of the power demand should be derived. This is realized by the load duration curve of the estimated heating demand (Figure 4.5). According to these calculations the maximal heating load is 22.6 kW, while the heating system is going to operate over than 6,500 hours throughout the year. For 250 hours the thermal load will be more than15 kW, while for 2,000 and 4,500 hours will be more than 10 kW and 5 kW, respectively.

25.0

20.0

15.0

10.0 thermal load load (kW) thermal

5.0

0.0 0 0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 6 6.5 7 7.5 8 8.5 hours in thousands

Figure 4.5 The load duration curve of the heating demand

The solar thermal system will be designed with the target to cover the base load of the heating system. In the following section options of different base loads will be examined to better understand the operation of a solar thermal system with UHV collectors.

4.2.2 Solar thermal system The model that will be used for the calculations of the solar thermal system has been described in sub- chapter 3.3. For the design of the solar system oriented for solar heating, there are two main constraints that should be taken into account. The first is the fact that the demand for space heating in summer months is low compared to winter months, thus over-production of energy will occur when heat is not required for space heating. The second one is related to the available roof space of the building for the installation of the solar thermal UHV collector, which cannot exceed the 24 collectors. Under these constraints the cases that will be examined are, as follows:

1. The design of a solar thermal system with a limited generation based on the heating demand during summer months.

By studying this case it is possible to highlight the importance of the variation of the heating demand throughout the year and especially in summer months. The lack of high heating loads during summer combined with the large variation of the solar irradiance throughout the year in Erlangen, constraints the performance of the solar thermal installation in winter months. This comes in contrast with the fact that during the summer months the solar potential is the most favorable period of the year for heat generation. In principal, a solar thermal installation should be tailored based on the desired energy demand; thus the present design is constrained to cover the energy demand during the summer months June, July and August. As expected, during the winter period, when heat production for space heating is needed most, a solar installation consisting of a small number of UHV collectors cannot cover a substantial heating fraction.

2. The design of a solar thermal system that exploits all the available roof space of the building. (240 m2, 24 collectors).

In this case, the total available surface is utilized in order to identify the maximal amount of energy that can be extracted for the given building. By increasing the size of the solar field, a higher heating demand can be covered in winter, although overproduction of heat will occur in the summer season. The latter scenario will not be an issue, if the excess generated energy of the solar system can be consumed either by a heating or a cooling load.

Design for limited generation during the summer months

In Figure 4.4 has been shown that the heating demand of the building for space heating in summer months experiences low values (with a minimum in July, 465 kWh). This heating demand value is translated to the number of collectors that will be implemented at the solar field and at the present design the solar field will consist of four UHV collectors. Based on the load duration curve (Figure 4.5) and the limited number of collectors that will be utilized, the base load will be considered as 5 kW.

Moreover, in sub-chapter 4.1.3 is given that the operation temperature of the existing space heating installation is Tfeed=45°C and Tret=40°C (the radiators load). Combining the above elements, the solar system will be designed to deliver to the secondary loop 5 kW with Tout,load=45°C and Tin,load=40°C. The solar field will consist of four collectors connected in series and the generated heat will be ejected in a buffer tank with capacity Vbuffer=0.5 m3 (Figure 4.6).

Figure 4.6 Design of the solar system according to the minimal monthly energy demand

Exploitation of all the available roof space As discussed in paragraph 4.1.1 the institute is characterized by a high demand of cooling. Taking into account this characteristic and assuming the existence of a heat driven chiller (e.g. absorption machine), the over-production of heat in summer months could be directed to the absorption chiller. In this way, the generated heat in summer could be utilized and a higher solar fraction could be obtained in winter months when demand for space heating raises dramatically. In this case a condition that should be fulfilled is the solar thermal installation to operate on the same temperature level with the operation temperature of the absorption cooling system (e.g. 85˚C/80˚C).

For the design of this solar system all the available roof space will be exploited, utilizing 24 collectors. As discussed above, the operation temperature is assumed Tout,load=85°C and Tin,load=80°C. Due to the larger size of the solar field, the capacity of thermal power to be covered and the volume of the buffer tank used, will be investigated further (Figure 4.7).

Figure 4.7 Design of the solar system according to maximal roof space for the solar field

By applying a larger solar field (24 collectors compared to 4 collectors) under the same base load (5 kW), the pace of the generated energy increases significantly, while the heat consumption remains approximately the same (different operation temperatures). As discussed in paragraph 3.2.1, this mismatch between the generated and consumed heat can be adjusted by the buffer tank of the system. Since the solar field utilized in the solar system is 6 times larger, it is quite normal that the buffer tank that accumulates the generated energy will be influenced, as well. The change at the buffer tank is related to the ability of the tank to accumulate heat and it can be translated either by experiencing higher temperatures inside the buffer tank, or by increasing the size of the buffer tank. The first option would affect the normal operation of the installation, while using a larger tank is always function of the available space. An alternative could be to increase the capacity of the thermal load (change of the base load), so more energy can be consumed, therefore less storage mass is needed. The installation of the 24 collectors will be examined for thermal loads of 5 kW, 10 kW and 15 kW.

4.2.3 Results Solar field productivity The energy productivity of a solar field is defined as the energy generated by the solar field and is mainly affected by the number of collectors utilized, the mass flow of the heat carrier and targeted operation temperatures. Figure 4.8 presents the solar field productivity of the systems discussed in the previous paragraph compared to the heat demand of the building. The productivity of the system comprised by 4 collectors generates annually 8,354 kWh, while the solar field with the 24 collectors 35,621 kWh more (43,975 kWh). As can been seen, the energy produced by the 4 collectors-solar field exceeds the heat demand of the building in months June, July and August by 261 kWh, 621 kWh and 215 kWh, respectively. On the contrary, overproduction of energy is experienced by the 24 collector solar field from April to September, counting in total 24,340 kWh of excess heat. The monthly figures of both solar fields are provided in APPENDIX D2.

10000 Heat Demand [kWh] 4 collectors [kWh] 24 collectors [kWh] 9000 8000 7000 6000

5000 kWh 4000 3000 2000 1000 0

Figure 4.8 The productivity of the solar field consisting of 4 and 24 collectors compared to the energy demand for space heating

The solar factor, SFsf, k , that expresses the amount of energy generated by the solar field during a month k , Qsf, k , compared to the heat demand of the thermal load, Qdem, k , for that month is given by:

Qsf, k SFsf, k = (4.4) Qdem, k

The solar factor covered by the compared solar fields for each month is given in Table 4.2. Although the solar factor for both solar fields increases significantly in summer, in winter months the solar factors are low. In more detail, the solar field consisting of 24 collectors generates in July almost 10 times more heat than July’s heat demand, while in December the energy productivity reaches only 4% of the heat demand. Similarly, the 4 collector-solar field covers only the 1% of the heat demand in December, but the production of energy in July is double the heat demand (2.14). Furthermore, it is worth to highlight that the annual solar factor of the small and the big solar field is 0.16 and 0.86, respectively.

Month Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec Year 4 collectors 0.02 0.04 0.10 0.22 0.54 1.25 2.14 1.29 0.58 0.19 0.03 0.01 0.16 24 collectors 0.06 0.16 0.50 1.17 2.93 6.88 11.72 6.97 3.10 0.98 0.13 0.04 0.86 Table 4.2 Solar factor of the compared solar fields throughout the year

Energy delivery of the solar systems Although the solar field productivity gives valuable information on how much energy a solar field can generate, the end user is mainly interested on how much energy the solar thermal system can deliver, taking into account the heat losses occurred at the buffer tank and at the heat exchanger. This is referred as the energy delivery of the solar systems. As expected, the energy delivered to the end user will be lower compared to the productivity of each solar field.

Figure 4.9 shows the energy delivered by four different systems compared to the heating demand of the reference building. The solar system consisting of 4 collectors (orange color) and the system with the 24 collectors designed in three different capacities of thermal load: 5 kW, 10 kW and 15 kW, in green, purple and blue, respectively. The energy delivered by the small installation (4 collectors) is 7,625 kWh on an annual basis, while the energy delivered by each configuration of the big solar field is affected by the connected thermal load of each system configuration. The delivered energy is 32,865 kWh, 28,650 kWh and 20,520 kWh for the systems connected to a thermal load of 15 kW, 10 kW and 5 kW, respectively. The monthly figures of the compared solar systems are provided in APPENDIX D2. As becomes obvious, the variation of the thermal load influences the performance of the overall system.

10000 Heat Demand [kWh] 4 collectors, 5 kW [kWh] 24 collectors, 5 kW [kWh] 9000 24 collectors, 10 kW [kWh] 24 collectors, 15 kW [kWh] 8000

7000

6000

5000 kWh 4000

3000

2000

1000

Figure 4.9 Energy delivered by the investigated solar thermal systems

Furthermore, as illustrated, the months that experience the two extremes in terms of heat demand and energy production are the months January and July. In January, when there is a high demand for heating (8,996 kWh) the solar systems comprised by the 24 collectors cannot deliver energy to the system. On the other hand, in July all the four investigated solar systems are able to generate high amounts of energy when space heating is not a necessity. A better overview of the mismatch between the generated energy and the heat demand for space heating is given by the solar factors achieved by the solar systems.

The solar factor of a solar system, SFsys, k , as defined at the equation (4.19), is the energy delivered by the solar system in a month k , Qdel, k , divided by the heat demand of the thermal load, Qdem, k , for that month.

As can been seen in Table 4.3 the same size of solar field (24 collectors) operates more efficient when the thermal load is 15 kW. In this case the annual solar factor reaches 0.65, while the systems with a 10 kW and 5 kW achieve a yearly solar factor 0.56 and 0.40, respectively. The reason for this variation is accounted to the fact that the given size of the solar field (24 collectors) matches better to the load of 15 kW. In more detail, the thermal power delivered from the solar field to the buffer tank under the 2 reference irradiance of Erlangen (600 W/m ) is Qsf=31.7 kW. This is more than six time higher compared to a thermal load of 5 kW, while is almost the half of the 15 kW load. Moreover, by employing a higher power demand to the solar system (thermal load), less energy has to be stored in the buffer tank; thus the size of the latter becomes smaller. The smaller the buffer tank, the faster it is getting charged, thus more energy is consumed by the end user “instantly” and less energy is “trapped” in the thermal mass of the tank.

Month Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec Year 4 collectors 0.01 0.03 0.09 0.20 0.51 1.20 2.06 1.22 0.54 0.17 0.02 0.01 0.15 24 coll., 5 kW 0 0 0.04 0.59 1.54 3.48 6.83 4.27 1.87 0.35 0 0 0.40 24 coll., 10 kW 0 0 0.21 0.78 2.20 5.40 9.02 4.96 2.20 0.55 0 0 0.56 24 coll., 15 kW 0 0.02 0.32 0.90 2.41 5.86 9.84 5.60 2.44 0.68 0.03 0 0.65 Table 4.3 Solar fractions of the compared solar systems for each month

The main characteristics of the system configurations for the big solar field (24 collectors) are presented in Table 4.4. As can been seen, the increase of the capacity of the thermal load, decreases dramatically the volume of the buffer tank that is needed. Although the operation hours of the system configuration A (4,104 hours) are almost double compared to the configuration C (2,191 hours), the total energy delivered to the end user by configuration C (32,865 kWh) is nearly 30% more compared to system A (20,520 kWh). Moreover, the maximal temperature experienced in the buffer tank for systems B and C follows the rule of thumb that the maximal temperature of the tank should not be more than 25 to 30˚C, higher than the operation temperature of the thermal load (Tout,load=85°C).

system size of the max. overall energy thermal load 3 operation hours configuration tank (m ) temperature (˚C) delivered (kWh) A 5 kW 20 143 4,104 20,520 B 10 kW 10 112 2,865 28,650 C 15 kW 5 113 2,191 32,865 Table 4.4 Main parameters of the system configurations A, B and C

The variation of the temperature of the buffer tank throughout the year (8,760 hours) for the three system configurations is illustrated in Figure 4.10. The buffer tank of system A (20 m3) experiences higher temperatures compared to systems B (10 m3) and C (5 m3) in summer months June, July and August when the energy production is maximized. Since the delivered thermal load is only 5 kW, the generated energy that is not fed to the load is accumulated to the buffer tank and increases the temperature of the water inside the tank. As explained above, the thermal loads 10 kW and 15 kW match better to the investigated fixed solar field; thus for systems B and C the temperature of the tanks is limited to a lower range with maximal values of 112°C and 113°C, respectively. Furthermore, the graph shows a large fluctuation of the tank temperature regardless the system configuration in winter months January, February, November and December. The very low temperature inside the buffer tank is a result of the limited solar potential during these months and of the relatively large thermal mass of the tank compared to the generated energy. In other words, the energy that is generated by the solar field is accumulated to the buffer tank as normal, but due to the limited generation of heat the latter is not enough to charge the buffer tank; thus heat is “trapped” at the bulk of the buffer tank.

160.0 A B C 140.0

120.0

100.0

80.0

60.0

40.0 temperature of the tank [˚C ] [˚C tank the of temperature

20.0

0.0 0 720 1440 2160 2880 3600 4320 5040 5760 6480 7200 7920 8640 hours

Figure 4.10 Temperature variation of the buffer tank for the system configurations A, B and C

4.3 Solar cooling

4.3.1 Heating demand profile of the applied chillers For the investigation of solar cooling with the UHV collectors heat driven chillers (absorption machines) will be used. Since there is no documented information about the cooling profile of the institute, the latter will be determined by a number of assumptions mainly related to the solar system and to the needs of the institute. Moreover, the technique of free cooling will be taken into account in order to determine the cooling demand profile.

Solar system assumptions Due to the absence of a documented cooling demand profile and based on the constraint of the available roof space of the institute, the cooling load covered by the solar cooling system will be considered based on what the solar thermal installation can provide to, instead of the other way around. This is a way of determining the boundary conditions of the problem, when the available surface for the solar field is a fixed value.

Based on an empirical rule concerning the installations with the UHV collectors, in order a solar system with a given solar field to achieve high number of operation hours during the year, the thermal load connected to the solar system should be 2 to 3 times less than the power of the solar field under the reference irradiance.

At the present investigation the solar field has been selected to be installed in an area of 1,400 m2 (orange surface in Figure 4.3), which is equivalent to a solar field of 140 UHV collectors. Under the 2 reference irradiance of Erlangen (Iref=600 W/m ), the solar field is generating power of Qsf,ref=160 kW. Using the rule of thumb described above for a decreased factor 2, the approximation of the thermal

51 load would be 80 kW. It is highlighted that the 80 kW, is the power of the thermal load of the solar thermal installation and not the delivered cooling capacity of the solar cooling system.

The cooling capacity of the total installation is influenced by the type of the absorption chiller used. The performance of a refrigeration cycle (cooling system) is given by the coefficient of performance (COP) and is expressed as the cooling output ( Qcool ) divided by the required energy input in order to achieve the cooling output ( Qdel ):

Q COP =cool (4.5) Qdel

For instance, a 1-step absorption chiller with a maximal coefficient of performance COP=0.7 can deliver

56 kWpeak of cooling load, while a 2-step absorption chiller with a maximal COP=1.2 can provide a cooling load of 96 kWpeak. As mentioned in sub-chapter 4.1, the centralized cooling installation of the institute is characterized by a high consumption of cooling that is provided by compressor chillers with a peak cooling capacity approximately 1.3 MW (RTHA 450 and RTHA 215). Based on this information, the under investigation solar cooling systems with a 1-step and a 2-step absorption chiller, will be able to cover the 4.3% and 7.3% of the peak cooling capacity of the institute, respectively.

Free cooling A technique that is used quite often for cooling in regions of Central Europe is the use of an economizer cycle, or free cooling. Free cooling is an approach to lowering the air temperature in a building by using naturally cool air or water instead of mechanical refrigeration. The atmospheric air at many latitudes and elevations can be considerably cooler during certain seasons and times of the day than the air that is warmed inside a building. The cooler outdoor air can be introduced directly into the cooling system of a building, in order to help cooling the building interior. Many air conditioning systems operate with a fixed minimum amount of outdoor air. The mechanical refrigeration load on these systems can be reduced by modifying the system to utilize outdoor air–up to 100% of its supply airflow–when outdoor air is cooler than return air. The sensors attached to the HVAC system activate the economizer cycle when the temperature is low enough to be able to assist in cooling the interior of the building. This allows the HVAC system to operate less and use less energy, as well. In actual practice, free cooling is not entirely free, because pumps, fans, auxiliary equipment and periodic maintenance are needed.

In the current investigation is assumed that the economizer cycle operates between temperatures Tcut,in

=-10°C and Tcut,out =10°C . When the temperature of atmospheric air is within this range, Tcut,in

Tcut,out , the cooling system is turned off and replaced by the economizer cycle. Due to the low temperatures experienced in Erlangen throughout the year, this strategy has a large impact on the cooling demand pattern. Based on the temperatures Tcut,in ,Tcut,out and the variation of Tamb, the economizer cycle is possible to operate for 4,863 hours in a year cycle.

As discussed previously, the type of the absorption chiller employed in a solar cooling system influences the COP of the cooling installation. Since in the present investigation a comparison between a 1-step and a 2-step absorption chiller will be performed, the demand profile of the system is expressed as a

52 thermal energy demand, instead of presenting a cooling demand profile. The thermal energy demand profile of the solar cooling system is calculated by using the solar system assumptions and the free cooling technique and is presented in Figure 4.11, while the monthly figures are presented in APPENDIX D3. The annual heat demand is 311,760 kWh and the highest monthly quantities are present from May to September. In more detail, 78% of the annual heating demand of the solar cooling system (243,200 kWh), is required during the months that the solar potential is favorable for heat production (May to September).The above fact shows the match that takes place between the availability of solar energy and the demand for cooling.

60,000 Energy Demand [kWh] 50,000

40,000

30,000 kWh

20,000

10,000

Figure 4.11 The heating demand of an absorption chiller, sized according to the applied solar field and the free cooling technique

4.3.2 Solar cooling with absorption chillers The technology that is employed in solar cooling is a 1-step and a 2-step type of absorption chillers. An absorption chiller is thermodynamically similar to a conventional vapor compression chiller, with the difference that the refrigeration cycle instead of using a compressor, utilizes a complex group of devices, in order the working fluid of the cycle to upgrade from a low pressure to a high pressure state. The driving force of an absorption chiller is thermal energy, thus it is a heat driven machine, rather than an electrically-driven compressor chiller. Depending on the type of the absorption chiller used, the coefficient of performance varies in a range of 0.6 to 1.2, while the conventional compressor chillers operate with an average COP=3.0. As becomes clear the first ones are less efficient compared to the compressor chillers; thus absorption machines rely upon the energy input that is relatively cheap, like solar energy or waste heat.

The key technical disadvantage with absorption chillers is that they require quite high temperatures to drive them, typically upwards of about 80°C (the higher the temperature, the higher the efficiency). For a conventional solar-thermal system with flat plate or vacuum tube collectors, it can be very difficult to maintain such high-grade heat during the year. Result of that is solar cooling in Central Europe mainly to

53 be implemented with technologies that operate on lower temperatures like adsorption chillers or desiccant cooling (DEC) systems (40-50°C) [10].

One-step absorption chiller For a solar system connected with a 1-step absorption chiller, it is assumed an inlet and outlet temperature of the absorption machine Tin,load=90°C and Tout,loadr= 85°C (Figure 4.12). As described above, the thermal load that the available solar field can deliver has been considered 80 kW, which is equivalent to a cooling capacity of 56kW assuming the chiller to operate on a COPmax=0.7. According to the model used for the investigations with the solar thermal UHV collectors in sub-chapter 3.3, the solar  field has been designed based on a Tin,field=86.5°C and Tout,field=100°C generating Qsf, ref =188.6 kW , 2 under the reference irradiance Iref=600 W/m . The solar field consists of 144 collectors, connected in 12 strings of 12 collectors each. The maximum thermal power generated by the designed solar field, under 2  3 Imax=950 W/m , is Qsf ,max =312.2 kW . The buffer tank used for the operation of the system is a 30 m tank, while the heat transport fluid (HTF) of the solar system is water.

Figure 4.12 Solar cooling with a 1-step absorption chiller

Two-step absorption chiller On the other hand, the system with the 2-step absorption chiller operates on a higher temperature with result to deliver higher cooling capacity (Figure 4.13). The solar system delivers energy with a pace of 80 kW that can be transformed in cooling capacity of 96 kW (assuming a COPmax=1.2). Due to the high temperatures needed in order a 2-step absorption chiller to operate, water cannot considered an option as the heat transport fluid (HTF), therefore thermal oil is utilized. The overall solar system has been designed to deliver 80 kW of thermal power to the absorption machine on a operation temperature

Tin,load=175°C and Tout,load=160°C. The solar field has been designed on a Tin,field=162.5°C and  Tout,field=185°C and is possible to generate Qsf, ref =163.6 kW , under the reference irradiance Iref=600 W/m2. The solar field consists of 140 collectors, connected in 10 strings of 14 collectors each. The 2 maximum thermal power generated by the designed solar field under Imax=950 W/m is  3 Qsf ,max =312.8 kW , while the volume of the integrated buffer tank used is 110 m .

Figure 4.13 Solar cooling with a 2-step absorption chiller

4.3.3 Results One-step absorption chiller Figure 4.14 illustrates the thermal energy demand of the cooling system (grey color) compared to the solar energy delivered by the solar system (orange color) and the maximum cooling capacity that can be produced by the 1-step chiller (green color). The annual values of the quantities are 311,760 kWh, 210,160 kWh and 147,112 kWh, respectively. As discussed in paragraph 4.3.1, 78% of the annual thermal energy demand of the cooling system occurs in the period May to September. Similarly, 73% of the solar energy delivered throughout the year (153,360 kWh) is available in the same period, offering a great match between demand and supply of solar energy. The monthly figures of the studied solar cooling system are given in APPENDIX D4.

60000 Thermal Energy Demand [kWh] Solar Energy Delivered [kWh] Cooling Produced [kWh]

50000

40000

30000 kWh

20000

10000

Figure 4.14 Solar energy delivered to the 1-step absorption chiller throughout the year

The solar factor, as defined in the equation 4.4, for each month is given in Table 4.5. The annual solar factor reaches 0.67, with a maximum of 1.95 and a minimum of 0.35 in March and November, respectively. In more detail, in March the demand for cooling (expressed as thermal energy demand of the chiller), is limited due to the operation of the free cooling sub-system; thus the available solar potential in March is enough to cover the needs for cooling. On the contrary, the solar irradiation in

November is not enough to provide the chiller with the required thermal energy. Furthermore, in July, the month with the highest demand for thermal energy (56,160 kWh), the solar system is able to cover 77% of the required energy.

Month Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec Year Solar Factor 0 0.44 1.95 1.06 0.80 0.77 0.59 0.48 0.54 0.66 0.35 0 0.67 Table 4.5 Solar factor of the solar cooling system (1 step absorption chiller)

The designed solar system is able to deliver 80 kWth to the connected absorption chiller for 2,627 hours, annually. The distribution of the operation hours throughout the year is presented in Figure 4.15. As expected, during the periods with poor solar potential, January to February and November to December the solar system is not able to deliver heat to the chiller; thus heat should be supplied to the absorption machine by an auxiliary energy source. On the contrary, the favorable solar irradiation in June allows the solar system to deliver thermal power for 22 hours per day (days: 170th to 173th).

24 22 operation hours per day 20 18

16 14 12 10

hours per day per hours 8 6 4 2 0

1 10 19 28 37 46 55 64 73 82 91 100 109 118 127 136 145 154 163 172 181 190 199 208 217 226 235 244 253 262 271 280 289 298 307 316 325 334 343 352 361 day of the year Figure 4.15 Operation hours of the solar thermal installation connected to a 1-step absorption chiller

Two-step absorption chiller Similarly, the solar energy delivered and the maximum cooling that can be generated by the solar system connected to a 2-step absorption machine is illustrated in Figure 4.16. The annual energy delivered to the chiller is 117,040 kWh, while the cooling that can be extracted by the chiller is 140,448 kWh. Compared to the values of the 1-step absorption chiller, the energy delivered by the solar system is 93,210 kWh less, but the maximal cooling delivered is decreased only by 6,664 kWh. The declined energy delivered by the solar system, sources from the fact that the given solar field operates in a much higher temperature (design temperature: Tin,field=162.5°C and Tout,field=185°C ), compared to the operation temperature of the system equipped with a 1-step chiller (design temperature: Tin,field=86.5°C and Tout,field=100°C). On the other hand, the 2-step chiller is possible to perform on a COPmax=1.2 compared to the COPmax=0.7 that is being considered for a 1-step chiller; thus the decreased energy delivery of the solar system is compensated by the higher COP of the 2-step absorption machine. The monthly figures of the studied solar cooling system are provided in APPENDIX D4.

60000 Thermal Energy Demand [kWh] Solar Energy Delivered [kWh] Cooling Produced [kWh]

50000

40000

30000 kWh

20000

10000

Figure 4.16 Solar energy delivered to the 2-step absorption chiller throughout the year

The solar factor of the solar system is presented in Table 4.6. The annual factor of the solar system connected to the 2-step absorption machine is 0.38, which is 43% less than the solar factor of the system with the 1-step chiller (0.67). The highest solar factor occurs is April (0.60), while the lowest value is experienced in November (0.03). Furthermore, in months January, February and December is not possible for the solar system to deliver thermal energy to the 2-step absorption chiller.

Month Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec Year Solar Factor 0 0 0.30 0.60 0.48 0.50 0.37 0.28 0.32 0.32 0.03 0 0.38 Table 4.6 Solar factor of the solar cooling system (2 step absorption chiller)

The solar system is able to deliver thermal energy of 80 kWth for 1,468 hours annually, with a distribution shown in Figure 4.17. Compared to the solar system connected to the 1-step chiller (2,627 hours), the total number of operation hours presents a decline by 1,159 hours (or 44%). Moreover, the maximum hours of operation that can be realized during a day (24 hours) is 16 hours, against the maximum value (22 hours per day) that can be achieved by the solar system connected to the 1-step chiller. The large variation between the two systems sources from the fact that, although the two systems utilize almost the same number of collectors 140 (system with a 1-step chiller) and 144 (for the system with the 2-step chiller), the sizes of the buffer tanks applied to the systems are 30 m3 and 110 m3, respectively. According to the model used to perform the simulation of the buffer tank (paragraph 3.3.2), the generated energy from the solar field is maintained at the thermal mass of the buffer tank, until the buffer tank becomes fully charged (fully mixed buffer tank). Therefore the system has to overcome the thermal inertia of the buffer tank, in order the solar system to deliver energy to the thermal load (absorption chiller).

24 operation hours per day 22 20 18

16 14 12 10 hours hours day per 8 6 4 2 0

1 10 19 28 37 46 55 64 73 82 91 100 109 118 127 136 145 154 163 172 181 190 199 208 217 226 235 244 253 262 271 280 289 298 307 316 325 334 343 352 361 day of the year Figure 4.17 Operation hours of the solar thermal installation connected to a 2-step absorption chiller

Chapter 5. Small scale steam cycle powered by solar thermal UHV collectors

5.1 Power generation with a small scale steam cycle Exploitation of the solar thermal UHV collector for power generation will be realized in a small scale Rankine cycle. For the operation of a small scale steam cycle, suitable sized turbines should be selected. Such small turbines are commercially available from the company GREEN TURBINETM. The specific manufacturer offers turbines in two variations: 1.2 kW and 15 kW of shaft power. The first one is expected to be integrated in a cycle with a boiler of 13 kWth, while the second one with a boiler of 110 kWth. Since the system is oriented to be installed on the roof of an office building, the availability of surface area for the installation of the solar thermal UHV collectors plays important role for the operation of the power cycle; thus for the selection of the cycle’s turbine. Based on the constraint that on the roof of the given building in Erlangen (240 m2) up to 24 collectors can be integrated to the studied system, the turbine which will be realized to the power cycle is the one with the 1.2 kW shaft power.

5.1.1 Description of a Rankine cycle The steam cycle which will be studied is a small scale Rankine cycle (Figure 5.1) oriented to operate continuously throughout the year (24 hours per day, 7 days per week). The Rankine cycle is equipped with a steam turbine of 1.2kW shaft power, and suggestions concerning the main components of the cycle are given by the steam turbine manufacturer (GREEN TURBINETM). The mass flow of the cycle is determined to 18 kg/h (5x10-3 kg/s), while the feed pump of the cycle delivers the cycle’s working fluid to the heat addition process (boiler) at 45˚C and 5.5 bar abs. The boiler should produce superheated, dry steam that enters to the turbine under 180˚C and 5 bar absolute pressure. The working fluid is expanded in the turbine (1.2kW shaft power) which is connected to a generator and then is directed to the condenser of the cycle. The condenser rejects part of the heat content of the working fluid and drops the temperature of the latter to 45˚C. After the condenser, the working fluid is directed to the feed pump and the cycle is completed.

Based on the assumptions that the steam turbine is able to deliver Pturb=1.2 kW of shaft power, the efficiency of the electro-generator is ηgen= 0.9 and that the parasitic electricity consumption of the cycle is Ppar=0.08 kW, the net electricity production of the cycle, Pnet,cycle , is calculated:

Pnet, cycle= P turbη gen − P par =1 kWe ( 5.1)

As a result, since the cycle is going to operate continuously throughout the year (8,760 hours), the net production of electricity the cycle is going to be:

EPannual= net, cycle ⋅8,760 hours = 8,760 kWhe ( 5.2)

Considering the cost of kWhe under an industrial tariff of €0.143/kWhe, the annual savings achieved by the cycle’s power generation will be approximately €1,250.

Figure 5.1 A Rankine cycle

5.1.2 Heating demand of a Rankine cyle During the boiling process of a steam cycle, compressed liquid water enters to the boiler and exits as superheated steam. The boiling process of the working fluid consists of three steps: a) heat addition before boiling, b) boiling of the water and c) heat addition to the superheated state. The overall thermal energy needed to realize the heat addition to the cycle is considered:

Q= Q + Q + Q = mCp ∆Τ +m ∆ h + mCp ∆Τ 5.3 dem econ evap sup f, H22 O econ f→ g g, H O sup ( )

Since in the studied Rankine cycle the states of the working fluid before and after the heating process are known, the heat demand of the cycle can be determined. For the rest of the chapter, is assumed there is no pressure drop of the working fluid as it passes through the heat exchanger. From thermodynamic tables, is obtained that:

• for water the saturated temperature under 5.5 bar is Tsat= 155˚C

• the latent heat of water is Δhfg=2,096 kJ/kg

• the heat capacity of water assumed Cpf=4.2 kJ/kg/K and Cpg=2 KJ/kg/K

The thermal power needed to be provided to the working fluid of the cycle can then be calculated as  Qdem =13 kW .

Taking into account that the energy demand of the cycle is Qdem=13 kWth and the net electricity production Pnet,cycle=1 kWhe, the efficiency of the cycle, ηnet,cycle can be determined:

Pnet, cycle ηnet, cycle = =7.6% ( 5.4) Qdem

5.1.3 Variations of the heating process in a Rankine cycle The integration of a solar thermal installation in the heating process of a power cycle can be realized by different ways and is affected by several variables. One of them is the type of the solar thermal technology used, since it influences the maximal temperature the working fluid can obtain. For instance the use of trough mirrors could lead to direct steam generation, while the installation of flat plate collectors could preheat the water, before it enters to the evaporator (substitution of the economizer). Another important variable is the point of the cycle at which the ejection of solar energy will take place (preheating or main heating process). Furthermore, is very important the designer to identify if the heat will be delivered to the working fluid of the cycle, or will be realized through an intermediate circuit, as applied in Organic Rankine Cycles (ORC).

An investigation on the integration possibilities of the solar thermal UHV collectors has been realized and the available options are presented in Figure 5.2. Starting point of the investigation is to determine if the heating process of the steam cycle is taking place directly to the cycle’s working fluid (through a fired boiler), or by using a suitable heat exchanger (unfired boiler). In the first option heat is added directly to the working fluid of the cycle (water), while in the latter an intermediate heat transport fluid (HTF) is being used to transfer the heat to the working fluid of the cycle. With this configuration, flexibility is offered to the integration of solar thermal system, since an intermediate HTF is used. Discussion related to the heat addition to the working fluid of the cycle (water) will be conducted in Chapter 6.

Figure 5.2 Possible ways for heat addition in a steam cycle

Given the daily and annually variation of solar irradiation, the heat source of the power cycle should include an auxiliary heat source which will be able to compensate the lack of generated heat by the solar thermal installation. The integration of the solar thermal system to the cycle can be realized either via a connection in parallel, or in series with to auxiliary heat source. In the first option (connection in parallel), the solar thermal system will operate in a relatively high temperature (e.g. 210-230˚C), fact that decreases the efficiency of the solar thermal UHV collectors and limits the operation hours of the

61 solar installation. On the contrary, by connecting the solar system in series to the auxiliary heat source, the operation temperature of the former can be adjusted in the desired temperature range of the power cycle; thus more freedom is given to the design of the system.

In the present study, focus will be given to the solutions provided by an unfired boiler employing thermal oil as a HTF (intermediate circuit). Moreover, the connection of the solar system to the unfired boiler will be realized in series.

5.1.4 Heat demand of a Rankine cycle with an unfired boiler Heat addition to the power cycle through an unfired boiler will be employed with a heat exchanger device. The heat source is not included in the primary components of the cycle, therefore in order the heat to be transferred from the heat source(s) to the heat exchanger, an intermediate heat transport fluid (HTF) is needed (Figure 5.3).

Figure 5.3 Heat addition to the cycle through a heat exchanger (unfired boiler)

Although, in big scale power plants these steps are taking place in different compartments of the boiler (economizer, evaporator and superheater, respectively), for the studied steam cycle a shell and tube heat exchanger, that includes the three stages mentioned above, will be employed. Assuming the effectiveness coefficient of the heat exchanger ε=0.8, the thermal power should be delivered to the heat exchanger is calculated:

Q Q =dem =16.25 kW ( 5.5) oil ε

The heat addition of the cycle will be realized through a two phase shell and tube heat exchanger. At the hot stream of the heat exchanger (shell) thermal oil will circulate, in order to provide the necessary heat to the steam cycle, while the cycle’s working fluid will flow through the tubes of the heat exchanger. The energy source(s) should be designed in such a way that the thermal oil loop will deliver thermal power of 13 kW to the heat transport fluid of the cycle (water). Since there the operation temperatures of the heat exchanger are not determined, is assumed that the thermal oil will operate in Tin,boiler=200˚C and

Tout,boiler=80˚C. The thermal oil that will be used is a widely used heat medium for industrial processes, called Therminol 66.

Given the average operation temperature Tav= 140˚C and the heat capacity of the selected thermal oil,

Cpoil=1.98 kJ/kg/k , the mass flow of the thermal oil loop can be determined with:

 Qoil= m oil Cp oil ∆Τ boiler (5.6)

 Resulting in a value of m oil = 0.07 kg / s and in a power demand of the overall process, Qoil =16.6 kW .

In the given study the auxiliary heat source will be a natural gas burner connected in series to the solar thermal installation. In this way the natural gas burner will provide the necessary heat to thermal oil, when the solar system is not able to operate. Solar thermal energy will be ejected at the thermal oil circuit (intermediate circuit), before enters to the burner. The process, at which the integration point is taking place, could be determined as a medium temperature preheating process of the thermal oil at the return stream of the heat source.

5.2 Powering a Rankine cycle with solar thermal UHV collectors

5.2.1 Connection of the solar system to the natural gas burner In most of the cases, storage of the generated heat in a solar thermal system is a necessity, in order the system to be less sensitive to the solar irradiance variation. In the present investigation the utilization (or not) of a storage tank will be taken into account. A solar installation without buffer tank (a solar field coupled to the intermediate circuit) is less complex, but is limited to deliver the generated energy instantly. The option of integrating only a solar field to the heating process of the cycle can be employed either through a heat exchanger (existence of temperature gradient between the solar loop and the intermediate loop), or via the direct circulation of thermal oil from the solar collectors to the heat exchanger of the cycle. In the first option heat losses occur at the heat exchanger interface of the two streams, while in the second option suitable match to the intermediate circuit should be realized. On the other hand, by using a storage tank, the appropriate match between demand and supply of solar thermal energy can be accomplished and at the same time flexibility at the size of the solar field is given.

The layouts that are going to be studied for the integration of the solar thermal UHV collectors are: a) the direct coupling of the solar field to the burner, and b) the integration of the solar system with a storage tank.

The system configurations described above will be presented for each case in the following sections. The physical constraints that are taken into account for the realization of the study are: i) the climate data, considering the installation location at Erlangen in South Germany, and

63 ii) the available surface that can be exploited for the installation of the solar field (240 m2), which is equivalent to 24 solar thermal UHV collectors.

5.2.2 Direct coupling of the solar field System_Description The proposed system configuration for direct coupling of a solar field to the power cycle is given in Figure 5.4. The heating process consists of the solar thermal subsystem (solar field) and the main heat source sub-system (burner), connected each other in series. The connection between the two sub- systems is being realized through a 4-way valve that is able to isolate the thermal oil circuit in two parts, the solar loop and the process loop. The normal operation of the overall installation is taking place when the burner heats up the thermal oil that circulates at the shell side of the heat exchanger. During this mode, the 4-way valve does not allow the two loops to be mixed. When the available solar irradiation is able to heat up the solar loop, the 4-way valve changes position in such a way that the thermal oil circuit is heated up by the solar field. After that, the thermal oil enters to the burner to receive the remaining energy needed before enters to heat exchanger (unfired boiler). The condition that has to be fulfilled, in order the 4-way valve to allow the mixture of the two loops is the outlet temperature of the solar field,

Tsf,out , to be higher than the return temperature of the process loop, Tpl,ret .

Figure 5.4 Direct coupling the solar field with the auxiliary burner of the cycle through a 4-way valve

Since no buffer will be installed at the solar thermal installation, all the generated heat from the solar field has to be consumed instantly by the thermal loads of the cycle (thermal oil loop of the cycle). Therefore the design of the solar field is limited to deliver its maximal thermal power, under the maximal irradiance of the location that is applied. Moreover, the way the 4-way valve is used on this system configuration, constraints the mass flow of the solar loop to be the same with the one of the process loop:

msl= m pl = m oil =0.07 kg / s ( 5.7)

It should be mentioned that the operation of the valve could create instant pressure peaks in the two loops as it switches from the one mode to the other, fact that could limit the lifespan of the component (4-way valve).

Modeling As discussed in paragraph 5.1.4, the feed and the return temperature of the thermal oil circuit have been determined as Tpl,feed=200˚C and Tpl,ret=80˚C, respectively. The solar field should operate with a mass flow of 0.07 kg/s and its outlet temperature under the maximal solar irradiation should not exceed the temperature Tpl,feed=200˚C. Given the latter constraint (mass flow and operation temperature), the designed solar field will consist of seven solar thermal UHV collectors, connected each other in series.

Due to the limitations of the model used for the design on the solar field, the generated power and the temperature is not possible to be expressed as function of the instant irradiance of the location. For this reason the global irradiance data has been grouped in eleven ranges covering the local solar potential throughout a calendar year. By calculating the thermal power generation of the solar field for the average value of each of the eleven ranges, an estimation of the generated energy throughout the year can be extracted.

The eleven ranges of the solar potential in Erlangen are presented in Figure 5.5. The irradiance levels assumed to remain constant on an hourly basis and the global irradiance is expressed in W/m2. From the presented chart can be seen that at the given location (Erlangen, South Germany) there are 4,199 hours during the year (in total 8,760 hours) that there is no solar irradiance available (0 W/m2), therefore the solar thermal system cannot generate any energy. The two ranges with the lowest global irradiance are the [0-100] and [100-200] W/m2, which are applicable for 1,886 and 722 hours. On the other hand, the ranges with the highest irradiance levels [800-900] and [900-1,000] W/m2 appear for 179 and 93 hours, respectively.

179 93 192 209 229 0 288 0,-100 343 100-200 420 200-300 300-400 4.199 400-500 722 500-600 600-700 700-800 800-900 1.886 900-1000

Figure 5.5 The solar potential in hours of irradiance throughout a calendar year in Erlangen, Germany

The unknown variables needed in order to determine the performance of the solar field, are the temperature output (Tsf,out) and the thermal power generated from the solar field (Qsf,out), while the inlet temperature of the thermal oil is fixed in Tsf,i n=80˚C. The unknown parameters are calculated in function of the solar irradiance and presented in the average values of the ranges discussed earlier (Table 5.1).

average value of the Qsf,out Tsf,out irradiance range [W/m2] [kW] [˚C] 50 -0.1 79 150 1.9 95 250 3.9 110 350 5.9 126 450 7.9 140 550 9.8 154 650 11.8 168 750 13.7 182 850 15.6 195 950 17.5 207

Table 5.1 The power (Qsf,out) and the temperature (Tsf,out) output of the solar field at the average values of the eleven irradiance ranges

As expected, by keeping the mass flow of the solar field constant, the generated heat (Qsf,out) and the temperature (Tsf,out) of the thermal oil at the exit of the solar field increases as the solar irradiance becomes higher. Under irradiance of 50 W/m2, the solar field is not possible to add heat to the thermal oil that circulates inside the collectors; thus the operation of the solar field will act beneficially to the heating process, for irradiance higher than 50 W/m2. On the contrary, under 950 W/m2, the solar field is able to deliver 17.5 kWth and the temperature that the thermal oil can obtain is Tsf,out=207˚C.

By correlating the power output of the solar field (Qsf,out) from Table 5.1 to the hours that each irradiance range is available throughout the year in Erlangen (Figure 5.5), an estimation of the generated heat by the solar field can be extracted. The results of this correlation are presented in paragraph 5.4.1.

Results The energy generated by the seven solar thermal UHV collectors on a monthly basis is presented in Figure 5.6. In grey color is shown the monthly energy required to run the Rankine cycle continuously, while in orange color the energy generated by the solar installation is illustrated. The fluctuation of the energy demand per month throughout the year is result of the different number of days included in each month. For example, in February (28 days) the energy demand is 1,196 kWh less than the energy demand of January (31 days); thus there is a fluctuation of the energy demand before and after February, as illustrated in the graph. The solar field supports the heating process mainly the period from April to September, when more than 2,000 kWh generated per month during this period. In contrast, the heat generation is limited the rest months, with the lowest values to be experienced in January (250 kWh) and December (152 kWh).

14000 Energy Demand [kWh] Solar Energy Produced [kWh] 12000

10000

8000

kWh 6000

4000

2000

Figure 5.6 Generated energy by the solar field coupled to the auxiliary burner of the heating process

Despite the total energy demand of the heating process is 145,504 kWh, the annual generation of heat by the solar installation is around 18,892 kWh. As a result, the rest energy needed -almost 126,614 kWh- will be provided by the natural gas burner. Since most of the demanded energy is generated by the natural gas burner (almost 87%), becomes understandable that the solar field plays a supplementary role at the heating process.

Since the solar system operates as a direct feed energy source, without having the ability to store energy, the generation of heat follows the pattern of the solar irradiation throughout the year. The solar fraction of the replaced energy by the solar system has been defined in Chapter 3 by the equations (3.18) and (3.19). Since in the given design the solar system consists only from a solar field:

Qsf, k SFsys,, k= SF sf k = ⋅100% ( 5.8) Qdem, k

The overall solar fraction of the energy generated reaches 13% of the total energy demand (Table 5.2). Most of the energy is generated during the summer months and especially in June with coverage that reaches 25.5% of the energy demand of the month. On the other hand, the months with the lowest solar fraction are January, February, November and December, with 2%, 4.9%, 2.9% and 1.2%, respectively.

Month Energy Demand [kWh] Solar Field Productivity [kWh] Solar Fraction [%] January 12,358 250 2 February 11,162 550 4.9 March 12,358 1,411 11.4 April 11,959 2,043 17.1 May 12,358 2,716 22 June 11,959 3,054 25.5 July 12,358 2,692 21.8 August 12,358 2,134 17.3 September 11,959 2,014 16.8 October 12,358 1,531 12.4 November 11,959 345 2.9 December 12,358 152 1.2 Year 145,504 18,892 13 Table 5.2 Solar fraction of the system consisting of 7 UHV collectors

Assuming the energy source of the burner is natural gas, the equivalent energy savings can be calculated. Taking into account the heating value of natural gas 9.72 kWh/m3 and that the efficiency of the boiler is 0.8 , the natural gas that can be saved annually, by using the solar field with the 7 solar thermal UHV collectors, is 2,429 m3. Considering the cost of natural gas €0.048/kWh (industrial tariff, in Germany), the equivalent annual savings are €1,134 /year.

5.2.3 Connection through a buffer tank System Description In most of the solar thermal installations the use of a storage tank is a necessity. By adding a buffer tank at the solar thermal installation, the thermal power consumer and generator are decoupled via an intermediate storage tank. The tank is able to store all the generated heat from the solar field and deliver it to the secondary loop, under fixed temperature difference at the heat exchanger. This fact allows the solar field to generate power higher than the power demand of the process (>16.6 kW), depending on the size and the design of the solar field. The excess generated heat is stored at the buffer tank, in order to be used in a later time. In other words, by using a buffer tank in a given solar field a higher amount of operation hours can be achieved by the solar thermal installation. The integration of the buffer tank can be utilized as presented in Figure 5.7. In the given configuration, the energy generated by the solar thermal UHV collectors is stored at the buffer tank and then through a heat exchanger is added to the return stream of the process loop, before enters to the burner.

Figure 5.7 Solar system with an integrated buffer tank connected in series to the burner of the cycle’s heating process

Modeling The following design of the solar system will be realized according to the model, described in sub- chapter 3.3. Starting point of the design of the solar system (Figure 5.7) is to set the temperatures of the hot and the cold side of the heat exchanger. This is realized by considering an effectiveness value of the heat exchanger and by setting the inlet and outlet temperatures of the hot (solar loop) and the cold side (process loop). From paragraph 5.1.4 can be extracted that the inlet temperature of the cold side is

Tpl,ret=80˚C, while the outlet temperature (Tpl,bur) should be equal or less than 200˚C. On the other hand, as discussed in paragraph 3.3.2, the temperatures of the hot side of the heat exchanger (solar loop) can be determined based on the rule of thumb:

Τsl,, feed =T pl bur +10 ( 5.9)

, and by using the effectiveness-NTU method, assuming an effectiveness value ε=0.9:

Τsl,, ret = Τ sl feed −ε ( Τ sl , feed −T pl , ret ) (5.10)

From the above becomes clear that the only value needed in order to determine the four sides of the heat exchanger is the Tpl,bur. This temperature will be determined indirectly by the solar field configuration. In this system configuration, the available surface area constraints the solar field to 24 solar thermal UHV collectors. Given the fact that the target of the solar system is to demonstrate a high operation temperature, the limited size of the solar field (24 collectors) and after several trials with different Tpl,bur; has been chosen that the solar field will consist of 2 strings comprised by 12 collectors each. The solar field configuration presents an acceptable pressure drop of 0.85 atm, and the inlet and outlet temperature of the solar field under the design irradiance (600 W/m2) are 160˚C and 88˚C, respectively.

Considering the model used for the description of the heat exchanger (paragraph 3.3.2), the inlet and the outlet temperatures of the hot side of the heat exchanger are determined as Tsl,feed=160˚C and Tsl,ret

=88˚C, respectively. The solar system is able to deliver 9.2 kWth, and to raise the temperature of the

69 process loop from 80˚C (Tpl,ret) to 150˚C (Tpl,bur), (Figure 5.7). Based on the design guideline that the maximal temperature of the buffer tank cannot be more than 30˚C higher than the outlet temperature of the cold side of the heat exchanger (Tpl,burn) :

Τbuf ,max =Tpl , bur +30 ( 5.11)

3 , the volume of the buffer tank is selected to be 20 m with a maximal temperature Tbuf,max=180˚C.

Since the solar thermal installation delivers 9.2 kWth to the process loop, the use of a auxiliary burner is considered mandatory, in order the process loop to reach the desirable temperature, Tpl,feed=200˚C.

From the analysis of the heat demand of the cycle, thermal power of 16.6 kWth is needed, which means the burner should provide 7.4 kWth. It is worth mentioning that the capacity of the burner should be ranged between 7.4 kWth and 16.6 kWth, so the power cycle will operate, regardless of the operation of the solar thermal installation.

Results The energy delivered from the solar thermal system to the process loop is given in Figure 5.8. The solar system consists of 24 solar thermal collectors and an integrated buffer tank of 20 m3. As expected, the energy demand of the process (heat addition to the cycle) is the same to the results presented in paragraph 5.2.2 (145,504 kWh). Since there is an integrated buffer tank at the solar system, the productivity of the latter is distinguished to the solar energy produced (in orange color) and to the solar energy delivered (in purple color). The first quantity represents the energy generated by the solar field and ejected to the buffer tank, while the second one shows the amount of heat delivered by the solar system to the process loop through the heat exchanger.

14000 Energy Demand [kWh] Solar Energy Produced [kWh] Solar Energy Delivered [kWh] 12000

10000

8000

kWh 6000

4000

2000

Figure 5.8 Delivered energy by the solar system with an integrated buffer tank, connected in series to the auxiliary burner of the heating process

Table 5.3 shows the monthly values of the quantities presented in Figure 5.8 and the monthly solar fractions throughout the year. The solar fraction is calculated as the solar system’s delivered energy to the energy demand of the process, according to the definition given in Chapter 3 (equation 3.19):

Qdel, k SFsys, k = ⋅100% ( 5.12) Qdem, k

The solar fraction experiences the highest value in June (44.6%), followed by the months July and May with 39.5% and 36.5%, respectively. It is highlighted that in months January, February and December, there is no energy delivered by the solar system; thus the solar fraction is 0 for these months. Given the annual delivered energy of the solar system is 28,636 kWh, the annual solar fraction reaches 19.5% of the energy demand of the process (145,504 kWh).

Energy Demand Solar Field Energy Delivered Solar Fraction Month [kWh] Productivity [kWh] [kWh] [%] January 12,358 251 0 0 February 11,162 791 0 0 March 12,358 2,378 1,509 12.2 April 11,959 3,895 3,175 26.5 May 12,358 5,263 4,517 36.5 June 11,959 6,135 5,331 44.6 July 12,358 5,423 4,887 39.5 August 12,358 4,287 3,563 28.8 September 11,959 3,926 3,332 27.9 October 12,358 2,814 2,102 17 November 11,959 462 148 1.2 December 12,358 174 0 0 Year 145,504 35,798 28,636 19.5 Table 5.3 Solar fractions of the system consisting of 24 UHV collectors and a 15 m3 buffer tank change as well

The equivalent savings of the solar thermal system is 3,682 m3 of natural gas or €1,718 , annually. In paragraph 5.2.2 has been shown that the solar field connected directly to the process loop achieves annual natural gas savings of 2,429 m3, equivalent to €1,134. Compared to the solution with the 7 collectors, the present system (24 collectors) saves annually 1,253 m3 of natural more (or €584). On the contrary, the latter option utilizes 17 collectors more (2.4 times larger solar field) and a 20 m3 buffer tank filled with thermal oil, facts that increase the initial investment of the installation dramatically.

As discussed before, the solar thermal installation has been designed in order to raise the thermal oil of the process loop with a mass flow of 0.07 kg/s, from 80˚C to 150˚C. The operation hours of the installation throughout the year are 2,993. Taken into account that the overall hours per year are 8.760, the annual utilization factor of the installation is 0.34 (2,993/8,760). Figure 5.9 shows the operation hours of the solar installation per day throughout a calendar year. During summer the available irradiation allows the solar thermal installation to deliver energy to the process loop on a 24 hour basis during a day. On the other hand, during the first 60 days of the year (January and February) and after

71 the 313th day of the year the solar installation cannot deliver the designed power demand (9.2 kW) to the process loop. This is happening because during the winter months (January, February, November and December) energy is generated by the solar field and accumulated to the buffer tank as normal, but due to the limited generation of solar energy the buffer tank is not possible to be charged fully (full mixed buffer tank, paragraph 3.3.2); thus the generated heat is “trapped” at the bulk of the thermal mass.

24 22 operation hours per day 20 18

16 14 12 10

hours per day per hours 8 6 4 2 0 1 31 61 91 121 151 181 211 241 271 301 331 361 day of the year

Figure 5.9 Operation hours of the solar thermal installation for the designed thermal load

From Figure 5.9 becomes clear that due to the limited solar potential during winter months the solar thermal system is not able to perform as has been designed to. The reasons for this anomaly, source from the large variation of the irradiation throughout the year in Erlangen and from the model used to describe the buffer tank of the solar system (paragraph 3.3.2). By choosing not to operate the power cycle (and subsequently the solar thermal installation),during the months November, December, January and February the annual energy demand is decreased to 97,667 kWh (instead of 145,504 kWh); thus the average annual solar fraction rises to 29.1% (Table 5.4), compared to the value of 19.5% that had been presented in Table 5.3.

Energy Demand Solar Field Energy Delivered Solar Fraction Month [kWh] Productivity [kWh] [kWh] [%] January - - - - February - - - - March 12,358 2,378 1,509 12.2 April 11,959 3,895 3,175 26.5 May 12,358 5,263 4,517 36.5 June 11,959 6,135 5,331 44.6 July 12,358 5,423 4,887 39.5 August 12,358 4,287 3,563 28.8 September 11,959 3,926 3,332 27.9 October 12,358 2,814 2,102 17 November - - - - December - - - - Year 97,667 34,121 28,414 29.1 Table 5.4 Influence of the solar fractions of the solar system with the integrated buffer tank

5.3 Comparison to the solar systems’ productivity in Spain As discussed previously, the performance of the studied solar systems has been calculated with reference the existing solar potential in South Germany (Erlangen). An interesting investigation would be the influence of the local weather conditions on the productivity of the studied solar thermal systems. The comparison of the energy productivity of the two systems will be performed between the reference location (Erlangen) and a location with higher solar irradiation, such as Valencia (Spain). The availability of the solar irradiance (W/m2) measured in hours throughout the year for both locations, is given in Figure 5.10. For both locations the default tilt angle of the collector (30˚) and the south orientation have been taken into account. As shown, all the ranges above the [300-400] W/m2 range, experience more hours in Valencia compared to Erlangen. For example, the ranges [700-800], [800-900] and [900-1,000] W/m2 are present 4.2 ,1.8 and 1.7 times more in Valencia against Erlangen, respectively. On the other hand, as expected the sun is shining less hours in Erlangen compared to Valencia, resulting in 4,199 and 4,065 of night hours (0 W/m2), respectively. More information regarding the solar potential in Erlangen and Valencia is given in APPENDIX C.

Figure 5.10 Solar potential in Erlangen (Germany) and Valencia (Spain), respectively

The first installation that will be compared at the two reference locations is the system that consists of seven collectors and coupled directly to the process loop. The energy productivity and the solar fraction for both locations are presented in Figure 5.11 and Table 5.5, respectively, and the monthly figures of the compared solar system are given in APPENDIX E1. As shown earlier, the annual generated energy by the solar field in Erlangen is 18,892 kWh, while the same solar field in Valencia generates 33,254 kWh. In other words the annual heat generated in Erlangen is 56% less compared to the one in Valencia. In more detail, the highest amount of energy for both locations is generated in June with 3,054 kWh and 3,721 kWh, in Erlangen and Valencia, respectively. Similarly, the month with the lowest amount of energy is experienced in December with 152 kWh in Erlangen and 1,474 kWh in Valencia.

Regarding the achieved solar fractions, the same installation in Valencia is possible to cover 22.9% of the demand of the cycle’s heating process, while in Germany 13%, annually. It is worth mentioning that the solar fraction per month in Spain remains higher than 10% throughout the year, but in Germany the lowest solar fraction is 1.2 (in December). Moreover, it is observed that during June the solar fraction in Erlangen (25.5%) is only 18% less compared to Valencia (31.1%). On the contrary, in December which is the month with the lowest solar fractions for both locations, the solar fraction in Erlangen (1.2%) is approximately 10 times less compared to the one in Valencia (11.9%).

14000 Energy Demand [kWh] Erlangen [kWh] Valencia [kWh] 12000

10000

8000

kWh 6000

4000

2000

Figure 5.11 Comparison of the direct coupled solar field (7 collectors) in Erlangen and Valencia

Month Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec Year Erlangen [%] 2 4.9 11.4 17.1 22 25.5 21.8 17.3 16.8 12.4 2.9 1.2 13 Valencia [%] 16.6 17.3 23.4 28 28.4 31.1 29.4 27.3 26.3 20.5 13.8 11.9 22.9 Table 5.5 Solar fractions of the direct coupled solar field (7 collectors) in Erlangen and Valencia

Similar behavior presents the solar system with the integrated buffer tank (Figure 5.12 and Table 5.6). The annual energy delivered to the heating process of the cycle by the solar system is 27,609 kWh and 55,033 kWh, in Erlangen and in Valencia, respectively. The productivity of the latter is almost double compared to the energy delivered in Erlangen. In June the solar system feeds the process loop with 6,516 kWh in Valencia and 5,256 kWh in Erlangen. Furthermore, the month with the lowest contribution for Spain is December (1,601 kWh), while energy is not delivered during January, February and December in Erlangen. The monthly figures of the compared solar system are given in APPENDIX E2.

The solar fractions in Erlangen and Valencia are presented in Table 5.6 with annual values 19% and 37.8%, respectively. The highest values experienced are 44% in Erlangen and 54.5% in Valencia (June). Due to the favorable solar potential in Valencia, in months May, June, July and August, the solar fraction remains over 50%. As expected, since there is no energy delivered in Erlangen in months January, February and December, the solar fraction is 0.

As discussed previously, in a solar thermal installation the total number of operation hours is mainly affected by the sizes of the solar field, the buffer tank and the thermal load. Although the present comparison performed under the same solar field configuration (2x12 collectors) and under the same load at the heat exchanger (9.2 kW), the size of the buffer tank in Valencia has been adjusted to 28 m3 (instead of 20 m3 used in Erlangen). The reason for this adjustment is the higher solar potential experienced in Valencia; thus a larger buffer tank has been employed with target to maintain the maximal temperature in the buffer tank to 180˚C. The annual number of the operation hours of the

75 solar system in Valencia, reaches 5,946 hours with a utilization factor 0.67 (5,926/8,760), while in Erlangen the operation hours are 2,993 and the utilization factor 0.34 (2,993/8,760).

14000 Energy Demand [kWh] Erlangen [kWh] Valencia [kWh] 12000

10000

8000

kWh 6000

4000

2000

Figure 5.12 Comparison of the solar systems with an integrated buffer tank (24 collectors) in Erlangen and Valencia

Month Jan Feb Mar Apr May Jun Jul Aug Sep Oct Nov Dec Year Erlangen [%] 0 0 10.3 25.7 35.9 44 38.5 28.2 27.1 16.1 0.9 0 19 Valencia [%] 24.2 24.4 37.2 47.9 51.2 54.5 51.4 51.2 47.1 33 18.1 13 37.8 Table 5.6 Solar fractions of the solar systems with an integrated buffer tank (24 collectors) in Erlangen and Valencia

5.4 Combined heat and power Although the primary function of the power cycle is to generate electricity, the high amount of heat rejected by the condenser, make wise the utilization of the rejected heat for other purposes (heat recovery). In the given sub-chapter, the exploitation of the waste heat of the cycle for space heating of the so called new building (space heating) will be investigated.

  From the power cycle analysis in sub-chapter 5.1 is known that Qboiler =13 kW , Wpump =0.08 kW and  Wturb =1.2 kW . From the energy equilibrium of the power cycle, the thermal power extracted from the  condenser, Qcond , can be derived by the rest components of the steam cycle:

  QQWWcond= boiler + pump − turb (5.13)

 The rejected thermal power from the condenser is calculated to Qcond =11.7 kW . The power cycle is assumed to operate continuously throughout the year; thus a constant rejection of heat from the condenser is realized. The total thermal energy rejected by the heat condenser is 102,607 kWh, annually.

Furthermore, in sub-chapter 4.2 the annual heat demand for space heating of the reference building had been calculated to 50,859 kWh (Figure 4.4), with a maximal power demand of 22.6 kW. From the presented load duration curve for space heating (Figure 4.5), becomes clear the thermal loads lower than 11.7 kW are applicable for more than 5,000 hours per year (or for more than 57% of the year). The monthly heat demand which is subject to thermal loads up to 11.7 kW is illustrated in Figure 5.13. 95% of the total energy demand is up to 11.7 kW, while only 5% of the required energy exceeds the power demand of the condenser. Thermal loads higher than 11.72 kW are mainly present in January, February and December, while they are absent in months June, July and August (detailed values are given in APPENDIX E3). Due to the reason that the condenser rejects heat under a constant power of 11.7 kW, the condenser is able to support the space heating system up to this capacity, while the remaining thermal power (11.7 up to 22.6 kW) should be supplied by other energy sources, such as an auxiliary heater or through the local district heating network.

10000 >11,7 kW up to 11,7 kW 9000 8000 7000 6000 5000 4000 3000 2000 1000

Heating demand for demand [kWh] for heating space Heating 0

Figure 5.13 Energy demand for space heating with a higher and lower capacity of 11.7 kW

As discussed previously, the upper threshold that the condenser is able to support the space heating system is 11.72 kW and the annual thermal energy rejected by the heat condenser is 102,607 kWh. Figure 5.14 presents the overall heat rejected by the condenser as a summation of the rejected heat utilized by the heating system of the new building (in grey color) and the excess of the rejected heat (in orange color). As expected the latter mainly occurs in summer months (June, July and August), when the demand for space heating is very limited (detailed values are given in APPENDIX E3). The highest heat recovery takes place is months January and February, with 88% and 85.8% of the condenser’s heat to be fed to the space heating installation. On the other hand, only 6% of the waste heat is recovered in July. In total, 47.2% (48,431 kWh) of the annual heat rejected by the condenser (102,607 kWh) can be utilized for space heating.

10000 Excess of the rejected heat Heat recovery by the space heating system 9000 8000 7000 6000 5000 4000 3000 2000 1000 Heat rejected rejected the by [kWh] Heat condenser 0

Figure 5.14 Heat rejected by the condenser of the cycle throughout the year

Chapter 6. Conclusions and Recommendations

6.1 Solar heating and cooling

6.1.1 Conclusions The following conclusions can be extracted after employing the solar thermal UHV collector as a heat source for heating and cooling in Erlangen (Germany):

As shown in paragraphs 4.2.1 and 4.3.1, the reference heating demand is mainly distributed in winter months (November to March), but most of heat required to operate an absorption chiller is needed from May to September. In more detail, 73% (37,227 kWh) of the annual heating demand profile (50,859 kWh) is distributed from November to March, and 78% (243,200 kWh) of the annual heating demand for solar cooling (311,760 kWh) occurs in the period May to September. As becomes clear, the vast majority of the demand of the two systems takes place in different periods and could be roughly described as the winter and summer period. On the other hand, due to the annual variation of the solar irradiation in Erlangen, the delivery of energy by the designed solar systems is affected crucially, as well. For instance, the energy delivered for solar heating by the examined system configuration C is 7% (2,340 kWh) of the heat delivered annually (32,865 kWh) in the period November-March. Moreover, during the period that 73% of the annual heat demand for solar cooling is being experienced (May to September), the energy delivered by the solar system to the 1-step absorption chiller is 73% (153,360 kWh) of the total energy delivered throughout the year (210,160 kWh). The above comparison shows clearly that the heat demand for solar cooling coincides with the productivity of the designed system, while this is not true for the systems designed for solar heating. From May to September the solar potential in Erlangen is favorable for heat production; thus there is a better match between the available solar irradiation and the solar cooling demand.

Solar Heating Concerning the solar systems designed to cover part of the heat demand of the building, the annual solar factor is affected crucially by the size of the solar field. The solar system comprised by 4 collectors is able to cover 15% of the annual demand for space heating. On the contrary, a solar system with 24 collectors is able to cover up to 65% of the space heating, depending on the size of the buffer tank and the capacity of the thermal load that the solar field is designed to operate with. Since a buffer tank is utilized in the designed systems, thermal losses occur from the buffer tank to the environment; thus the heat delivered by a solar system is less than the generated energy by the system’s solar field.

Furthermore, due to the large variation of the solar potential throughout the year in Erlangen, the energy delivered by a solar system with an integrated buffer tank is limited in winter months. In more detail the solar factors in winter months fluctuates from 0 to 0.03 depending on the studied number of collectors utilized in each system.

Solar Cooling The exploitation of the solar thermal UHV collectors for solar cooling has been realized through absorption chillers. Depending on the type of the chiller used (1-step or 2-step), the coefficient of

79 performance of the cooling system and the temperature level of the heat source are being affected. The former influences the amount of cooling that is provided by the installation throughout the year, while the second influences the characteristics of the solar system. Although the designed solar fields of the two systems operate in very different temperature levels, the power generated by both solar fields is similar. In more detail, at the reference irradiance (600 W/m2) the solar field that generates energy for the system equipped with a 1-step chiller, operates at temperatures Tin,field=86.5°C and Tout,field=100°C, while in the case of the 2-step chiller, the solar field operates at Tin,field=162.5°C and Tout,field=185°C. The 2 power delivered by the solar fields at an irradiance of 600 W/m (Qsf,ref) is 188.6 kW and the 163.6 kW, respectively. The above fact, combined to the relatively high annual solar factors experienced by the two cooling systems, show that the solar thermal UHV collector can be exploited efficiently for solar cooling in the studied location.

Another component of the solar system that is influenced by the type of the chiller utilized is the buffer tank. As shown, the system equipped with a 1-step chiller utilizes a buffer tank of 30 m3, while the system with the 2-step chiller a buffer tank of 110 m3. Due to the different operation temperatures of the two systems, the heat carrier of the first system is water (with a heat capacity, Cpwater,25˚C=4.2 kJ/(kgK)), while the heat transport fluid of the high temperature system is thermal oil (Cpthermal oil,25˚C=1.6 kJ/(kgK)). This large difference in the heat capacity of the two fluids, is compensated by using a larger storage volume (buffer tank) when thermal oil is being employed as a heat transport fluid.

6.1.2 Recommendations The suggestions proposed for further improvement of the studied solar systems are mainly related to the systems that the solar thermal UHV collector is being integrated to. Although, the solar thermal UHV collector is able to generate energy efficiently in medium to high temperatures (80˚C-200˚C), the design of the rest of the solar system influences crucially the operation of the overall installation. As discussed previously, the solar systems designed in Erlangen present limited delivery of heat in winter months, regardless if the system is oriented for heating or cooling purposes. Despite the fact that the main reason is the limited solar potential of Erlangen at the given period, the solar fields of the designed systems are possible to generate some amounts of energy (Figure 4.8). This energy is “trapped” in the utilized buffer tank as discussed in paragraph 4.2.3, due to the model of the buffer tank assumed (fully mixed buffer tank, paragraph 3.3.2). An alternative would be that in winter months the solar field ejects heat directly to the thermal load, instead of delivering the generated heat to the buffer tank. Another option would be the solar systems to utilize a variable buffer tank, with the target to use the optimal size of storage volume depending-for example-on the monthly solar irradiation. The variable buffer tank could be comprised by a number of smaller tanks connected in series. In this way, even the lowest amounts of heat generated could be exploited by the thermal loads, increasing the solar factor of the overall system. Realization of the suggested options could be used in the months January, February November and December, when the solar irradiation is limited.

Furthermore, as discussed in paragraph 4.3.2, the solar systems connected to a 1-step and to a 2-step absorption chiller utilize a buffer tank of 30 m3 and 110 m3, respectively. The large volume of the buffer tanks arises issues regarding the available space needed to install such a large storage tank, but also related to the costs involved (large volumes of thermal oil). An alternative would be other storage

80 options to be investigated for their integration to the solar thermal systems. Such solutions could be potentially provided by applying phase change materials (PCMs) or thermo-chemical materials (TCMs) as a storage medium.

6.2 Power generation

6.2.1 Conclusions The following conclusions can be extracted after employing the solar thermal UHV collector as a heat source for powering a small scale steam cycle in Erlangen (Germany):

The studied Rankine cycle is equipped with a steam turbine of 1.2 kW shaft power. The net electricity production Pnet,cycle=1 kWhe, while the heat demand of the cycle’s boiler is Qdem=13 kWth (ηnet,cycle =7.6). The cycle is characterized by a relatively low net efficiency and by high thermal power at the heating process.

Integration of the solar thermal UHV collectors can be realized in different ways, depending on the targets and the constraints applied. In the given investigation, two possible integration scenarios had been studied. The design of the solar field coupled to the process loop (7 collectors) and a solar system with comprised by 24 solar thermal collectors and an integrated buffer tank of 20 m3.

Although for both systems the annual solar fractions experience relatively low values, the solar system with an integrated buffer tank allows the overall installation to cover higher solar fraction of the cycle’s total demand for heat (19.5%). On the other hand the annual solar fraction of the direct coupled solar field (7 collectors) is 13%. As expected both systems operate more efficiently in summer with the solar fraction of June to reach 44.6% and 25.5%, respectively.

The productivity of the studied solar thermal installations is affected by the local weather conditions that the systems are applied to. By applying the systems in an area with higher solar irradiation such as in Valencia (Spain), the productivity of the solar systems increases significantly. For the solar field coupled to the process loop (7 collectors), the annual generated energy by the solar field in Valencia covers 22.9% of the demand of the cycle’s heating process, while by choosing the solution with the integrated buffer tank (24 collectors) the annual contribution of the solar system is 37.8%.

Furthermore, exploitation of the cycle’s rejected heat can be implemented to the heating system of the reference building. In this way, the heat generated by the solar system can be utilized further, instead of only powering the steam cycle. By supporting the heating installation of the building, heat recovery of 47% can be achieved.

The current study shows that although the applicability of the solar thermal UHV collectors for power generation is technically feasible, the low solar fractions achieved in Erlangen are relatively low. Possible realization of the studied systems, should take into account both the heat recovery that could be achieved by the cycle’s rejected heat and the financial feasibility of the overall system.

6.2.2 Recommendations Solar system with an integrated buffer tank and with a bypass operation of the solar field As shown in paragraph 5.2.3, due to the large irradiance variation between winter and summer in Central Europe, the solar system does not deliver thermal energy in months January, February, November and December because of the exploitation of a large buffer tank (20 m3). On the other hand, the buffer tank cannot be smaller due to the high temperatures that will be experienced by the thermal oil inside the tank. An alternative could be that in winter months, instead of buffering the generated heat to the storage tank, it could be ejected directly to the secondary loop of the solar system, bypassing the buffer tank (Figure 6.1). In this way, heat would not be “trapped” at the thermal mass of the buffer tank, but exploited directly. The size and the design of the solar field should be such that even at low irradiance levels (e.g. 350-450 W/m2) would generate enough energy to be ejected at the secondary loop. In summer, when the sun is shinny and the solar field generates enough heat, this energy could be temporarily stored in the solar system’s buffer tank.

Figure 6.1 Solar field that bypasses the buffer tank during winter months

The proposed solution combines elements from the systems proposed in sub-chapter 5.2. In winter months the system acts as solar field connected to the process loop via the heat exchanger, while in summer months as a regular solar system installation with an integrated buffer tank. Shifting between the two modes of the system could be achieved either by a threshold irradiance value, or by a pre- selected time period (e.g. winter and summer season).

Heat addition to the working fluid of the Rankine cycle As discussed in paragraph 5.1.3, an alternative of using an intermediate circuit and a heat exchanger for heat addition to the power cycle, is to add thermal energy directly to the working fluid of the cycle (water). In the present paragraph two options will be considered. The first one is achieved by ejecting heat at one (or more) specific section(s) of the heating process, with target to cover heat content of the economizer, evaporator or superheater. In the second option the working fluid of the cycle (water) is pressurized in much higher pressure compared to the feed pump of the cycle, then heat is added to the

82 pressurized water and finally is flashed in a vessel with lower pressure with target to generate steam that will run the turbine (Figure 6.2).

Figure 6.2 Heat addition to the working fluid (water) of the steam cycle

From the analysis of the power demand performed in paragraph 5.1.2, has been shown that the most of the energy is demanded during the evaporation process. In more detail the three different stages of the economizer, evaporator and superheater require 2.31 kW, 10.48 kW and 0.25 kW, respectively. This fact, highlights that the heat addition at the economizer (from 45˚C to 155˚C) and at the superheater (from 155˚C to 180˚C) stages, is not as energy intense as at the evaporation process. Therefore, a possible integration of a solar field with UHV collectors should be realized at the evaporation stage of the cycle’s heating process, while the rest energy demand could be supported by an auxiliary burner. This could be achieved by using three plate heat exchangers (PHE) connected in series that would perform each stage of the heating process (Figure 6.3). The intermediate PHE, which plays the role of the evaporator, will be connected to the solar system. Depending on the design of the overall configuration, the solar system could also feed partially the economizer, in order the cycle to exploit more the solar potential. It is worth mentioning that the proposed configuration, would demand the solar field to operate on high temperature (>Tsat_water@5,5bar=155˚C), something that would affect the efficiency of the solar thermal UHV collectors. Furthermore, suitable design of the auxiliary burner should be realized, in order the cycle to maintain the continuous operation when the solar system cannot deliver energy.

Figure 6.3 Solar energy ejected to the evaporator of the heating process of the Rankine cycle

Another recommendation, in regards to the direct heat addition to the working fluid of the cycle, would be the exploitation of a flash device for steam generation (Figure 6.4). In this case, the working fluid is pressurized on a higher level than the cycle’s feed pump normally would do (circuit in red color), in order to remain in liquid state during the heating process of the cycle and until enters to the flash vessel. Due to the pressure difference inside the flash vessel (lower pressure environment), the pressurized liquid water will be evaporated, with result the generated steam to be fed to the steam turbine. The heat addition could be realized, either by a solar field where the pressurized water could be heated up directly by the UHV collectors, or by an auxiliary burner, controlled by a 4-way valve. One of the limitations of this setup could be the constraint to use water as a heat transport fluid of the solar installation, thus low temperatures in winter could limit drastically the operation hours of the solar installation. Moreover, design constraints of the UHV collector limit the maximal pressurized liquid that can be applied to the collectors to a pressure of 10 bar (for water: Tsat@10bar=180˚C).

Figure 6.4 Exploitation of a flash vessel for steam production of the Rankine cycle

References

[1] C. Benvenuti, VACUUM FOR THERMAL INSULATION: THE SRB SOLAR PANEL AS AN EXAMPLE, SRB Energy Research SARL, c/o CERN, CH-1211 Geneve 23, Switzerland

[2] A. Binz, A. Moosmann, G. Steinke, U. Schinhardt, F. Fregnan, H. Simmler, S. Brunner, K. Ghazi, R. Bundi, U. Heinemann, H. Schwab, H. Cauberg, M. Tenpierik, G. Johannesson. T. Thorsell, M. Erb, B. Nussbaumer, Vacuum Insulation in the Building, Edition 2013, HiPTI- High Performance Thermal Insulation IEA/ECBCS Annex. [Online]. Available: http://www.ecbcs.org/docs/Annex_39_Report_Subtask-B.pdf

[3] V. Lindberg, Chapter 5 Kinetic Theory and Vacuum, April 26, 2012. [Online]. Available: http://people.rit.edu/vwlsps/LabTech/KineticTheory.pdf

[4] P. Danielson, Choosing the Right Vacuum Materials, THE VACUUM LAB, vol. 36. [Online]. Available: http://www.vacuumlab.com/Articles/VacLab36.pdf

[5] P. Danielson, How to Use Getter and Getter Pumps, THE VACUUM LAB, vol. 23. [Online]. Available: http://www.vacuumlab.com/Articles/VacLab23.pdf

[6] C. Benvenuti, NON-EVAPORABLE GETTERS: FROM PUMPING STRIPS TO THIN FILM COATINGS, CERN, CH-1211 Geneva 23. [Online]. Available: http://accelconf.web.cern.ch/AccelConf/e98/PAPERS/THZ02A.PDF

[7] C. Benvenuti, PARTICLE ACCELERATORS AND SOLAR PANELS, EVOLUTION OF THE GETTER TECHNOLOGY DURING THE LAST YEARS [Online]. Available: http://prometeo.sif.it/papers/online/sag/029/01-02/pdf/05-fisica-e.pdf

[8] P.Kovac, A guide to the standard EN 12975, Quality Assurance in solar thermal heating and cooling technology- keeping track with recent and upcoming developments, QAiST-IEE/08-593/SI2.529236, Deliverable D2.3, 28.05.2012. [Online]. Available: http://www.estif.org/fileadmin/estif/content/projects/QAiST/QAiST_results/QAiST%20D2.3%20Guide% 20to%20EN%2012975.pdf

[9] F. Mauthner, W. Weiss, Solar Heat Worldwide Markets and Contribution to the Energy Supply 2011, Edition 2013, IEA Solar Heating & Cooling Programme, (May 2013). [Online]. Available: http://iea-shc.org/solar-heat-worldwide

[10] TECSOL, Solar Cooling systems with cooling power larger than 20 kW, SHC-IEA, Task 38 Solar Air Conditioning and Refrigeration. [Online]. Available: http://www.tecsol.fr/Rafrsol2/downloads/large_scale_SCP_lists_31Aug09_TECSOL.pdf

APPENDIX A: Project parameters questionnaire

APPENDIX B: Solar and geometric equations

Variable Symbol Unit Equation Optical G GΤ− TT Τ− η ηη=b IAM +b 0.95 ++αav amb α av amb efficiency of opt - opt 012 GG+ GG +GG ++ GG the collector bd bd bd bd 11  1− 0.16− 1 , for 0 <1 − 0.16 − 1 < 1 Incidence cosθθ cos Angle IAM Deg. IAM =   1  Modifier  0 ,for 1 − 0.16−∉ 1( 0,1)   cosθ  Direct irradiance on a tilted Gb W GIb= DNI cosθ horizontal surface Diffuse irradiance 1+ cos β on a tilted Gd W GId= diff  horizontal 2 surface Average temperature TT+ Τ Deg. in,, load out load of the av Τ=av thermal Celsius 2 load

θ=+−aacos sinss cosβ cos a sin β cos( γγ s) Angle of  θ Deg. incidence ° asz+=θθ90 Ca ⇒ sins = cos z  −cosδω sin tanδ 180− a sin  ,for cosω ≥  Solar cos as  tanϕ azimuth γ s Deg. γ s =  angle  −cosδω sin tanδ asin  ,for cosω <   cos as tanϕ Solar altitude as Deg. aas = sin[ sinϕδ sin+ cos ϕ cos δ cos ω] angle Hour angle ω Deg. ω =(H −12) 15 284 + N Declination δ Deg. δ = 23,45sin 360 365

Τamb Ambient temperature [Deg. Celsius]

Tin, load Inlet temperature of the thermal load [Deg. Celsius]

Tout, load Outlet temperature of the thermal load [Deg. Celsius]

IDNI Direct normal irradiance [W]

Idiff Diffuse horizontal irradiance [W]

γ Surface azimuth angle [Deg]

β Slope of the collector [Deg]

ϕ Latitude of the location [Deg]

H Hour of the day

N Day of the year

APEENDIX C: Solar Irradiation data

The calculations have been performed for the locations: Erlangen (Germany) and Valencia (Spain). The irradiation data for the calculations has been taken from the SODA database (http://www.soda-is.com/eng/index.html).

Erlangen Valencia units (Germany) (Spain) Latitude 49.58 39.46 N Longitude 11.03 -0.36 E Orientation 180 deg Tilt 30 deg 2 Maximal Irradiance through the year 969.15 992.28 W/m 2 Normal Direct Irradiation, annual 1,039.0 1,850.3 kWh/m /year 2 Diffused Horizontal Irradiation, annual 479.0 557.1 kWh/m /year 2 Total Irradiation fixed tilt, total 1,108.8 1,757.2 kWh/m /year

hours

2 0˚ 30˚ W/m Erlangen Valencia Erlangen Valencia 0 4,200 4,066 4,199 4,065 0.-99. 1,889 1,245 1,886 1,256 100-199. 763 525 722 503 200-299. 465 459 420 403 300-399. 377 508 343 331 400-499. 316 454 288 406 500-599. 252 367 229 322 600-699. 222 327 209 476 700-799. 158 343 192 336 800-899. 117 301 179 319 900-999. 1 165 93 343 1000-1,099. 0 0 0 0 2 100 W/m 2,671 3,449 2,675 3,439 2 <100 W/m 6,089 5,311 6,085 5,321 total hours per year 8,760

C1. Erlangen (Germany)

Irradiation Data

Diffused Irradiation fixed Normal Direct Month Num Days/month 2 Horizontal tilt [kWh/m /month] 2 2 [kWh/m /month] [kWh/m /month] January 1 31 20 13 24 February 2 28 32 21 39 March 3 31 74 39 85 April 4 30 106 51 118 May 5 31 145 65 151 June 6 30 167 70 168 July 7 31 136 68 150 August 8 31 109 58 124 September 9 30 115 41 114 October 10 31 96 28 89 November 11 30 25 15 29 December 12 31 15 10 18

260 DNI Irradiation Fixed Tilt Diffused Horizontal 240 220 200 180 160

2 140 120 kWh/m 100 80 60 40 20 0

C2. Valencia (Spain)

Irradiation Data

Diffused Irradiation fixed Normal Direct Month Num Days/month 2 Horizontal tilt [kWh/m /month] 2 2 [kWh/m /month] [kWh/m /month] January 1 31 125 26 109 February 2 28 106 32 103 March 3 31 153 48 152 April 4 30 180 56 175 May 5 31 194 65 186 June 6 30 215 64 195 July 7 31 207 64 192 August 8 31 182 62 177 September 9 30 169 49 164 October 10 31 139 40 133 November 11 30 95 27 90 December 12 31 86 24 80

260 DNI Irradiation Fixed Tilt Diffused Horizontal 240 220 200 180 160

2 140 120 kWh/m 100 80 60 40 20 0

APPENDIX D: Solar heating and solar cooling results tables

D1. Heating demand of the new office building

Energy Month Demand [kWh] January 8,996 February 7,645 March 6,274 April 4,068 May 2,134 June 1,034 July 544 August 739 September 1,526 October 3,585 November 6,044 December 8,268 Year 50,859

D2. Solar heating: Energy generated and delivered by the solar fields and the systems

Energy generated by the solar fields

Heat Demand 4 collectors 24 collectors Month [kWh] [kWh] [kWh]

January 8,996 146 535 February 7,645 271 1,210 March 6,274 624 3,155 April 4,068 896 4,762 May 2,134 1,152 6,247 June 1,034 1,295 7,117 July 544 1,166 6,377 August 739 954 5,155 September 1,526 878 4,728 October 3,585 670 3,519 November 6,044 197 806 December 8,268 104 360 Year 50,859 8,354 43,972

Energy delivered by the solar systems

Heat 24 24 24 4 collectors, Month Demand collectors, 5 collectors, collectors, 5 kW [kWh] [kWh] kW [kWh] 10 kW [kWh] 15 kW [kWh]

January 8,996 70 0 0 0 February 7,645 200 0 0 150 March 6,274 555 230 1,290 1,995 April 4,068 830 2,415 3,160 3,660 May 2,134 1,095 3,280 4,700 5,145 June 1,034 1,240 3,600 5,580 6,060 July 544 1,120 3,720 4,910 5,355 August 739 905 3,155 3,670 4,140 September 1,526 825 2,850 3,350 3,720 October 3,585 610 1,270 1,960 2,445 November 6,044 130 0 30 195 December 8,268 45 0 0 0 Year 50,859 7,625 20,520 28,650 32,865

D3. Heating demand profile of the cooling system

Thermal Energy Free cooling Thermal Energy Demand Month Demand without Free (Tcut,in=-10˚C, Tcut,out=10˚C) with Free Cooling Cooling [kWh] [kWh] [kWh] January 59,520 55,280 4,240 February 53,760 49,600 4,160 March 59,520 52,560 6,960 April 57,600 35,520 22,080 May 59,520 19,120 40,400 June 57,600 8,800 48,800 July 59,520 3,360 56,160 August 59,520 5,600 53,920 September 57,600 13,680 43,920 October 59,520 34,720 24,800 November 57,600 52,560 5,040 December 59,520 58,240 1,280 Year 700,800 389,040 311,760

D4. Solar Cooling: Energy generated and delivered by the solar fields and the systems

1-step absorption chiller

Thermal Solar Energy Solar Energy Cooling energy Solar Fraction Month Produced Delivered Produced Demand [%] [kWh] [kWh] (energy delivered) [kWh] [kWh] January 4,240 2,976 0 0.0 0 February 4,160 6,918 1,840 44.2 1,288 March 6,960 18,399 13,600 195.4 9,520 April 22,080 28,005 23,360 105.8 16,352 May 40,400 36,858 32,400 80.2 22,680 June 48,800 42,072 37,760 77.4 26,432 July 56,160 37,660 33,360 59.4 23,352 August 53,920 30,374 26,080 48.4 18,256 September 43,920 27,855 23,760 54.1 16,632 October 24,800 20,642 16,240 65.5 11,368 November 5,040 4,553 1,760 34.9 1,232 December 1,280 2,002 0 0.0 0 Year 311,760 258,314 210,160 67.4 147,112

2-step absorption chiller

Thermal Solar Energy Solar Energy Cooling energy Solar Fraction Month Produced Delivered Produced Demand [%] [kWh] [kWh] (energy delivered) [kWh] [kWh] January 4,240 371 0 0.0 0 February 4,160 2,317 0 0.0 0 March 6,960 8,756 2,080 29.9 2,496 April 22,080 16,040 13,200 59.8 15,840 May 40,400 22,305 19,520 48.3 23,424 June 48,800 26,928 24,240 49.7 29,088 July 56,160 23,391 20720 36.9 24,864 August 53,920 17,808 15,040 27.9 18,048 September 43,920 16,624 14,240 32.4 17,088 October 24,800 10,948 7,840 31.6 9,408 November 5,040 1,277 160 3.2 192 December 1,280 191 0 0.0 0 Year 311,760 146,956 117,040 37.5 140,448

APPENDIX E: Power generation results tables

E1. Energy delivered by the direct coupled solar field in Erlangen and Valencia

Solar energy delivered Energy [kWh] Month Demand [kWh] Erlangen Valencia [kWh] [kWh] January 12,358 250 2,051 February 11,162 550 1,928 March 12,358 1,411 2,889 April 11,959 2,043 3,350 May 12,358 2,716 3,507 June 11,959 3,054 3,721 July 12,358 2,692 3,633 August 12,358 2,134 3,370 September 11,959 2,014 3,147 October 12,358 1,531 2,531 November 11,959 345 1,653 December 12,358 152 1,474 Year 145,504 18,892 33,254

E2. Energy delivered by the solar system with an integrated buffer tank in Erlangen and Valencia

Solar energy delivered Energy [kWh] Month Demand [kWh] Erlangen Valencia [kWh] [kWh] January 12,358 0 2,990 February 11,162 0 2,721 March 12,358 1,268 4,600 April 11,959 3,073 5,729 May 12,358 4,433 6,321 June 11,959 5,266 6,516 July 12,358 4,757 6,349 August 12,358 3,480 6331 September 11,959 3,239 5,627 October 12,358 1,990 4,082 November 11,959 102 2,166 December 12,358 0 1,601 Year 145,504 27,609 55,033

E3. Heat Recovery by the cycle’s condenser for space heating

Heating Total Heating Heating Demand Heat Recovery Excess of Demand (up to Month Demand ( >11,7 kW) from the cycle waste heat 11,7 kW) [kWh] [kWh] [kWh] [kWh] [kWh] January 8,996 7,677 1,319 8,720 1,042 February 7,645 6,690 955 7,876 1,186 March 6,274 6,005 269 8,720 2,714 April 4,068 4,014 55 8,438 4,425 May 2,134 2,134 1 8,720 6,586 June 1,034 1,034 0 8,438 7,405 July 544 544 0 8,720 8,175 August 739 739 0 8,720 7,980 September 1,526 1,526 0 8,438 6,912 October 3,585 3,574 11 8,720 5,146 November 6,044 5,914 130 8,438 2,525 December 8,268 7,480 788 8,720 1,240 Year 50,859 48,431 3,528 102,667 54,236

100