Solar Cooling Handbook: A Guide to Solar Assisted Cooling and Dehumidification Processes 9783990434390, 9783990434383

Table of contents :
Contents
Preface
Notes from the editors
1 Introduction
2 Meteorological data, heating and cooling loads and load sub-systems
2.1 Solar radiation, ambient temperature and humidity
2.1.1 Average quantities
2.2 Availability of climatic data, sources of weather data
2.3 Building space heating, domestic hot water and air conditioning needs
2.3.1 Efficient building design practice
2.3.2 Heating and cooling load: definitions and calculation methods
2.3.3 Domestic hot water load profiles
2.4 Industrial heating and cooling
2.4.1 Preliminary analysis required data
2.4.2 System design data: thermal load profile
2.5 The load sub-system – air-conditioning equipment
2.5.1 All-air systems
2.5.2 Water systems
2.5.3 Air-water systems
3 Components of solar thermal systems
3.1 The solar thermal collector
3.1.1 Assessment of the collector’s thermal performance
3.1.2 C ollector yield for long term performance prediction
3.2 Solar thermal collector technologies
3.2.1 Flat-plate collectors
3.2.2 Solar air collector
3.2.3 Evacuated tube collectors
3.2.4 Evacuated flat plate collectors
3.2.5 PV-thermal hybrid collectors
3.2.6 Stationary concentrating collectors
3.2.7 Solar concentrating tracking collectors (PTC, LFR)
3.2.8 Summary
3.3 Testing and certification of solar thermal collectors
3.3.1 Applicable test standards
3.3.2 C ertification schemes
3.4 Heat storage
3.4.1 Hot water stores
3.4.2 Storages with phase change materials
3.5 Backup heater
4 Heat driven cooling technologies: closed cycles
4.1 Principles of absorption and adsorption cooling
4.1.1 Absorption chillers
4.1.2 Adsorption chillers
4.2 Other closed cycles
4.3 Complementary components – Heat rejection systems
4.3.1 The challenge of heat rejection
4.3.2 Types of heat rejection devices
4.3.3 C old storage
5 Heat driven cooling technologies: open cycle systems
5.1 Principles and materials of desiccant cooling systems
5.2 Solid desiccant systems
5.2.1 System performance
5.2.2 Solar desiccant cooling systems (SDEC): examples, control and operation
5.2.3 Possible operational problems
5.2.4 Main components of solid DEC air handling units
5.3 Liquid dessicant systems
6 Solar cooling system characterization
6.1 Generic system schemes
6.1.1 Basic system topology
6.1.2 C omposition of generic systems
6.1.3 System control and hydraulics
6.1.4 Selection guide and system examples
6.2 Pre-engineered systems
6.3 Custom-made systems
6.3.1 L arge-capacity installations
6.3.2 Desiccant cooling systems
7 Energy and economic figures for solar cooling
7.1 Performance of conventional chillers
7.2 Performance of thermally driven chillers
7.3 Energy performance of solar driven cooling systems
7.3.1 Fractional PE savings
7.3.2 Primary energy sensitivity analysis of solar cooling systems
7.3.3 O ther useful energy performance parameters
7.4 Environmental impact analysis
7.5 Economic figures of solar cooling systems
8 Overall system design, sizing and design tools
8.1 Suitability analysis of a targeted building for a defined solar air-conditioning application
8.1.1 Presentation and objectives of the check-list
8.1.2 Selection of the appropriate system technology: the SAC decisionscheme
8.1.3 Selection of the proper type of solar collectors for the selected air-conditioning system and thermally driven cooling equipment
8.2 System sizing
8.2.1 Guidelines
8.2.2 Simple pre-design tools
8.2.3 Detailed simulation tools
9 Solar thermal system design
9.1 Field configuration parallel/series, high/low-flow
9.1.1 General characteristic of high/low-flow systems
9.1.2 Heat needs of solar cooling systems
9.1.3 Heat needs of domestic hot water and space heating preparation
9.1.4 Possible layouts and control strategies for collector fields for solar cooling systems with DHW and SH production (solar combi-plus-systems)
9.2 Stagnation of solar plants
9.2.1 Stagnation in collector fields
9.2.2 Implications of stagnation on the solar pump group
9.3 Stratification and necessary hot water storage tank volume
9.3.1 Heat input from solar collectors to the heat stores
9.3.2 Heat input from solar collectors into the heat store for solar combi-systems with solar cooling
9.3.3 Necessary volumes in the tank for solar combi-systems without cooling
9.3.4 Storage volume for solar combi-systems with solar cooling
9.3.5 Stratification
9.4 Other components of the solar loop for solar cooling systems
10 Pre-engineered systems: built examples and experiences
10.1 What can be expected from a pre-engineered system?
10.2 Built examples
10.3 Experiences
10.3.1 Installation issues
10.3.2 Commissioning
10.3.3 Maintenance issues
10.3.4 C ontrol issues
10.4 Recommendations for system suppliers
10.4.1 Electricity consumption of auxiliary components
10.4.2 Heat rejection components
10.4.3 Part load operation
10.4.4 Pressure drop in the system
10.4.5 Nominal flow rates – high temperature differences
10.4.6 Use of a cold store
10.4.7 Influence of heat rejection temperature
11 Experiences from installed custom made systems
11.1 Introduction
11.2 Built examples
11.2.1 Example 1: office building in Gleisdorf – Austria
11.2.2 Example 2: education centre in La Reunion island – France
11.2.3 Example 3: Industrial application in Grombalia – Tunisia
11.3 Experiences
11.3.1 C omponents integration and layouts
11.3.2 C omponent sizing
11.3.3 C ontrol strategies
11.3.4 Commissioning
12 DEC systems: built examples and experiences
12.1 Built examples
12.1.1 ENERGY base
12.1.2 Munich Airport
12.1.3 DREAM Unipa
12.2 Experiences
12.3 Control strategy definition
13 Summary and outlook
13.1 Overall technology status
13.2 Energy performance
13.3 Basic design guidelines and operation principles
13.4 Economics
13.5 Outlook
14 Appendix
14.1 The IEA Solar Heating & Cooling Programme
14.2 TASK 38 Solar Air-Conditioning and Refrigeration
14.2.1 Objectives
14.3 TASK 38 management structure
14.3.1 O perating Agent
14.3.2 Subtask Leaders
14.4 Institutions participating in Task 38

Citation preview

Editors: Hans-Martin Henning (Fraunhofer Institute for Solar Energy Systems ISE, Germany) Mario Motta (Politecnico di Milano, Italy) Daniel Mugnier (Tecsol, France)

This work is subject to copyright. All rights are reserved, whether the whole or part of the material is concerned, specifically those of translation, reprinting, re-use of illustrations, broadcasting, reproduction by photocopying machines or similar means, and storage in data banks. Product liability: The publisher can give no guarantee for the information contained in this book. The use of registered names, trademarks, etc. in this publication does not imply, even in the absence of a specific statement, that such names are exempt from the relevant protective laws and regulations and are therefore free for general use. © 2013 AMBRA | V AMBRA | V is part of Medecco Holding GmbH, Vienna Printed in Austria Cover Design: WMXDesign GmbH, Heidelberg, Germany Layout: Martin Gaal, Vienna, Austria Typesetting: Lena Appl, Vienna, Austria Printing and binding: Holzhausen Druck GmbH, Wolkersdorf, Austria Printed on acid-free and chlorine-free bleached paper With 226 (mainly colored) figures and 40 tables ISBN 978-3-99043-438-3  Ambra |V ISBN 978-3-211-73095-9 ISBN 978-3-211-00647-8

2. Auflage SpringerWienNewYork 1. Auflage SpringerWienNewYork

Hans-Martin Henning Mario Motta Daniel Mugnier (Eds.)

Solar Cooling Handbook A Guide to Solar Assisted Cooling and ­Dehumidification Processes 3rd Completely Revised Edition

Contents 9 Preface 11 Notes from the editors 13

1 Introduction

21

2 Meteorological data, heating and cooling loads and load sub-systems

21 26

2.1 Solar radiation, ambient temperature and humidity 2.1.1 Average quantities

27

2.2 Availability of climatic data, sources of weather data

29

2.3 Building space heating, domestic hot water and air conditioning needs

29

2.3.1 Efficient building design practice

30

2.3.2 Heating and cooling load: definitions and calculation methods

37

2.3.3 Domestic hot water load profiles

38

2.4 Industrial heating and cooling

39

2.4.1 Preliminary analysis required data

40

2.4.2 System design data: thermal load profile

41

2.5 The load sub-system – air-conditioning equipment

43

2.5.1 All-air systems

46

2.5.2 Water systems

49

2.5.3 Air-water systems

53 53

3 Components of solar thermal systems 3.1 The solar thermal collector

54

3.1.1 Assessment of the collector’s thermal performance

58

3.1.2 Collector yield for long term performance prediction

58

3.2 Solar thermal collector technologies

59

3.2.1 Flat-plate collectors

61

3.2.2 Solar air collector

63

3.2.3 Evacuated tube collectors

64

3.2.4 Evacuated flat plate collectors

65

3.2.5 PV-thermal hybrid collectors

66

3.2.6 Stationary concentrating collectors

68

3.2.7 Solar concentrating tracking collectors (PTC, LFR)

71

3.2.8 Summary

73

3.3 Testing and certification of solar thermal collectors

73

3.3.1 Applicable test standards

74

3.3.2 Certification schemes

75

3.4 Heat storage

76

3.4.1 Hot water stores

79

3.4.2 Storages with phase change materials

82

3.5 Backup heater

87 90

4 Heat driven cooling technologies: closed cycles 4.1 Principles of absorption and adsorption cooling

90

4.1.1 Absorption chillers

100

4.1.2 Adsorption chillers

104

4.2 Other closed cycles

105

4.3 Complementary components – Heat rejection systems

106

4.3.1 The challenge of heat rejection

106

4.3.2 Types of heat rejection devices

110

4.3.3 Cold storage

117

5 Heat driven cooling technologies: open cycle systems

117

5.1 Principles and materials of desiccant cooling systems

119

5.2 Solid desiccant systems

123

5.2.1 System performance

124

5.2.2 Solar desiccant cooling systems (SDEC): examples, control and operation

134

5.2.3 Possible operational problems

135

5.2.4 Main components of solid DEC air handling units

147

153 154

5.3 Liquid dessicant systems

6 Solar cooling system characterization 6.1 Generic system schemes

154

6.1.1 Basic system topology

155

6.1.2 Composition of generic systems

166

6.1.3 System control and hydraulics

169

6.1.4 Selection guide and system examples

171

6.2 Pre-engineered systems

173

6.3 Custom-made systems

173

6.3.1 Large-capacity installations

173

6.3.2 Desiccant cooling systems

175

7 Energy and economic figures for solar cooling

175

7.1 Performance of conventional chillers

179

7.2 Performance of thermally driven chillers

182

7.3 Energy performance of solar driven cooling systems

182

7.3.1 Fractional PE savings

185

7.3.2 Primary energy sensitivity analysis of solar cooling systems

189

7.3.3 Other useful energy performance parameters

191

7.4 Environmental impact analysis

191

7.5 Economic figures of solar cooling systems

207 209

8 Overall system design, sizing and design tools 8.1 Suitability analysis of a targeted building for a ­defined solar air-conditioning application

209

8.1.1 Presentation and objectives of the check-list

216

8.1.2 Selection of the appropriate system technology: the SAC ­decision scheme

231

8.1.3 Selection of the proper type of solar collectors for the selected air-conditioning system and thermally driven cooling equipment

234

8.2 System sizing

234

8.2.1 Guidelines

236

8.2.2 Simple pre-design tools

241

8.2.3 Detailed simulation tools

249 249

9 Solar thermal system design 9.1 Field configuration parallel/series, high/low-flow

249

9.1.1 General characteristic of high/low-flow systems

251

9.1.2 Heat needs of solar cooling systems

251

9.1.3 Heat needs of domestic hot water and space heating preparation

252

9.1.4 Possible layouts and control strategies for collector fields for solar cooling systems with DHW and SH production (solar combi-plus-systems)

254

9.2 Stagnation of solar plants

254

9.2.1 Stagnation in collector fields

257

9.2.2 Implications of stagnation on the solar pump group

258

9.3 Stratification and necessary hot water storage tank volume

258

9.3.1 Heat input from solar collectors to the heat stores

260

9.3.2 Heat input from solar collectors into the heat store for solar combi-systems with solar cooling

260

9.3.3 Necessary volumes in the tank for solar combi-systems without cooling

262

9.3.4 Storage volume for solar combi-systems with solar cooling

262

9.3.5 Stratification

263

9.4 Other components of the solar loop for solar cooling systems

265

10 Pre-engineered systems: built examples and experiences

266

10.1 What can be expected from a pre-engineered system?

267

10.2 Built examples

281

10.3 Experiences

281

10.3.1 Installation issues

282

10.3.2 Commissioning

282

10.3.3 Maintenance issues

283

10.3.4 Control issues

283

10.4 Recommendations for system suppliers

284

10.4.1 Electricity consumption of auxiliary components

285

10.4.2 Heat rejection components

287

10.4.3 Part load operation

287

10.4.4 Pressure drop in the system

288

10.4.5 Nominal flow rates – high temperature differences

288

10.4.6 Use of a cold store

288

10.4.7 Influence of heat rejection temperature

291

11 Experiences from installed custom made systems

292

11.1 Introduction

293

11.2 Built examples

294

11.2.1 Example 1: office building in Gleisdorf – Austria

299

11.2.2 Example 2: education centre in La Reunion island – France

304

11.2.3 Example 3: Industrial application in Grombalia – Tunisia

309

11.3 Experiences

309

11.3.1 Components integration and layouts

320

11.3.2 Component sizing

324

11.3.3 Control strategies

330

11.3.4 Commissioning

333 334

12 DEC systems: built examples and experiences 12.1 Built examples

334

12.1.1 ENERGY base

337

12.1.2 Munich Airport

340

12.1.3 DREAM Unipa

343

12.2 Experiences

347

12.3 Control strategy definition

349

13 Summary and outlook

349

13.1 Overall technology status

350

13.2 Energy performance

351

13.3 Basic design guidelines and operation principles

353

13.4 Economics

354

13.5 Outlook

357

14 Appendix

357

14.1 The IEA Solar Heating & Cooling Programme

358

14.2 TASK 38 Solar Air-Conditioning and Refrigeration

358 359 359 359 360

14.2.1 Objectives 14.3 TASK 38 management structure 14.3.1 Operating Agent 14.3.2 Subtask Leaders 14.4 Institutions participating in Task 38

Preface There is no doubt that drastic reduction of greenhouse gas emissions is needed in order to limit global warming and related dramatic changes in conditions for human life on earth. Consumption of fossil fuels is recognized as the most important cause of greenhouse gas emissions. In almost all countries buildings are among the main sources of fossil fuel consumption, directly and indirectly through the use of electricity. Therefore, reducing fossil fuel consumption of buildings will play a key role in transforming our energy system towards sustainability. Energy efficiency and the use of renewable energy are two important measures to significantly reducing the consumption of conventional energy in buildings. It is becoming increasingly obvious that buildings of the future have to be as energy neutral as possible, at least on a seasonal level. This can only be achieved by exploiting measures for energy conservation and efficiency and by covering the remaining demand by locally available renewable energy sources. Here solar energy plays a major role. Solar thermal energy can be converted into heat and used directly for heating and hot water production or indirectly for cooling through the use of thermally driven cooling devices. Therefore solar thermally driven heating and cooling systems represent a major technical solution for energy neutral buildings. This handbook aims to provide comprehensive information about solar thermal energy systems used for air-conditioning in buildings and also for applications in industrial refrigeration. We wish to contribute towards a continuous deployment of solar thermally driven cooling technology by providing information on suitable components and systems. The main focus is on technologies and equipment which are commercially available today or which are undergoing pilot tests. The key components covered are thermally driven cooling systems and solar collectors, as the major heat source to drive the cooling system. This book also puts a major focus on the entire systems including all auxiliary components and experiences derived from real life installations. The main goals of this handbook are to encourage planners and potential users to consider the installation of solar-assisted cooling systems and to provide them with helpful information during the decision-making and the design processes. A properly designed and carefully operated installation will give a high degree of satisfaction by providing a high level of indoor comfort to the users while using environmentally friendly technologies.

9

Notes from the editors This handbook is the product of a co-operative initiative carried out by experts from fourteen countries in Task 38 “Solar-Assisted Air-Conditioning and Refrigeration” of the Solar Heating & Cooling Programme (SHC) of the International Energy Agency (IEA). All contributing authors are grateful to the national funding authorities that enabled work within Task 38, as well as support for the production of this handbook. Each chapter has been produced by one or more responsible authors. In some cases co-authors contributed on particular issues. The editors provided the overall structure of the handbook and went through the whole text with the aim of streamlining the whole content in a coherent way without neglecting the individual approach and style of each single responsible author. Thanks are due to all Task 38 participants, authors and co-authors, who followed the iterative process of writing and reviewing this handbook. A process which included extensive discussions and sometimes long iteration loops. Of all the persons who contributed to the production of this book, some should be named personally. Particular thanks are due to: Marcus Ben Jones, Sufia Jung and Lucio Mesquita who made a very comprehensive review of the entire book and added British units. Elena Tröndle had a huge task in gathering all the figures, tables, working on copyright issues and making all the text and material available to the publisher in a usable form and format. David Marold and Angelika Heller from ʻAMBRA ǀV publishingʼ were not only very flexible with regard to timing but also provided us with many valuable tips on the transfer of technical information to the target audiences.

11

Chapter 1

Introduction Responsible Authors:

Hans-Martin Henning, Fraunhofer Institute for Solar Energy Systems, Germany Mario Motta, Politecnico di Milano, Italy Daniel Mugnier, Tecsol, France

Buildings represent one of the dominating energy-consuming sectors in industrialized societies. In Europe about 40% of primary energy consumption is due to services in buildings. Residential, commercial and industrial buildings consume energy for applications such as heating, hot water, air-conditioning, lighting and other – mainly electrically operated – equipment. Consequently in several countries legislation which aims to reduce energy consumption in buildings has been adopted in the last decade. For example, the European Parliament and the Council of the European Union adopted (in 2010) an update of the 2002 Energy Performance of Buildings Directive (EPBD) /1.1/. It includes a significant strengthening of the energy performance requirements of new and existing European buildings. It fixes 2020 (and 2018 for the public sector) as a deadline for new buildings to be “nearly zero energy”. A “nearly zero energy building” is a building that requires a very low amount of energy, which should be covered to a very significant extent by energy from renewable sources, produced on-site or nearby. A pre-condition to achieving a “nearly zero energy” standard is to maximize energy saving and energy efficiency in buildings. This minimizes the remaining energy demand in such a way that it becomes realistic to (a very large extent) cover it by renewable energy sources. Solar energy is often the most important renewable energy source available on-site and therefore is one of the favored candidates to be employed in buildings. As such applications of solar energy have to play a major role in covering the energy demand for heating and cooling in buildings. Energy consumption for air-conditioning and refrigeration During the last few decades the energy consumption for air-conditioning purposes has increased dramatically in most industrialized countries, even in heating-dominated climates. Total residential air-conditioners’ electricity consumption in EU-25 in year 2005 was estimated to be between 7–10 TWh (2.39e +13 BTU to 3.41e +13 BTU) per year /1.2/. This figure covers only small room air-conditioners and does not include the centralized air-conditioning plants or chilled water systems that are generally installed in large commercial buildings. The main reasons for the increasing energy demand for summer air-conditioning are the increased internal loads due to the increased

13

1 Introduction

use of electric and electronic equipment, increased living standards, occupant comfort demands and architectural characteristics and trends, such as an increasing ratio of transparent to opaque areas and to the growing popularity of glass buildings. All this generates the need for more active cooling. The share of commercial buildings in Europe equipped with cooling devices is expected to reach 60% of the stock by 2020. The European cooling potential for the total useful building stock is estimated as 1370 TWhc (4.67e+15 BTU) per year, of which 560 TWhc (1.91e +15 BTU) are due to the service sector and 810 TWhc (2.76e +15 BTU) to the residential sector /1.3/. Load reduction and passive concepts Three major steps can be taken to minimize the need for active cooling: (1) reduction of cooling loads by optimizing the building envelope and using energy efficient equipment, (2) using high efficient conversion technologies and auxiliary components and (3) make use of heat sinks in the environment such as outdoor air, water or the ground. Nowadays, the know-how on building design concepts leading to energy load reduction is broadly based and well developed. It is well established in practice, based on results from extensive research studies, using technology aimed at reducing cooling energy needs, for example, external shading devices, improved daylight concepts in combination with intelligent control of artificial lighting and energy-saving equipment. A further step has been taken towards the use of “cheap” cooling sources, for example, heat sinks such as the outdoor air for night cooling or evaporative cooling, radiative cooling using the cold sky and ground cooling using earth-to-air heat exchangers. However, in general the cooling capacity of these passive (or natural) cooling techniques is limited and cannot fulfill the cooling requirements of all building types, with different end user needs and climatic conditions. Solar thermal collector systems During the last thirty years and particularly in the past decade, growing environmental concerns and consistent effort in research and product development have opened up the market for active solar systems. Significant advances and improvements have been made, resulting in a wealth of experience with large installations using solar thermal collectors (examples are given in ­Figures 1.1 and 1.2). The solar thermal collector capacity in operation worldwide equalized 172.4 GWth (5.88e +11 BTU/h) corresponding to 246.2 million m² (2,650 million ft²)1 by the end of the year 2009 /1.4/. In spite of a significant and growing market penetration rate, the main obstacle preventing broad application of solar thermal collectors beyond their use in domestic hot water production has been the seasonal mismatch between heating demand and solar energy gains. Long-term storage units have to be employed in order to overcome this. The need for seasonal storage does not arise if solar thermal energy can be exploited for airconditioning of buildings during summer, i.e., sensible cooling and air dehumidification. The 1 0.7 kWth/m² (222 BTU/h*ft²) have been considered to derive the nominal capacity from the area of installed collectors.

14



great advantage for this kind of application is that the seasonal cooling loads coincide with high solar radiation availability. An example of the annual distribution of monthly heating and cooling loads and the available solar radiation is shown in Figure 1.3. The Figure shows cooling and heating loads (kWh per m² floor area) and the available solar radiation (kWh per m² collector area) (1 kWh/m² = 317 BTU/h*ft²) for a site in southern Europe. The exploitation of solar energy throughout the year, for both heating and cooling, thus has the potential to improve the performance output of a solar thermal installation and the economics of the investment and cover a larger part of the building energy needs with renewable energy at the same time

Fig. 1- 1 Flat-plate collector field for air-conditioning of a factory building in Inofita Viotias, Greece

Fig. 1- 2 Evacuated tube collector field for ­ air-conditioning of a wine cellar in Banyuls, France

Why use solar energy for cooling? The first demonstration of a solar-assisted absorption cooling machine was made during the Paris World Exhibition in 1878 by Augustin Mouchot, based on a technique developed by Edmond Carré /1.5/. Following intensive research and development activities in this field in the United States and Japan mainly during the 1980s (see e.g., /1.6/, /1.7/), there was a slow-down during the 1990s. However, recently this topic has attracted new interest in many countries. The reasons for this are manifold. On the one hand, there is an increased consciousness of the environmental problems which are created by the use of fossil fuels for generating electricity used by conventional cooling systems. In addition, common working fluids (refrigerants) exhibit a significant global warming potential. And in several countries the use of air-conditioning has an important impact in increasing the peak power demand and in some cases contributing to shortages in the electricity supply. This underlines the need to implement advanced, new concepts in building air-conditioning. In many countries and economic zones (e.g. Europe) policies for an increased use of renewable energies have been implemented. After the implementation of policies focusing on renewable electricity production and renewable heating today also measures on renewable cooling are taken into consideration. In this framework solar cooling becomes the most efficient technological option to renewable cooling in terms of primary energy, today available.

15

1 Introduction

On the other hand, the current conditions for the application of solar technology are much more favourable than in the past decades. Industrial production of solar components – solar thermal collectors as well as photovoltaic arrays – is well established in many countries. Reliable products with improved efficiency and competitive costs are available, which guarantee a reliable energy supply for many years without disproportionate maintenance costs. In addition, automated control technology has advanced and improved greatly due to new developments in electronics, which makes it feasible to handle even complex systems and installations efficiently. Subsidy programmes for application of solar technology in many countries document the political intention to increase the use of renewable energy sources as a substitute for fossil fuel consumption. Solar cooling is the logical consequence of these two developments. It provides a sensible synergy between a steadily increasing need for environmentally sound cooling concepts and a steadily increasing availability of high quality solar energy components and systems. The physical principle In principle, there are many different ways to convert solar energy into cooling or air-conditioning processes; an overview is given in Figure 1.4. A main distinction can be made between thermally and electrically operated systems. The latter can – at least in principle – be based on Peltier elements or vapour compression technology. Among the thermally driven processes, thermo-mechanical processes and processes based on heat transformation can be distinguished. The latter are all based on reversible thermo-chemical reactions with relatively low binding energies. These can again be classified in two categories. Open cycles are in contact with the atmosphere and always use water as the “refrigerant”. Closed cycles which are based on sorption technologies use sorption materials in liquid or solid form. While closed cycles provide only sensible cooling, i.e. they lower the temperature of a heat transfer medium (e.g. water or brine), open cycles treat both temperature and humidity of air in a single device. Thus an open sorptive cycle can be designed for direct treatment of ventilation air, while closed cycles need additional components such as cooling (condensing) coils to process latent loads.

Fig. 1- 3 Example of cooling and heating loads (kWh per m2 room area per month) and available solar radiation (kWh per m2 collector area per month) for a site in southern Europe. (1 kWh/m² = 317 BTU/h*ft²)



16



Solar electric and solar thermal cooling As mentioned, solar energy can be converted directly into electricity using photo-voltaic panels and the produced electric energy can be used to drive a vapour compression chiller with an electric motor. However, in industrialised countries, which have a well-developed electricity grid, the maximum use of photovoltaic has been (at present) achieved through feeding the produced electricity into the public grid without implementing any direct use within a building energy system. The competition between different types of solar technology, i.e., photovoltaic versus solar thermal systems, in the case of air-conditioning applications raises the question: is it more appropriate to install solar-thermally driven air-conditioning or to install a grid-connected PV power system and use a conventional air-conditioning system?



Fig. 1- 4 Conversion chains for the conversion of solar radiation energy into useful cooling

In countries such as Germany and Italy very attractive PV feed-in tariffs with a 20-years guarantee of the price of the electricity fed into the grid led to a strongly growing market for PV installations /1.8/. However, almost all these systems are simply connected to the grid and often do not interact with the building energy system. This may change in future with an increasing drop of the price for PV modules and corresponding reduced feed-in tariffs. As soon as the price for electricity fed into the grid gets lower than the price for electricity purchased from the utility it becomes interesting to make use locally of the electricity produced. Then electrically driven heating and cooling equipment such as vapour compression heat pumps, chillers or reversible heat pumps in connection with heat and/or cold storages will be interesting options for the energy supply in buildings. However, today only few complete system solutions using photovoltaics for energy supply in buildings are available on the market and therefore no solid information on the overall cost performance of such systems is available. Thus it is currently impossible to state that one of the two options is preferable in all conditions in economic and energy terms and as a consequence a careful analysis of the annual energy output and economic performance has to be carried out for each single case. This requires a comprehensive analysis of all energy needs within a building – electricity, heating, cooling, domestic hot water – and a detailed analysis on the possible contributions of either technology solution (photovoltaics or solar thermal).

17

1 Introduction

Technical scope This book deals with solar-thermally driven cooling technology. Moreover, it covers only mature forms of technology for installation, i.e., all necessary components are commercially available or at least in the phase of pilot applications. This also justifies the emphasis on centralised systems that provide conditioned air and/or chilled water to an entire building or certain areas of a building, since small (2–4 kW (0.568 to 1.136 ton)) thermally driven systems for decentralised appli­ cation in a single room are not yet available on the market. The general type of systems covered in this handbook is demonstrated in Figure 1.5. The system boundaries considered here are the incoming solar radiation on one side and the chilled water and/or conditioned air that is supplied to the building or the industrial process on the other side. This book does not address issues on how to reduce the building loads for air-conditioning or energy conservation; many textbooks and handbooks on this subjects are available (see e.g., /1.10/).



Fig. 1- 5 General scheme of systems covered in this handbook

Present technology – water chillers The dominating type of thermally driven cooling technology to produce chilled water is absorption cooling. Absorption chillers have been in commercial use for many years, mainly in combination with cogeneration plants or using industrial waste heat or district heat. For air-conditioning applications, absorption systems commonly use the water/lithium bromide working pair. Another closed-cycle sorption technology to produce chilled water uses the physical process of adsorption. This kind of chiller has a much lower market share in the overall cooling market at present even though there are numerous installations that use solar-thermally driven adsorption chillers. Present technology – desiccant cooling systems Another type of technology which has gained increasing attention over the last 30 years is desiccant cooling technology. Using this technology, air is conditioned directly, i.e., cooled and dehumidified. Desiccant cooling systems exploit the potential of sorption materials, such as silica gel, for air dehumidification in an open cooling cycle. This dehumidification effect is generally used for two purposes: to control the humidity of the ventilation air in air-handling units and – if possible – to reduce the supply temperature of ventilation air by evaporative cooling.

18



Target audience The goal of this handbook is to support planners in designing a solar-assisted cooling system which uses solar energy as a heat source and a thermally driven cooling / air-conditioning system to meet the load. Typical questions, which the book intends to answer, are: –– Is the use of solar-assisted air-conditioning feasible for a given building at a specific site? –– Which technology can be used? –– How can the best system for a given application under the conditions of the specific site be identified? –– Which solar collector types should be used for the selected air-conditioning system? –– What dimensions of the solar collector area and other system components result in the best energy cost-performance? –– Which tools are available that can support a user to further address these questions? Handbook structure This book consists of three main parts: part I (chapters 2 to 5) covers components, part II (chapters 6 to 9) describes systems and methods for design and performance assessment of solar-assisted cooling systems and part III summarizes lessons learned through practical applications (chapters 10 to 12). Chapter 2 provides an overview on meteorological data, heating and cooling loads and the load sub-system. In chapter 3 solar collectors, the main driving energy source in solar-assisted air-conditioning, are depicted. Hot water storages, being an important part of solar heat production systems, are also discussed, as well as other heat sources which may serve as a back-up. In chapters 4 and 5 main types of technology to transform heat into cold or conditioned air are described, i.e., thermally driven chillers (chapter 4) and desiccant cycles (chapter 5). Standard components such as vapour compression chillers and cooling towers are also addressed briefly, since they may play an important role in solar-assisted plants. In chapter 6 a step is taken from the component level to the system level. Different complete systems are compared and general operation strategies are discussed. A novel generic system approach is presented which helps to classify the different system solutions. In chapter 7 performance figures are defined, which allow the comparison of different technical solutions for solar-assisted air-conditioning with each other and the comparison with conventional systems that are used as a reference. Key figures to evaluate a system in terms of energy and cost performance are defined. In chapter 8 the design of complete systems is presented, whereby the main focus is placed on the design of the solar system in combination with the cooling system. Diverse design approaches are presented which provide different levels of detail, depending on the available information. In chapter 9 a more detailed view is made on the design of the solar thermal system used for heating and cooling applications. In chapter 10, 11 and 12 experiences from existing installations and their experimental analysis are presented. This includes an overview of the achieved performance and suggestions for a correct operation. Information about relevant technical and financial data is also provided. chapter 10 deals with small-scale pre-engineered systems, chapter 11 deals with custom made plants employing closed cycle technologies and chapter 12 covers desiccant systems.

19

1 Introduction

/1.1/

/1.2/

/1.3/ /1.4/

/1.5/

/1.6/ /1.7/ /1.8/

/1.9/

/1.10/

20

DIRECTIVE 2010/31/EU OF THE EUROPEAN PARLIAMENT AND OF THE COUNCIL of 19 May 2010 on the energy performance of buildings (recast). Brussels/Belgium. P. Bertoldi, B. Atanasiu, Institute for Environment and Sustainability (2007) – ­Electricity Consumption and Efficiency Trends in the Enlarged European Union – Status report 2006. ECOHEATCOOL b (2006), Work Package 2, The European Cold Market – Final Report, Euro Heat & Power. W. Weiss, F. Mauthner – Solar Heat Worldwide – Markets and Contribution to the Energy Supply 2009, Edition 2011 – IEA Solar Heating & Cooling Programme, May 2011 Mouchot, A.: “La chaleur Solaire et ses Applications Industrielles”. German translation “Die Sonnenwärme und ihre industriellen Anwendungen”, Olynthus Verlag, Oberbözberg, 1987. Löf G. – editor, (1993). Active Solar Systems, Volume 6 in Solar Heat Technologies series. (Bankston C.A. – editor) MIT Press, Cambridge, MA/USA. Sayigh A.A.M. and McVeigh J.C. – editors, (1992). Solar Air Conditioning and Refrigeration. Pergamon Press, Oxford/UK. EPIA (2011). “Solar Photovoltaics Competing in the Energy Sector – On the Road to Competitiveness”. Available: http://www.epia.org/index.php?id=18 [Access: 18 June 2012]. EPIA; Greenpeace (2011). “Solar Generation 6 Solar photovoltaic electricity empowering the world”. Available: http://www.epia.org/index.php?id=18 [Access: 18 June 2012]. Santamouris M. and Asimakopoulos D., – editors (1996). Passive Cooling of Buildings, James & James, London/UK.

Chapter 2

Meteorological data, heating and cooling loads and load sub-systems Responsible Author:

Livio Mazzarella, Politecnico di Milano, Italy Contributing Authors: Rien Rolloos, TNO, The Netherlands, Mario Motta, Politecnico di Milano, Italy

This chapter is devoted to the heating and cooling load calculations required for proper system sizing. An important consideration is that the dimensioning of a solar driven heating and/or cooling system is based on integrated energy usage, while the standard dimensioning of an HVAC (heating, ventilating, air conditioning) system is based on peak load conditions. The calculation procedure is therefore quite different; hourly based dynamic energy simulations should be used for solar thermal energy systems while design day based static heating and cooling calculations are typically and traditionally used for the latter (although a dynamic simulation and integrated energy sizing also plays an increasingly important role in HVAC design). It is evident that the required meteorological data is also quite different: hourly data for solar system sizing, design day peak values (static) and design day hourly profiles (dynamic) are required for traditional HVAC design. Modern simulation software for evaluating the performance of solar energy systems requires accurate and high-resolution meteorological data series /2.1/. Information concerning solar irradiation and meteorological variables – such as air temperature, relative humidity, and wind speed – are necessary for the analysis of renewable energy systems and to simulate the evaluation of heating and cooling loads in buildings and the performance of solar thermal systems.

2.1

Solar radiation, ambient temperature and humidity

The most important meteorological data needed to perform heating/cooling load calculations for solar thermal system simulation and sizing are: –– air temperature –– air humidity –– direct, or beam solar irradiance –– diffuse solar irradiance –– wind speed and direction

21

2  Meteorological data, heating and cooling loads and load sub-systems

Secondary data in the form of microclimatic data can also impact the design and sizing of a solar thermal energy system. Microclimatic data is normally derived from the main meteorological data via experimental and analytical relationships, which consider the local specific orography and building/system placement. Such secondary data are: –– sky temperature –– outdoor ambient temperature (outdoor operating temperature) All these data should be provided on at least an hourly basis (one data point per hour). Air temperature is the temperature of the outdoor air in degrees Celsius, θ (°C), recorded near the construction site or available from the nearest weather station. Air humidity is the water content of the outdoor air, recorded close to the construction site or available from the nearest weather station, and can be provided through different quantities, the most common of which are: –– relative humidity, φ , not dimensional (–) –– partial water vapour pressure, p in Pascal (Pa) –– wet bulb temperature, θwb, in (°C) –– humidity by mass, also denoted as humidity ratio, x (kg water vapor/kg dry air) Direct solar irradiance is the directional component of the solar radiation over the area unit of the receiving surface (flux). This component of irradiance is always directed according to the geometrical relationship between the sun, earth and surface, (the solar beam) and can be provided in terms of both normal direct solar irradiance (or normal beam, Gbn) or horizontal direct solar irradiance (or horizontal beam, Gbh). In the first case the receiving (measuring) surface is always perpendicular to the solar beam, in the second case the receiving (measuring) surface is fixed on the horizontal plane. The unit in both cases is W/m². The effective direct solar irradiance on a generic surface, Gb, depends on the relative position between the surface and the solar beam and it can be derived via “solar” trigonometric calculations from both normal direct or horizontal direct /2.2/, as:

Gb = Gbn ⋅ cos(ϑ ) = Gbh ⋅

cos(ϑ ) = Gbh ⋅ Rb cos(ϑ z )

Eq. 2- 1

where ϑ is the incident angle and ϑz the zenithal angle; both of which are functions of the time, latitude and surface azimuth and tilt angle over the horizon. Diffuse solar irradiance is the non-directional component of the solar radiation by area unit of the receiving surface, Gd. Diffuse irradiance is the compound of the total solar radiation coming from all possible directions with the exclusion of the solar beam direction. The diffuse solar ­radiation is also called the “indirect component” because it is not coming directly from the sun

22

2.1 Solar radiation, ambient temperature and humidity

but it is produced by atmosphere scattering and surface reflections. It is normally further divided into two main components: the diffuse sky solar irradiance, Gd,S, and the diffuse ground solar irradiance, Gd,G. The diffuse sky solar irradiance is the solar radiation scattered by the atmospheric components and conveyed to the receiving surface from any direction. The irradiance Gd,S depends on the relative position between the surface and the scattering medium, the “sky”: if the surface is exposed to the whole sky dome (i.e. horizontal surface) the irradiance is maximum. If instead it is exposed only in part to the sky, the irradiance value will be lower, until zero with no exposure to the sky. The diffuse ground solar irradiance is due to reflections of solar radiation from any surface surrounding the receiving surface. The irradiance Gd,G also depends on the relative position between the receiving surface and the reflecting surfaces in the same but complementary way to the sky component. Normally the available data are referred to as the horizontal diffuse solar irradiance, Gdh, which is the diffuse sky component for a horizontal receiving surface and represents the maximum possible amount of diffuse sky irradiance for a certain site at a specific time. The unit for horizontal diffuse solar irradiance is W/m² (1W/m²=0.317BTU/(h*ft²). A relationship among these quantities, based on the isotropic sky model, can be given as /2.2/;

Gd = Gd ,S + Gd ,G = Gdh ⋅ Fs ,S + Gd ,G ⋅ Fs ,G Eq. 2- 2

where Fs,S is the black body view factor between the surface and the sky dome and Fs,G is the black body view factor between the surface and the ground. Global solar irradiance is just the sum of the direct and the diffuse irradiance components, GT, ;

GT = Gb + Gd Eq. 2- 3

The last relation is very important when solar irradiances are measured through the use of two pyranometers, one with a shadow band and one without. When placed on a horizontal surface, the first measures the horizontal global irradiance, while the second measures the horizontal diffuse irradiance, i.e., the diffuse sky component (the direct solar component is blocked from the pyranometer by a shading device). Thus the normal direct solar irradiance can be derived by a simple calculation from the two directly measured quantities as:

23

2  Meteorological data, heating and cooling loads and load sub-systems

Gbn =

Gbh G − Gdh = Th cos(ϑ z ) cos(ϑ z )

Eq. 2- 4

Alternatively, the solar normal direct irradiance can be measured with a tracking pyroheliometer. All quantities are to be considered radiative total quantities, because spectral radiative quantities are not normally available in standard weather data. Wind speed is the horizontal component of wind velocity v (m/s), as measured at the reference condition – at 10 meters (32.8 ft) height and in open terrain (defined as an area where the horizontal distance between the anemometer and any obstruction is at least 10 times the height of the obstruction /2.3/). The wind direction D is expressed in degrees and is defined as the direction from which the wind blows, and is measured clockwise from geographical north, namely, true north. The wind speed is a function of the local environment (topography, ground roughness and nearby obstacles); then to have an estimation of wind conditions over a specific site from the reference wind speed vr, this speed shall be corrected following the relationship (assuming that there are no nearby obstacles):

v = v r ⋅ C R ⋅ CT Eq. 2- 5

where –– ––

C R CT

is the roughness coefficient is the topography coefficient

Values of roughness and topography coefficients can be found in various reference books, e.g. /2.4/. The sky temperature is the black body equivalent temperature of the atmosphere in Kelvin, Tsky (K), taking into account the cloud cover, i.e., the equivalent black body temperature which results in a radiative black body flow equal to the real longwave radiative flow from the atmosphere to the earth. The longwave irradiance from the atmosphere is normally related to different atmospheric parameters as the air temperature T (K) and the sky emissivity εa, which depends on dewpoint temperatures, θdp (°C), and on the cloud cover, c /2.5/. Outdoor ambient temperature, θoa (°C), is a compound heat transfer potential which groups together in one “linear” term the effect of both convection and longwave radiation heat transfer. It is the superficial heat transfer coefficient weighted average of the air temperature and the mean radiant ambient temperature, as follows:

24

2.1 Solar radiation, ambient temperature and humidity

θ oa =

hcv ⋅ θ + hrd ⋅ θ r hCR

Eq. 2- 6

with

hCR = hcv + hrd Eq. 2- 7

where –– –– –– –– ––

θ θ r hcv hrd hCR

air temperature mean radiant ambient temperature convective superficial heat transfer coefficient radiative superficial heat transfer coefficient combined convective-radiative superficial heat transfer coefficient

Using such potential the resulting heat flow is apparently linear in the temperature difference, as

φ = hCR ⋅ (θ s − θ oa ) Eq. 2- 8

where θs is the surface temperature. All nonlinearities are included in the superficial heat transfer coefficient, which is a function of the air temperature, wind speed and mean radiant ambient temperature. In some cases the combined convective-radiative superficial heat transfer coefficient may be taken as constant, simplifying the heat transfer calculations. The mean radiant ambient temperature, θr (°C), should not to be confused with the equivalent black body temperature of the outdoor ambient. It should be consistent with the outdoor ambient temperature definition, which is defined as the linear average radiant temperature of the outdoor ambient defined as the surface-to-surface radiative superficial heat transfer coefficient weighted average of superficial temperatures of any surface bounding the outdoor ambient (included the sky) /2.6/, that is:

θ r = ∑ j =1 N

hs , j ⋅ θ j hs

Eq. 2- 9

25

2  Meteorological data, heating and cooling loads and load sub-systems

with

hs = ∑ j =1 hs , j N

Eq. 2- 10

where θj temperature of surface j hs,j surface-to-surface radiative superficial heat transfer coefficient between the reference surface s and surface j hs radiative superficial heat transfer coefficient between the reference surface s and the whole ambient In some cases the surface-to-surface radiative superficial heat transfer coefficients may be simplified and assumed to be constant resulting in a simple estimation of the mean radiant ambient temperature.

2.1.1 Average quantities The available weather data is never provided as instantaneous values, but is always related to a measuring time interval, tm, and a recording time interval, tr,. Both affect the data quality. The measuring time interval depends on the measure type itself and on the data acquisition hardware. Typical reading time is 3 minutes or less (normally never less than 1 minute). Lower reading times are not necessary for weather data which do not have significant high time variability (wind speed is normally the most critical one). The recording time interval is on the other hand dependent on the storage space for data and on the future potential use of the recorded data. If a numerical model will be used to assess solar system performance with a time step of one hour, it is not generally necessary to have 1 minute of recorded data. For these reasons, weather data today are normally provided at best on an hourly basis. This makes it important to clarify the meaning of the available (recorded) data. On an hourly recording basis, the data reading is normally every 3 minutes, i.e., 20 readings in an hour. The recorded data is then an average value of such readings over the considered hour; this fact needs some more attention in relation to the time stamp. A time stamp of 01:00 means that the data are referred to the preceding time interval. Thus at 12:00 the average value is the average over the time interval 11:00–12:00, but marked 12:00. This should be taken into consideration when using such data with a performance simulation computer simulation; the weather data at

26

2.2 Availability of climatic data, sources of weather data

time 12:00 are not representative for the forward calculation time step 12:00–13:00. If no interpolation or adaptation is done on the weather data reading, one hour time shift will be introduced between weather data and other boundary conditions. This point may be even worse for solar irradiation. For fixed receiving surfaces, the quantity of interest is the global irradiance on the surface, calculated through its direct and diffuse components. As was discussed, the direct irradiance is calculated from the normal direct irradiance (which we may suppose is the available data) multiplied by the cosine of the incident angle (which is a function of time). This is true for irradiance, however not for its average value over a time interval. In this case the formula is; t + ∆t

1 ⌠ Gb = Gbn ⋅ cos(ϑ ) = ∆t  ⌡t

Gbn ⋅ cos(ϑ ) dτ ≠ Gbn ⋅ cos(ϑ )

Eq. 2- 11

Thus the true value differs from the product of the averages, which is the best approximation which can be used.

2.2

Availability of climatic data, sources of weather data

Commonly available weather data differ in the process they are collected and compiled, they differ in the amount and type of offered data, in the accessibility of data from a desired location, in accuracy, availability, and applicability. Consequently, the analysis of renewable energy solar systems can lead to quite different conclusions according to what kind of weather data has been used in an analysis. Therefore it is very important to know what characteristics the applied specific weather data set has in order to be able to correctly interpret the simulation results. Starting from the basic common feature, each data set has to allow for dynamical simulation, at least at hourly frequency. The most common hourly data formats can be used: –– multiyear hourly data (MYHD-type) –– average reference year (ARY-type) –– test reference year (TRY-type) –– design reference year (DRY-type) –– future weather year (FWY-type) The use of multiyear hourly data (MYHD-type) involves a very large data set compound of at least 30 years (in order to be representative) of historically ordinated hourly weather data. These are the source data for other data sets and are not normally used (if available) to avoid the additional and expensive computational efforts their utilization requires over this larger dataset. Using

27

2  Meteorological data, heating and cooling loads and load sub-systems

such a data set allows to correctly qualify the analyzed solar system because the weather data span over a long term taking in consideration both very favourable and unfavourable conditions. To speed up calculations, the average reference year (ARY-type) may be used. It is a single year long-term average measured data set created from many years of available data series (again 30 or more years), and is a feasible series but with the drawback that the values are averaged and are not able to directly account for the extreme meteorological values reached during the year. Typical periods of poor weather, which occur to a greater or lesser extent from year to year, are not explicitly represented or are not pronounced enough. Thus solar system simulated performances resulting from its use are not completely qualified and a system designed with such a data set may be unable to work correctly under extreme conditions (too unfavourable or too favourable ones). To overcome such drawbacks the test reference year data set (TRY-type) can be used. It is a single year of hourly data (8760 hours), selected to represent the range of weather patterns that would typically be found in a multi-year dataset, i.e., it is an average or typical year for a given location and time frame. Definition of a Reference Year depends on its satisfying a set of statistical tests relating it to the multi-year parent dataset from which it is drawn. Historically, two main different methods have been applied. The first method identifies a continuous, 12-month period as the typical year. The second method applies the criteria to individual months which are subsequently assembled into a composite 12-month set. In the first case, the most average year out of a set of many years of data is searched for among the multiyear data set to match the characteristics of an entire year in terms of the means and standard deviations of its monthly data to the average monthly values for many years of data. In the second case, the test reference year is constructed by examining the weather records for a series of past years, and then calculating how “typical” the months in each year are. In other words, the records for all Januarys are compared to each other, and the “most average” January is combined with the “most average” February, and so on until 12 typical months are assembled in a single 8760-hour record. As an example, the “most typical” January may be from 1962; February may be from 1976 and so forth. Mathematical smoothing is applied to the data at the beginning and end of each month to avoid abrupt changes in values. In this way a typical (i.e. reference) weather data set (TRY-type) is built rather than a data set based on long term averaged or extreme conditions. TRY-type data series are best suited for energy studies of solar systems allowing performance comparisons of different types, configurations, and locations. Design reference year (DRY-type) is a dataset with a sequence of 8670 hourly data values of meteorological variables selected to be representative of a very hot and/or very cold year, i.e. nearextreme data set. It is mainly used for peak power sizing of non-renewable energy systems and/ or heating/cooling system emitters and air handling units (AHUs), or to assess natural ventilation performances or building overheating and discomfort. Future weather year (FWY-type) is a dataset obtained through morphing (TRY-type) data sets or through the use of a stochastic weather generator, both trying to account for possible future climate change. These data sets can be very useful to forecast the possible overheating of actual low energy

28

2.3 Building space heating, domestic hot water and air conditioning needs

buildings in a global warming scenario and an increasing cooling demand or significant changes in solar system performances compared to those derived by the use of the (TRY-type) weather data.

2.3

Building space heating, domestic hot water and air conditioning needs

In the following section basic information on the physical, background and calculation methods of heating and cooling loads as well as domestic hot water profiles are provided.

2.3.1 Efficient building design practice Efficient buildings are not just the result of integrating one or more efficiency measure (insulation, highly efficient ventilation, etc.). It rather requires an integrated process ensuring advocacy and action on the part of the design team throughout the entire project development process. Nowadays this approach is often termed just with one word: holistic. With energy efficient or low-energy buildings, it is important to realize that the real purpose of the building is neither to save nor to use energy. The building has to serve the occupants and their activities. An understanding of building occupancy and activities can lead to building designs that not only save energy and reduce costs, but which also improve occupant comfort and workplace performance. An increased performance not necessarily has to result in increased construction costs. The main objective of the holistic approach is to invest in the building’s shape and enclosure (e.g. windows, walls, roofs) so that the heating, cooling, and lighting loads are reduced and smaller and less costly heating, ventilating, and air conditioning systems are required. Incorporating solar collectors and/or photovoltaic panels in the building envelope from the early design stage can lead to significant cost reduction and result in lower heat gains through the envelope itself. The first step in the design process begins with the assessment of the occupants’ needs, which are the primary goal. The second step is to site the building and arrange its spaces in a way to reduce energy use for heating, cooling, and lighting as much as possible, attaining and maintaining the occupants’ needs. The heating and cooling loads can be minimized by designing standard building elements; windows, walls, and roofs, so that they control, collect, and store the sun’s energy to optimum advantage. These passive solar design strategies also require that particular attention be paid to building orientation and glazing. Taken together, they form the basis of integrated, whole-building design, the holistic design.

29

2  Meteorological data, heating and cooling loads and load sub-systems

The third step is the correct design (without over sizing) and the efficient use of mechanical systems, equipment and controls. The systems design team has to give feed-back to the building design team because by incorporating building-integrated photovoltaic panels or thermal solar collectors into the facility, some conventional building envelope materials can be replaced by energy-producing technologies. For example, photovoltaic panels as well as solar thermal collectors can be integrated into window, wall, or roof assemblies, and spandrel glass, skylights, and roof become both part of the building skin and a source of power generation. Thus the holistic design procedure of an efficient building is an integrated process in the design phases. Energy-saving techniques, efficiency strategies and mechanisms to be deployed will vary greatly, depending on building and space typology. Their selection and configuration will be influenced by: –– Climate –– Internal heat gains from occupants and their activities, lights, and electrical equipment –– Building size and mass –– Illumination (lighting) requirements –– Hours of operation When the primary goal is building energy reduction and a solar system is employed to further reduce the use of non-renewable energy sources and related CO2 production, the importance of such influencing factors is increased. Due to the daily variability of solar energy, the time profile of the loads is needed to maximize the system efficiency and minimize the storage needs. If the building envelope thermal capacity and the user’s provided internal load are in such a way that the total load profile matches widely with the solar radiation availability, the solar system will cost less and perform better. Holistic building design as described above is particularly useful for new buildings; however, a similar approach can also be applied in building refurbishment. In fact most measures to reduce building energy consumption are more cost efficient than the installation and operation of large cooling and heating equipment.

2.3.2 Heating and cooling load: definitions and calculation methods Cooling loads and energy demand can be calculated using different approaches. In engineering practice CEN and ASHRAE calculation procedures are often used; when an efficient holistic building design or analysis involving a solar driven energy system are implemented, hourly simulations should be used. Regardless of the methodology employed the first step in buildings heating and cooling analyses is the determination of heat transfer through the envelope.

30

2.3 Building space heating, domestic hot water and air conditioning needs

2.3.2.1 Heat and mass transfer through the envelope Heat transfer through the building envelope occurs when energy flows from outside to inside the building or vice versa, depending on the weather and internal conditions. System loads can occur depending on the ongoing boundary conditions both during winter and summer. For consistency purposes a sign convention has to be fixed on the energy flow through the envelope. In the following discussion it will be positive if incoming to the building, negative if outgoing flow from the building. The heat transfer through the envelope (walls, windows, roof and basement), measured at the inner envelope surface, has the following components: –– unsteady heat transfer through opaque building envelope component, ΦT,op –– unsteady heat transfer through transparent building envelope component, ΦT,tr –– heat transfer by air exchange between surroundings and building because of infiltration and direct natural ventilation through window and envelope openings, ΦV The heat transfer through the opaque building envelop depends on outdoor and indoor ambient temperatures, absorbed solar radiation (wall orientation and wall surface color), thermal conductivity and thermal capacity of the wall materials. In modern buildings the envelope elements are often quite well insulated, thus heat transfer through opaque elements is in most cases small irrespective of the thermal capacity. Due to the thermal capacity effect, the heat flux at the outer envelope surface differs from that at the inner surface at any given time. Contemporary buildings are also sufficiently tight to prevent significant infiltration of ambient air into the building. Nevertheless they must be ventilated to ensure good indoor air quality. Mechanical ventilation must be controlled according to demand to ensure lower heating or cooling load with supply air. The total heat transfer through the building envelope, ΦT,t , measured at the inner envelope surface, is given by:

Φ T, t = Φ T, op + Φ T, tr + Φ V Eq. 2- 12

If there is infiltration or direct natural ventilation through windows or envelope openings, the incoming outdoor air is a moist air and therefore brings in a certain amount of water vapour, which is normally quantified through the humidity ratio x (kg or g of water vapour over kg of dry air). But, if outdoor air is blown in, some indoor air has to flow out, leading to a water vapour flow through the building. To properly account for this phenomenon, this transfer of water vapour should be considered in the calculations. Water vapour transfer through the building envelope together with the heat transfer by infiltration/natural ventilation is coupled and is considered by moist air enthalpy flows.

31

2  Meteorological data, heating and cooling loads and load sub-systems

Since this coupling is weak, a separation between the two phenomena can be made, calculating the heat transfer due to infiltration/natural ventilation, ΦV , as a net dry air enthalpy flow, and calculating the related moisture transfer as a net water vapour enthalpy flow, Φh,wv,f .

2.3.2.2 Heat and mass gains Heat gains by definition are always positive (from outside to inside or inside “generated”) and are divided into sensible and latent heat gains. Sensible heat gains are originated by: –– solar radiation and heat transfer through windows, ΦS&H –– internal heat gains (occupants, lighting, appliances, ...), ΦI Heat gains through windows and transparent walls can be characterized by several optical parameters: –– solar radiation transmittance τ –– solar radiation absorptance α –– fraction of the absorbed solar radiation transferred to the inside (the secondary heat transfer factor of the glazing towards the inside) qi –– total solar energy transmittance g –– shading factor of shading devices Sf Solar radiation transmittance τ is the ratio between transmitted and incoming solar radiation as electromagnetic energy which will be converted into heat gains when released by the irradiated walls, floor and furniture. Since part of solar radiation is absorbed in glazing, radiation and convection heat fluxes from the inner glass layer into the interior represent additional heat gains. The sum of both leads to the total solar energy transmittance g:

g = τ + qi Eq. 2- 13

The g-value is the most adequate parameter to characterize windows, but for dynamical heating and cooling load determination it is better to maintain the two separate components. Cooling loads through the transparent building envelope component could be significantly reduced by selection of effective shading devices, which ideally allow radiation through the windows in a heating season to maximize the passive gains in that season. Taking into account the shading devices effect by the shading factor S_f, an effective total solar energy transmittance can be defined and used as:

g ef = S f ⋅ (τ + qi ) = S f ⋅ τ + S f ⋅ qi Eq. 2- 14

32

2.3 Building space heating, domestic hot water and air conditioning needs

According to the different heat transfer phenomena solar radiation and heat transfer though the window is broken down into;

[

]

Φ S&H = Φ S + Φ S,th = (S f ⋅ τ )T + (S f ⋅ qi )T ⋅ Aw ⋅ GT Eq. 2- 15

where ΦS is the solar radiation entering through the window, ΦS,th is the heat flux due to the glass overheating due to solar radiation absorption, Aw is the transparent window area and GT is the total solar irradiance on the windows surface (please note: the transmission coefficients are calculated for the total irradiance as function of the incident angle and thus of time). When dynamic performance is important as in the case of solar systems, only the thermal component of the solar energy flux through the window, ΦS,th , is a direct heat gain. The transmitted solar radiation ΦS, before becoming a heat gain, has to be absorbed by any material surface and then converted, after a time lag, to heat flux delivered to the room. Thus indicating with ΦS,TL such a component which is a function of ΦS and of the thermal and radiative properties of the room and which exhibits a time lag with respect to the incoming solar radiation, the time dependent solar gain, ΦS,G , can be expressed as:

Φ S,G = Φ S,TL + Φ S,th Eq. 2- 16

Internal heat gains are often a reason for overheating. The human body itself emits a heat flux from approximately 100W (341 BTU/h) to 250W (853 BTU/h) in a condition of heavy activity. Large numbers of appliances, as is usual in commercial buildings, contribute to large internal heat gains as well. Good daylighting design and use of high efficient compact and LED lamps can significantly reduce the internal cooling loads. If the internal gains are indicated by ΦI the total sensible gains ΦG,s are given by:

Φ G,s = Φ S,G + Φ I Eq. 2- 17

Latent heat gains are in general generated in buildings because of different water vapour sources and by moist air infiltration. For example, a human body emits up to 50g (772gr) of water vapour per hour, plants up to 20g (309gr) per day. Thus they correspond to a water vapour enthalpy rate due to water vapour sources, Φh,wv,s.

33

2  Meteorological data, heating and cooling loads and load sub-systems

2.3.2.2 Building heating and cooling loads The sensible heating load in building can be defined as the heat needed to fulfil the requirements of thermal comfort, especially regarding indoor air temperature, or, more precisely, the operative temperature when it is a provided rate of energy; while building sensible cooling loads indicate heat when it is a removed rate of energy. Thus, using the present sign convention, which can be unified for both in the term thermal load and the symbol ΦSL, which will be positive if cooling load and negative if heating load. This generalized building thermal load is then given by the balance between the heat transfer through the envelope and the sensible gains as:

Φ SL = Φ T, t + Φ G,s Eq. 2- 18

The thermal load is time dependent on the outside forcing conditions (weather) and the inside forcing conditions (occupancy, activities, etc.); it also depends on the indoor air temperature which can vary, or free float, if there is no heating or cooling system trying to balance the load itself. Assuming that it is possible to split sensible and latent loads, as we already have done, it is possible to write the following integral thermal balance equation:

C ai

d Tai (t) dt

= Φ SL (t) + Φ SYS,S (t)

Eq. 2- 19

The term on the left side represents the rate of change of internal air energy due to changing of the air temperature (Cai is the indoor air thermal capacity), and the new term on the right side, ΦSYS,S, is the provided or removed heat flow when trying to balance the thermal load by the system. It is clear that if there is no system, the volume average internal air temperature is going to follow the load; increasing if a cooling load exists, decreasing if a heating load exists. If we assume that a perfect control system exists, the average internal air temperature can be considered constant over time t regardless of the time dependence of the thermal load. Thus in this case the system is always compensating the load and this can be stated as:

Φ SYS,S ( t) = −Φ SL (t) Eq. 2- 20

The heat flux required by the system to constantly maintain the temperature is positive when the system is in heating mode, negative when in cooling mode; we call this term the system thermal load.

34

2.3 Building space heating, domestic hot water and air conditioning needs

Normally a building performance simulation computer program employs both equations, the former to account for time periods when neither mechanical heating nor cooling exists (free floating regime), the latter to calculate the maximum heating and/or cooling demand to size the system (design heating and cooling loads). If a performance assessment is required, the system is virtually already sized and the differential equation will be used together with a control strategy. This can be, for example a simple on/off controller, switching the system on and off allowing the temperature to swing across the control dead band, or a modulating controller, maintaining the temperature more constant and running the system at variable power. In this case, the time dependence of the thermal system load is important because this is what the solar system has to compensate with its renewable energy source. The building latent load indicates the water vapour enthalpy that has to be extracted or injected if a comfort condition has to be achieved and maintained with respect to air humidity. It is given by the algebraic sum of the net water vapour enthalpy flow, Φh,wv,f due to infiltration or direct natural ventilation plus the water vapour enthalpy grow rate due to water vapour sources, Φh,wv,s , that is;

Φ LL = Φ h, wv,f + Φ h, wv,s Eq. 2- 21

It is mainly dependent on the water vapour sources inside the building, especially if the building is very tight as in modern constructions. The water vapour energy balance can be written in general integral form as:

ρda ⋅ V ⋅ hwv

d xai (t) = Φ LL (t) + Φ SYS, L (t) dt

Eq. 2- 22

where the term on the left side represents the rate of change of air energy due to a change in water vapour content (ρda is the indoor dry air mass density, V the volume and h wv a reference water vapour specific enthalpy), and the new term on the right side, ΦSYS,L , is the system provided or removed water vapour enthalpy flow when trying to balance the latent load. If there is no system, the volume average internal humidity ratio is going to follow the load: increasing if a positive latent load exists, decreasing if a negative latent load exists. If it is assumed that a perfect control system exists, the volume average internal humidity ratio can be taken as constant to time regardless of the time dependence of the latent load. Thus in this case the system is always compensating the load and this can be stated as:

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2  Meteorological data, heating and cooling loads and load sub-systems

Φ SYS, L (t) = −Φ LL (t) Eq. 2- 23

The water vapour enthalpy flow required by the system to maintain the indoor air humidity constant is positive when the building latent load is negative, and vice versa. This term is called the system latent load. Assuming the same reference value, h wv , for the water vapour specific enthalpy in Φh,wv,f , Φh,wv,s , and in ΦSYS,L , the previous energy balance is equivalent to the following mass balance, for a . unique infiltration mass flow, m da(t), in terms of dry air, and a total water vapour mass source, . M wv(t):

ρda ⋅ V

d xai (t) dt

= m da (t) ⋅ (xin (t) − xai (t)) + M& wv (t) + m da,SYS (t) ⋅ (xin,SYS (t) − xai (t)) .

.

Eq. 2- 24

Assuming that a perfect control system exists, the system is always compensating the load which can be stated in terms of masses:

m da,SYS (t) ⋅ (xin,SYS (t) − xai (t)) = − m da (t) ⋅ (xin (t) − xai (t)) − M& wv (t) .

.

Eq. 2- 25

or in terms of humidity ratio, assuming as usual to include the net water vapour flow due to air infiltration in an equivalent water vapour mass source, as:

M& wv,eq = + m da (t) ⋅ (xin (t) − xai (t)) + M& wv (t) .

Eq. 2- 26

It can be written as

xin,SYS (t) = xai (t) − Eq. 2- 27

36

M& wv,eq (t) .

m da,SYS (t)

2.3 Building space heating, domestic hot water and air conditioning needs

which yields the needed information: that is the humidity ratio of the air stream supplied to the building for compensating the latent load. The quantity calculated in this way is the so called design latent load or design system provided humidity ratio. If, on the other hand, a performance assessment is required (the system is already sized), the differential equation has to be used together with a control strategy algorithm to calculate the actual system load. Even in this case, the time dependence of the system latent load is the goal, because this is what the solar system has to compensate for.

2.3.2.3 HVAC heating and cooling loads When system thermal and latent loads are known (as design values or as time profiles for a sized virtual system when we perform a performance simulation), the heating and ventilation air conditioning unit (also named Air Handling Unit, AHU) loads can be calculated or assessed and finally the energy profile provided by the solar system can be evaluated. The AHU loads are not equal to the system loads because of the mechanical ventilation loads, both sensible and latent. To provide the system loads, sensible and latent, treated external air is supplied to the building at the required temperature and humidity ratio, the AHU has to perform such thermodynamic treatment from the outdoor air conditions to the indoor conditions. Thus, it is normal that, depending on the amount of fresh outdoor air and its temperature and humidity, the heating and cooling power required by the AHU can vary and in some cases be higher than the sum of the thermal and latent system heat loads. The extent of these differences will depend on the AHU typology, on what kind of thermodynamic processes are employed, the coils efficiencies, etc. To be able to analyze and asses the solar driven system performance – systems which can be integrated into the AHU like DEC technologies or just be a cold or hot generator for a heat carrier fluid – the full design of the AHU and the whole HVAC system has to be carried out. Then, using a computer simulation tool like TRNSYS /2.7/, EnergyPlus /2.8/, esp-r /2.9/ or others, employing the right meteorological data set for specific locations and the planned goal (TRY-type or DRY-type or FWY), performance assessment of the solar system can be carried out in a holistic way. This leads to synergies between building and other system components, e.g. compression or absorption chillers or heat pumps, and the users’ habits and complaints to minimize both CO2 production and life-time system cost. Some references and application examples for heating and cooling of building are found in /2.10, 2.11, 2.12, 2.13/.

2.3.3 Domestic hot water load profiles For solar thermal systems which have to provide only domestic hot water (DHW), or for assessing the performance of a combined DHW and heating or/and cooling solar system, the domestic hot water profile is needed.

37

2  Meteorological data, heating and cooling loads and load sub-systems

The needed data is: –– the hourly profile of mass flow rate –– the delivered water temperature –– the required storage temperature –– the employed control strategy With all this data it is possible to calculate the required hourly thermal power to the solar system: .

Φ DHW (t) = m DHW (t) ⋅ cW ⋅ (θ dev − θ0 ) Eq. 2- 28

where cW is the specific thermal capacity by mass of liquid water and θdev is the domestic water main distribution temperature, assumed constant. Again using a computer simulation program and the right weather date set for the purpose, it is possible to assess the performance and to optimize the system. The required mass flow rate profile is strongly dependent on the specific application, while the delivered water temperature is less variable, but from the solar system point of view is important to know or to decide if there is a temperature booster (electrical or fuel boiler) or not to reach the required temperature. That is a matter of cost/energy optimal solution. Some information on the DHW demand can be found in /2.14/.

2.4

Industrial heating and cooling

Industrial heating and cooling is a field where solar energy can play an important role. To be able to correctly design a process solar system several steps have to be followed which can be shortly summarized as: Preliminary analysis This analysis includes the following main steps: analysis of building structure and surrounding area, analysis of process characteristics and integration of solar heat/cold in distribution networks. System design The detailed system design is made of the following main steps: calculation of thermal load profile, pre-dimensioning of collector field and storage(s), selection of collector type and system simulations. The main focus of this paragraph is to underline which kind of data have to be collected to correctly design the solar system in the preliminary steps up to definition of the thermal load profile, which is the main boundary condition together with the weather data.

38

2.4 Industrial heating and cooling

2.4.1 Preliminary analysis required data Prior to the system design phase, preliminary analysis of the existing conditions is needed in order to assess the potential benefits for the considered industrial process. The boundary conditions include the actual building structure and surrounding area, the process characteristics (schedule and thermal power), the heat and cold distribution network. As a general rule, a process suitable for solar thermal heating and cooling should meet the following general criteria: –– thermal energy demand throughout the year –– demand for 5 days a week minimum –– similar demand between summer and winter, with preferable larger demand in summer. Thus a collection of information on any process heat or cold use during the whole year is a fundamental requirement in the pre-design phase and in particular in order to assess if it makes any sense to install a solar system at all. Other important prerequisites are those related to the following: Analysis of building structure and surrounding area This step requires collecting information about: –– available unshaded area (roof and ground) –– roof materials and construction technique for structural static assessment –– available area for heat and cool storage(s) –– length of piping from collector field to storage and from storage to process Analysis of process characteristics This step requires information about: –– type of process (open, closed, with or without heat and mass recovery) –– process scheduling (continuous, discontinuous, seasonal dependency) –– temperature levels (supply and return temperature of heat transfer media) –– energy demand, at least on a seasonal basis Integration of solar heat / cooling in distribution network This step requires information about: –– temperature levels in the distribution network –– efficiency of the actual generating system –– parallel or series connection of solar heat/cold with respect to actual generating system)

39

2  Meteorological data, heating and cooling loads and load sub-systems

2.4.2 System design data: thermal load profile To be able to perform a detailed system design, an estimation, measurement or calculation of the thermal load profile has to be done. For solar thermal heating and cooling system design, the thermal load profiles should be known on a daily, weekly and annual basis. Since it is usually rather difficult to obtain such information through interviews with the process planners or the staff working at the plant, it is recommended to measure the mass flows and inlet- and outlet-temperatures of each process on each typical working day. After deciding which processes should be supported and how the solar heat/cold should be integrated, the thermal load should be determined in terms of fluid mass flow and temperature at the integration point, hour by hour, during each typical day. Depending on the process characteristics, which have been assessed during the pre-design phase, measurements of mass flow rate and temperature may be taken as given or must be made for a process characteristic time interval (hours, days, weeks), if not available from processes design, management or maintenance data. While mass flow rates are similar in each specific process plant, the temperature range of required heat or cold is more process specific. In the following a delivered temperature range for typical processes is given.

2.4.1.1 Low and medium temperature heat processes The list of industrial processes is listed below, classified by industry field and including usage temperature range (useful temperature level, the one of the process itself):

Food and beverages Drying

(30–90 °C) (86–194 °F)

Washing

(40–80 °C) (104–176 °F)

Pasteurizing

(80–110 °C) (176–230 °F)

Boiling

(95–105 °C) (203–221 °F)

Sterilizing

(140–150 °C) (284–302 °F)

Textile industry Washing

(40 –80 °C) (104–176 °F)

Bleaching

(60–100 °C) (140–212 °F)

Dyeing

(100–160 °C) (212–320 °F)

Chemical industry Boiling

(95–105 °C) (203–221 °F)

Distilling

(110–300 °C) (230–572 °F)

Chemical processes

(120–180 °C) (248–356 °F)

40

2.5 The load sub-system – air-conditioning equipment

All sectors Pre-heating of boiler feed water

(100 °C) (212 °F)

Heating of production halls

(30–80 °C) (86–176 °F)

2.4.1.2 Low and medium temperature refrigeration processes Food and beverages –– Food conservation ( –20 – +5 °C) (–4 – +41 °F) All sectors –– Cooling of production halls (20–30 °C) (68–86 °F) Some information on possible applications and temperature range are also reported in /2.15/.

2.5

The load sub-system – air-conditioning equipment

Air-conditioning, or indoor climate control, is the process of treating air so as to control simultaneously its temperature, humidity, cleanliness and distribution. The designer is responsible for considering various systems and recommending the one that will best provide the desired air treatment. It is imperative that the designer defines the goals of the design according to the user needs and project constraints. Before starting any planning activity for air-conditioning systems – for solar-driven systems in particular – it is always important to make sure that the estimated building loads are reasonable. It is always more economical and rational to apply energy-saving measures – at almost any planning stage – to achieve reasonable loads instead of over-dimensioning the plant in order to meet excessive loads. Once the main design constraints, such as use of the building, cooling loads, occupancy schedule, required ventilation rates and zoning requirements are clear, the air-conditioning design process can be initiated. In some cases, for example low-energy buildings, this will be the first step of an iterative process, which will lead to the optimization of the building envelope and installations. Given the project constraints, the designer will have to determine the feasible technology options that will fulfil the user’s requirements. Air-conditioning systems fall into one of four major categories: all-air, air-water, water and refrigerant-based systems. Each type has certain technical and economic advantages. Some are better than others for specific applications. This identification is based on the controllable fluids used in each zone for the appropriate control of the air conditions. Each system or combination of systems has specific capabilities concerning cooling capacity.

41

2  Meteorological data, heating and cooling loads and load sub-systems

The refrigerant-based systems have a distinct feature compared to the other three groups. They are normally unitary systems, meaning that the “cooling” process takes place at a very short distance from the delivery terminals; these systems are mainly window-mounted or wall-mounted. In the other, centralized systems, the refrigeration units may be located at some distance from the conditioned space, for instance in a central mechanical equipment room. The distribution system for the cold medium (supply and return) connects them to the delivery terminals. Only centralized solar-assisted air-conditioning systems are currently available on the market, although there are research efforts towards unitary systems. For the purpose of this book, only centralized systems will be considered further. The generic classification for centralized systems is shown in Figure 2.1.

Fig. 2- 1 Generic classification of centralised air-conditioning systems

42

2.5 The load sub-system – air-conditioning equipment

Table 2.1 lists specific cooling capacities of the different air-conditioning systems.

Tab. 2- 1 Specific (maximum possible) cooling capacities (W/m²) of different air-conditioning systems. For cooling loads higher than 45 W/m² choose: air cooling based on minimum required fresh air quantity (e.g., 30 to 50 /h per person) and secondary cooling (water-based); e.g,. system C or A + D. The aim is to save costs for ducting and energy for the transportation of air and to avoid possible draughts when introducing too much air into the room. Possible draught can be avoided by using, for example, high-induction air outlets in the wall or swirl outlets in a ceiling. 1 W/m² is equal to 0.317 BTU/(h*ft²) *) At higher ventilation rates  **) Values not known

2.5.1 All-air systems An all-air system provides complete sensible and latent cooling, and possibly humidification of the supplied air. No additional cooling or dehumidification is required in the zone. The same air stream is normally used for heating purposes. Air is delivered to the conditioned space through air ducts and distributed through air outlets (i.e., grilles, slot diffusers, ceiling diffusers, perforated ceiling panels, variable diffusers etc.) or mixing terminal outlets (mixing boxes). Indoor air is then usually returned to the main air-handling unit, where heat/energy is recovered, or exhausted to the outside. All-air systems may be adapted to many applications for comfort or process work. They are often used in buildings that require individual control of multiple zones, such as office buildings, schools, hospitals, laboratories and hotels.

43

2  Meteorological data, heating and cooling loads and load sub-systems

The advantages of all-air systems are that they have the greatest potential for use of outside air for “free” cooling, and they provide a wide choice of zoning, flexibility and humidity control under all operating conditions, with possible simultaneous heating and cooling, even during off-season periods. Heat-recovery systems can be easily incorporated into the main air-handling units, thus resulting in considerable energy savings. By adjusting the air-flow volume, they are able to maintain practically constant indoor space conditions within a very small tolerance. Their disadvantages include the requirement for additional duct space within the building for air distribution and their higher installation and operation costs. Their installation and design require closer co-operation between architectural, mechanical and structural designers. All-air distribution systems do not pose specific problems if applied for solar-assisted air-conditioning, and no components need to be specially adapted. Therefore the reader can refer to existing specific texts for more thorough treatment (e.g. /2.16/). The supply-air temperature and speed delivered to indoor spaces have to be carefully controlled in order to avoid thermal discomfort and draught problems. The maximum possible temperature difference between the supply and indoor air depends on the height of the indoor space, actually on the height of the mixing zone above the occupation zone (1.8 m above floor level). Furthermore, the way of introducing the air influences the temperature difference. From practical experience (at least in northern European countries), it has been documented that air-cooling complaints related to draught are minimized when the temperature difference between exhaust air and supply air, ΔT, is not greater than the values listed in Table 2.2. Typical capacity values for such systems are shown in Table 2.1.



Tab. 2- 2 Maximum temperature difference for supply of cooled air (1 °C = 33.8 °F)

All-air systems can be classified into five groups: single-duct systems in the versions variable air volume (VAV), or constant air volume (CAV), single-duct displacement systems, dual-duct systems and multizone systems. In general, single-duct systems consume less energy than dual-duct systems, and VAV systems are more energy-efficient than CAV systems. Energy waste due to leakage in ducting and terminal devices can be considerable, reaching values up to 20% /2.16/. Heat-recovery devices can save considerable amounts of energy and reduce the required capacity of primary cooling and heating. A brief overview of the most important all-air type systems is presented in the following sections.

44

2.5 The load sub-system – air-conditioning equipment

2.5.1.1 Single-duct systems: variable air volume (VAV) and constant air volume (CAV) system This system supplies a single stream of conditioned air (either warm or cold air) from the airhandling unit to the indoor space. Single-duct CAV systems change the supply air temperature and humidity in response to the space load, while maintaining a constant air-flow. The VAV system controls the indoor air temperature by varying the quantity of the supplied air rather than varying the supply air temperature. The supply air temperature is held relatively constant, depending on the season. The greatest energy savings associated with VAV systems occur in the building’s perimeter zones, where variations in solar load and outdoor temperature permit a reduction of the supplied air quantities.

2.5.1.2 Single-duct displacement systems This system supplies the conditioned air near floor level at a low speed. For cooling, the air-supply temperature is some degrees below the indoor air temperature. The supplied air spreads out over the floor. Indoor air movement is enhanced – from bottom to top – because of convective heat gains (i.e., from occupants, equipment, etc.). Cooling capacity is restricted to about 30 W/m² (9.51 Btu/(h*ft²)) at a room height of 2.7 m. Air movement may also be enhanced by the removal of the exhaust indoor air. In clean rooms and similar zones, pure displacement ventilation can be applied with high ventilation rates and high specific cooling capacities.

2.5.1.3 Dual-duct systems This system supplies two streams of air at different conditions (one warm air stream and one cold air stream) from the air-handling unit to the indoor spaces. The two ducts run in parallel and deliver air to a mixing box. The warm and cold air streams are mixed in correct proportions to satisfy the specific space loads at each conditioned space or zone. Dual-duct systems use more energy than single-duct VAV systems and require even more space for the ducts. Note that these systems may be designed as constant air volume or variable air volume systems.

2.5.1.4 Multizone systems This system supplies several zones from a single, centrally located air-handling unit. Different zone requirements are met by mixing cold and warm air through zone dampers at the central airhandling unit in response to zone thermostats. The mixed, conditioned air is distributed throughout the building by a network of single-zone ducts. This is a common system in humid climates; it often produces air at a temperature of 9–11 °C (48.2 °F – 51.8 °F) for humidity control. Supply air is then mixed with a return stream to maintain comfort conditions in different zones. Heat is only added if the zone served cannot be maintained by delivery of return air.

45

2  Meteorological data, heating and cooling loads and load sub-systems

2.5.2 Water systems Water systems heat and/or cool a space by direct heat transfer between water and the indoor air. They provide a means to cover indoor cooling (or heating) loads but cannot satisfy the ventilation loads. Therefore, fresh air has to be provided by other means, for example either by infiltration in naturally ventilated buildings or by other local fresh/exhaust air exchange or even a central, mechanical ventilation system. Depending on the type of chilled-water air-conditioning terminal unit installed in the space, it may be possible to handle latent loads. The most common terminal units that are used with chilled-water systems include fan-coils, chilled ceilings, chilled floors or systems using natural convection in combination with cooling coils (silent cooling, chilled beams). A brief overview of the most common terminal units is presented in the following sections.

2.5.2.1 Fan-coils A fan-coil system is a heat exchanger (coil) with a fan that simply circulates indoor air over it, housed in the same unit. The heat exchanger is supplied with hot or chilled water. The units have a thermostatically controlled built-in fan that draws air from the room and then blows it over finned tubes of the heat exchangers where hot water or steam for heating or chilled water for cooling is circulated. The hot or cold medium is centrally produced either from equipment located in the building or supplied by a district heating/cooling network. A sketch of a simple fan-coil unit is shown in Figure 2.2.

Fig. 2- 2 Cross-section of a typical simple fan-coil unit with one heat exchanger for air heating/cooling

46

2.5 The load sub-system – air-conditioning equipment

The system can be controlled by using a simple thermostat or a more complex electronic microprocessor control system (three-way valves, variable speed fan, etc.). Fan-coils are examples of single-zone systems. They are available in horizontal ceiling-mounted, concealed, or recessed vertical floor units. Two types of fan-coils are commonly available, namely two- or four-pipe systems. –– Two-pipe system: A two pipe unit uses one pipe for the supply and one pipe for the return of the hot/cold medium to the heat exchanger. The unit can then be used for either heating or cooling the indoor air, depending on the central system operation mode. Multiple chillers and/or boilers may be required for multiple zones. Unitmounted controls utilize a pipe-mounted sensor; it is used to determine the operation mode, i.e., heating or cooling, and the unit is controlled accordingly. –– Four-pipe-system: A four-pipe unit is equipped with two independent coils, one for heating and one for cooling. Cooling and heating valves for controlling coil capacities are often factory-installed, and their control devices are hidden inside the unit’s cabinet, or they are wall-mounted, or remotely mounted. Beside these systems, also two-pipe systems with an electric heater are available. Standard operating conditions for the use of fan-coils are for instance defined in Eurovent Standard 6/3/ /2.17/. Accordingly, in the cooling mode, the standard defines the water supply/return temperatures at 7/12  °C (44.6/53.6 °F) for an indoor air temperature of 27 °C (80.6 °F) (relative humidity 50%). In the heating mode, the standard defines the water supply/return temperature at 70/60 °C (158/140 °F) for a 4-pipe system and 50/40 °C (122/104 °F) for a 2-pipe system, for an indoor air temperature of 20 °C (68 °F). The main advantages of fan-coils are: –– The system requires only piping installation, which takes up less space than air ducts in all-air systems. –– Unoccupied building spaces may be separated by simply turning off the local fan-coils or diverting the cold/hot medium flow to the fan-coil. –– Zones can be individually controlled and managed with a centralized control unit. The main disadvantages are: –– Condensate must be removed from each unit –– Interior zones may require additional fresh-air ventilation with separate ducts –– Heat-recovery may be more difficult to achieve –– Potentially noisy system since the air fans are located inside occupied areas For the purpose of this handbook it is of particular interest to consider the energy performance of typical fan-coil units. Based on information from commercially available fan-coils, typical cooling capacities of a single fan-coil can vary between 0.5 kW and 10 kW. In general, the heating capacity is around 2 times higher than the cooling capacity. The typical air-flow is in the range of

47

2  Meteorological data, heating and cooling loads and load sub-systems

5.5 to 6.5 /h per kW of cooling capacity A fan-coil can be operated in a range of around 10% to 100% of the full cooling capacity by controlling the air-flow. The electric consumption for the fan is in the range of 3% to 7% of the cooling capacity if the fan-coil runs at the lowest air-flow and in the range of 1% to 2.5% of the cooling capacity if the fan runs at its maximum. Most fan-coils are equipped with a device to dispose condensate in those cases where control of indoor humidity is possible. Condensation occurs, depending on indoor conditions, when the air is cooled below its dew-point.

2.5.2.2 Chilled ceilings A cooling system which is based on chilled ceilings and ventilation separates latent and sensible cooling. The main part of the sensible cooling is delivered to the room by the chilled ceiling. Cooling is transferred to the room via radiation and/or convection. The relative proportion of each heat transfer mechanism depends on the design of the chilled ceiling. A ceiling with a closed surface involves primarily radiation for the heat transfer and is referred to as a radiative ceiling, whereas a convective ceiling presents a more “open” structure. A closed loop circulates water through the ceiling. The water inlet temperature is usually between 16 °C (60.8 °F) to 18 °C (64.4 °F). Changing the water flow-rate and/or the water inlet temperature controls the room temperature. Dewpoint sensors may be installed to avoid condensation by increasing the inlet water temperature or reducing the water flow-rate. A chilled-ceiling system is composed of a network of tubes in which cold water circulates; generally the tubes are either made of plastic materials or metal (usually copper). The network of tubes is either attached to the ceiling and covered by a false ceiling, or directly attached to the elements of false ceilings, or clipped onto metal diffusers. Chilled ceilings are available in two different generic types: –– For drop-in ceiling applications (panel with back insulation and acoustic perforation) –– For free-hanging designs (design for enhanced upper surface) A major disadvantage of chilled ceilings is that no latent loads can be covered. Also their cooling capacity is rather limited. Typical values are in the range of 70 W/m² (22.2 BTU/(h*ft²) for drop-in ceiling applications and up to 140 W/m² (44.4 BTU/(h*ft²) for free-hanging designs (both figures are valid for a difference between room air temperature and average water temperature of 10 °C (50 °F)). For chilled ceilings, typical time constants are in the range of three to five minutes, as result of the lightweight construction of the panels (approximately 5 kg/m²).

2.5.2.3 Chilled floors The cooling/heating floor system consists of a network of tubes, in which hot or cold water circulates, depending on the seasonal needs. The floor network is often made entirely of plastic materials. It is then placed on an insulated layer (to reduce ground heat losses) and the entire construction is embedded in a floating slab. Floor heating systems are generally a well-established technology and common practice in residential or large commercial buildings. There is a E ­ uropean Standard

48

2.5 The load sub-system – air-conditioning equipment

on how to design and dimension floor heating systems /2.18/. On the other hand, floor cooling is a relative new idea, although several successful applications have demonstrated their applicability to maintain comfort conditions in a variety of buildings. The main advantage of a chilled floor is that the device is completely integrated into the floor and thus is perfectly invisible and does not diminish the exploitable indoor space. Due to the inertia of the floor, a mismatch between the supplied cooling power and cooling loads is buffered by the system; in fact, care should be taken that the floating slab (including the floor covering) does not have a very high thermal inertia (mass above the top of the insulation limited to about 160 kg/m²). Another advantage is that the system operates silently since no air is moved actively. From an energy point of view, the main advantage is that the temperature of the chilled water supplied to the floor ranges between 12–20 °C (53.6–68 °F) (depending on the load) and is considerably higher than for fan-coil systems, where they typically range between 7–12 °C (44.6–53.6 °F). Similar advantages occur during the heating mode, when the floor water-supply temperature is considerably lower than for conventional systems. The main disadvantages are the limited capacity and that it is not possible to cover latent loads. Depending on local weather conditions, appropriate control of the chilled-water supply temperature and flow-rate to the floor system is very important in order to avoid condensation problems. The use of appropriate ventilation to enhance indoor air circulation is essential for optimizing space cooling. The cooling capacity of a typical system lies in the range of 35–40 W/m² (11.1– 12.7 BTU/(h*ft²)) (indoor temperature 26 °C (78.8 °F), floor surface temperature 20 °C (68 °F)).

2.5.3 Air-water systems Air-water systems condition spaces by distributing air and water to terminal units installed in the zones to be conditioned throughout a building. The air and water are cooled or heated in central equipment rooms and from there are distributed to the air-conditioned spaces. The following systems can be distinguished: –– Induction system (two-pipe or four-pipe) –– Fan-coil system with supplementary air –– Radiant panels with supplementary air Air-water systems apply primarily to perimeter building zones/spaces with high sensible loads. They may, however, be applied to interior zones/spaces as well. These systems work well in office buildings, hospitals, hotels, research laboratories and other buildings where their functions meet the performance criteria.

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2  Meteorological data, heating and cooling loads and load sub-systems

2.5.3.1 Induction systems These systems use a high speed, high-pressure and constant air volume supply to a high-induction type terminal unit. High pressure means, that narrow ducting is required. Induced air from the room is either heated or cooled within the terminal as required. The capacity control is implemented by means of water flow regulation or a bypass. This system may use two pipes (one water circuit; supply and return) or four pipes (two independent water circuits) for heating and cooling. The primary air (fresh air) is usually conditioned in a central air-handling unit, while the secondary air (indoor air) is conditioned in the terminal unit. Therefore, this system requires central production of chilled (hot) water that is supplied to the air-handling unit and the local terminal units. A schematic drawing of a two-pipe induction unit is shown in Figure 2.3. Induction units are usually installed in building perimeter zones/spaces, for example, under a window, although units designed for overhead plenum installation are gaining popularity.



Fig. 2- 3 Schematic drawing of a two-pipe induction unit; primary air comes from the central air-handling unit, secondary air is indoor air which is cooled and blown back into the room

2.5.3.2 Fan-coil systems with supplementary air This type of system is a combination of a fan-coil with a centralised air-distribution system to provide fresh air. The advantage in comparison to a simple fan-coil is that heat-recovery (heating season) and indirect evaporative cooling (summer) of ventilation air can be easily achieved. A simpler alternative is to have a wall opening where the fan-coil draws in outdoor fresh air. This applies to perimeter building zones only.

50

2.5 The load sub-system – air-conditioning equipment

2.5.3.3 Chilled building components with supplementary air The combination of chilled building components, for example, ceilings, floors or walls, with a central ventilation system can provide high-comfort air-conditioning. The conditioned fresh air is supplied using a central air-handling unit. Sensible loads not covered by the ventilation air are extracted by the system of chilled building components.

/2.1/ /2.2/ /2.3/ /2.4/ /2.5/ /2.6/

/2.7/ /2.8/ /2.9/ /2.10/ /2.11/ /2.12/ /2.13/

/2.14/

/2.15/

Crawley, D.B., and Huang, Y.J., (1997). Does it matter which weather data you use in energy simulations?, DOE-2 User News, 18, pp. 2–12. Duffie, J.A. and Beckman W.A. (2006). Solar Engineering of Thermal Processes, 3rd Ed. J.Wiley & Sons Guide to Meteorological Instruments and Methods of Observation, WMO–No. 8, Seventh edition 2008, World Meteorological Organization, Geneva 2, Switzerland EN ISO 15927–1 (2004). Hygrothermal performance of buildings — Calculation and presentation of climatic data – Part 1: Monthly means of single meteorological elements Aubinet, M. (1994). Longwave sky radiation parameterization. Solar Energy. 53. No. 2. pp. 147–154. Mazzarella, L. (2002). L’impiego della temperatura operante esterna per il calcolo di dimensionamento ed energetico. 43° AICARR International Conference, “Ambient quality and Sustainable solutions”, Milan 7–8/03/2002, Vol. “Heating”, pp 149–187. TRNSYS (TRaNsient SYstem Simulation Program – Solar Energy Laboratory at the University of Wisconsin, W, USA – http://sel.me.wisc.edu/ EnergyPlus, http://apps1.eere.energy.gov/buildings/energyplus/ ESP-r – Energy Systems Research Unit, University of Strathclyde, Glasgow, http:// www.esru.strath.ac.uk/Programs/ESP-r.htm International Energy Association – Solar Heating and Cooling Programme: Task 25 Solar Assisted Air Conditioning in Buildings. http://www.iea-shctask25.org/ International Energy Association – Solar Heating and Cooling Programme: Task 38 Solar Air Conditioning and Refrigeration. http://www.iea-shc.org/task38/ Henning H.M., (Ed) (2003), “Solar Assisted Air Conditioning in Buildings – A Handbook for Planners” ISBN 3-211-00647-8, Springer, Vienna Jaehnig D., Weiss W., (2007). Design Guidelines – Solar Space Heating of Factory Buildings with Underfloor Heating Systems, IEA Solar Heating & Cooling Programme, Task 33 “Solar Heat for Industrial Processes” Report EN 15316-3-1, “Method for calculation of system energy requirements and system efficiencies – Part 3–1: Domestic hot water systems, characterization of needs (tapping requirements)” Vannoni C., Battisti R., Drigo S. (2008). Potential for Solar Heat in Industrial Processes. IEA Solar Heating & Cooling Programme, Task 33 “Solar Heat for Industrial Processes” Report.

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/2.16/ /2.17/ /2.18/

ASHRAE (2008). Handbook 2000 HVAC systems and equipment – SI edition. American Society of Heating, Refrigeration and Air-Conditioning, Atlanta GA/USA. Eurovent Standard 6/3 (1996). Thermal Test method for Fan-Coil Units.Eurovent/ Cecomaf, Paris/France. European Standard: Floor heating – Systems and components (/EN-1264-1:2011, 2:2008, 3–4:2009, 5:2008.

Abbreviations AES – Atmospheric Environment Service (CA) AMeDAS – Automated Meteorological Data Acquisition System (Jap) ASHRAE – American Society of Heating, Refrigerating and Air-Conditioning Engineers CEC – Commission of the European Community or California Energy Commission CIBSE – Chartered Institution of Building Services Engineers (UK) CWEC – Canadian Weather for Energy Calculations DOE – USA Department of Energy DRY – Design Reference Year EWY – Example Weather Year (UK) IWEC – International Weather for Energy Calculations NCDC – National Climatic Data Center NOAA – National Oceanographic and Atmospheric Administration (USA) TMY – Typical Meteorological Year (USA, EU) TRY – Test Reference Year (EU, USA) WMO – World Meteorological Organization WRDC – World Radiation Data Centre WYEC – Weather Year for Energy Calculations

52

Chapter 3

Components of solar thermal systems Responsible Author:

Joao Farinha Mendes, National Laboratory of Energy and Geology (LNEG), Portugal Contributing Authors: Maria Joao Carvalho, National Laboratory of Energy and Geology (LNEG), Portugal Peter Schossig, Fraunhofer Institute for Solar Energy Systems ISE, Germany

The heat production sub-system is the part of the overall system which provides heat to a thermally driven air-conditioning system. With regards to the solar-driven equipment, the solar collector is the main component of the sub-system. A brief review on the present state-of-the-art with respect to performance characteristics and available certification schemes of solar collectors is given in the following chapter. Besides the solar collector field, the storage unit and the back-up heat source unit are also important components. An overview of these items and the identification of other relevant components of the solar thermal system are given in the following sections.

3.1

The solar thermal collector

A solar thermal collector has the absorber surface as its main component where the absorbed solar radiation is transformed to heat. Part of this heat is transferred to the heat transfer fluid and the remainder is lost to the environment. Solar thermal collectors used in solar cooling systems usually have a transparent cover which separates the absorber from the environment and simultaneously allows as much incident solar radiation as possible to pass through to the absorber. In nonconcentrating and non-tracking flat plate collectors, the back of the collector is insulated to reduce losses to the environment. Concentrating collectors include a reflector as an additional component which redirects the radiation incident on the collector aperture to the absorber. In a steady state, the incident radiation on the collector surface is equal to the sum of useful heat and the heat losses , as is evident from the absorber energy balance and is given by the following expression:

AAG = Q& use + Q& loss,opt + Q& loss ,convective + Q& loss ,conductive + Q& loss ,radiative Eq. 3- 1

53

3  Components of solar thermal systems

where: AA G

is the absorber area (m² (1 m²=10.764 ft²)); is the incident global solar irradiance on the collector aperture (W/m²) (1W/m² = 0.32 BTU/(h*ft²)); ̇Quse is the useful heating power of the collector (W); are the optical collector losses (W); this term includes all losses that are due to Q̇ loss, opt. reflection and absorption in the transparent cover, i.e., the part of the incident radiation that does not reach the absorber; Q̇ loss, convective expresses convective losses (W); natural convection occurs in the gap between the absorber and the transparent cover and creates a heat flux from the absorber to the cover; ̇Qloss, conductive expresses heat conduction losses (W); they occur in the gap between the absorber and the cover as well as through the back insulation (depending on the construction of the collector) and through the frame edges; ̇Qloss, radiative denotes radiative losses (W); because of the absorber temperature these losses occur mainly in the infrared region of the spectral range. The different losses contribute to the steady state energy balance depending on the operating temperature of the collector and thus the temperature of the absorber. Measures to reduce these losses can be taken in order to maximize the useful heat output from the solar collector. These measures become clear in sections 3.2.1 to 3.2.6 where the different collector technologies are presented.

3.1.1 Assessment of the collector’s thermal performance Several terms related to solar geometry and basic concepts referring to solar radiation and its collection are used but not explained in detail here. For knowledge on these matters the reader is referred to specific textbooks /3.1/ or /3.2/. In this section solar energy vocabulary as defined in /3.3/ is used. Collector thermal performance is characterized either by its instantaneous efficiency, h, or by the instantaneous power delivered by the collector, . In a steady state these two values are related by the following equation /3.1/, where Aa is the collector aperture area and G is the global irradiance incident on the collector aperture area: 2 Q& use ∆T  ∆T  = k (θ ) ⋅ η o − c1 − c2 G  η=  Aa G G  G 

Eq. 3- 2

This equation follows directly from the energy balance in Equation 3.1, if all non-linear losses are

54

3.1 The solar thermal collector

approximated by a quadratic expression. The symbols used in Equation 3.2 have the following meaning: k(Θ) is the incident angle modifier (IAM), which accounts for the influence of nonperpendicular incident radiation at incidence angle, Θ, in relation to normal incidence radiation Θ = 0.; DT=Tm–Ta, and Tm is the average fluid temperature in the collector; for typical flowrates, the average fluid temperature can be expressed by the arithmetic average between the inlet and the outlet temperature. Ta is the ambient air temperature; h0, c1 (with the unit W/m²K) and c2 (with the unit W/m²K²) are the collector efficiency values, whereby h0 denotes the optical efficiency value and c1 the linear and c2 the quadratic loss coefficients, respectively. It should be stressed that when referring to the collector area different definitions are available. It is important, to have a common understanding on the reference area, when efficiency curves or costs per m² collector or other area related issues are discussed. In general, there are three different area definitions as shown in figure 3.1: the gross area, the aperture area (indicating the projected light catching area of the collector), and the absorber area (see /3.3/ for correct definitions of these areas).

Fig. 3- 1 Definition of the different collector areas for a flat-plate collector.

For stationary or one axis tracking collectors, additional information on the optical behavior of the collector is needed. The IAM was defined for this reason. The expression above represents only an approximation for the influence of the incident angle on the optical performance of the solar collector. In a detailed physical description the impact of the incident angle has to be considered separately for direct and diffuse radiation. Its value is always 1 in the case of a two axis tracking collector and its value is also 1 at normal incident radiation for any collector. In the case of flat plate collectors, the IAM is well characterized by one single measurement for an incidence angle equal 50º and is given by the equation:

K (θ ) =

η 0 (θ ) 1 = 1 − b0 ( − 1) η 0 (θ = 0) cos(θ )

Eq. 3- 3

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3  Components of solar thermal systems

Fig. 3- 2 IAM for a typical flat plate collector. K50 is the IAM for an incidence angle of 50º and gives the reduction in optical efficiency at this angle.

For collectors with asymmetrical characteristics such as stationary concentrating collectors, evacuated tubular collectors (ETC) and one axis tracking collectors, the IAM has to be measured in two directions, longitudinal and transversal, and the following expression can be used to determine K(q):

K (θ ) = K (θ l ,θ t ) ≈ K (θ l ,0) K (0,θ t ) Eq. 3- 4

Fig. 3- 3 Example of longitudinal and transversal measured IAM for a collector with asymmetric optical behavior.

The instantaneous collector power is given by: 2 Q& use = η o K (θ )GAa − c1 (Tm − Ta ) Aa − c2 (Tm − Ta ) Aa

Eq. 3- 5

For the use of equation (3.5) in the calculation of long term performances of collectors based on an appropriate climatological data base, it is important to consider that the optical effects of the different collector types (accounted for in the IAM) affect the components of solar radiation (direct, diffuse) differently; this effect can be considered by the introduction of two IAM values, the first one for direct radiation and the second for diffuse radiation. A power correction methodology was proposed by /3.4/ for steady state efficiency curve based calculations.

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3.1 The solar thermal collector

In Figure 3.4 different power curves for different collector types are represented.

Fig. 3- 4 Power curves for different collector types (1 W=3.4 Btu/h, 1.0 K differential =1.8 °F differential).

When transient effects are considered, the useful heating power of the collector is given by /3.5/:

dT 2 Q& use = η o K b (θ )Gdirect Aa + η o K d Gdiffuse Aa − c1 (Tm − Ta ) Aa − c2 (Tm − Ta ) Aa − c5 m Aa dt Eq. 3- 6

This equation is applicable to glazed collectors, while for unglazed collectors additional terms have to be considered /3.6/. When the collector is characterized by this equation no additional corrections are needed since the collector thermal behavior is described by six parameters and the IAM values are based on beam radiation and diffuse radiation, respectively. The six parameters are: –– the optical efficiency for normal incident radiation ho, –– the IAM for beam radiation Kb, –– the IAM for diffuse radiation Kd, –– a linear heat loss coefficient c1, –– a quadratic heat loss coefficient c2 and –– a dynamic response coefficient c5 representing the effective heat capacity of the collector normalized to the aperture area Aa.

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3  Components of solar thermal systems

When characterizing the thermal behavior of collectors with special optical features, such as concentrating collectors, this equation is more adequate. Two test methodologies are presently available to determine: 1) a collector’s instantaneous efficiency (equation 3.5) which is a steady state test method, or, b) the instantaneous power delivered by a collector (equation 3.6) which is a quasi-dynamic test method. The steady state test method is widely used and described in several test standards (/3.6/, /3.7/, /3.8/). The quasi-dynamic test method is described in the European Standard (/3.6/; section 6.3).

3.1.2 Collector yield for long term performance prediction The calculation of the energy delivered by a solar thermal collector for one year can be made based on the parameters determined in both test methodologies and is given as:

Q year =

1 365( d ) 24( h ) & ∑ ∑ Quse ∆t 365 i =1 j =1

Eq. 3- 7

For this calculation, information on climatic data – ambient temperature and irradiance values (global and diffuse or beam and diffuse) on an hourly basis – is needed. The irradiance incident on the collector is determined with the knowledge of specific optical collector characteristics /3.1/. The most straightforward information on annual heat output for a given collector is based on the assumption of a constant operation temperature, Tm, of the collector during the entire year. Such values of the so called collector gross heat production can be used to compare different collectors at a given location characterized by the annual time series of hourly average values of key meteorological data. These data can be derived from specific databases available in the different countries (see chapter 2).

3.2

Solar thermal collector technologies

The different thermally driven cooling technologies used in combination with solar thermal energy usually require driving heat within a range of temperatures between 50 °C to 250 °C (122 °F to 482 °F). In the present review of solar technologies only those collectors typically judged to be able to achieve those temperatures will be included. They are generally grouped in the following categories: –– flat plate collectors –– air collectors –– evacuated tube collectors

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3.2 Solar thermal collector technologies

–– –– ––

PV-thermal hybrid collectors stationary CPC type collectors tracking concentrators.

In this chapter brief technical descriptions of the different collector types are presented as well as data pertaining to their typical optical and thermal performance. The data presented here have been obtained from literature surveys and is related to collectors available in the market which have, in most cases, been tested in accredited laboratories around the world. An example is the data available on the European Solar Thermal Industry Federation (ESTIF) web site where the performance characteristics are listed for all those solar collectors certified according to the Solar Keymark scheme /3.9–3.11/.

3.2.1 Flat-plate collectors The simplest flat solar collector has a black surface with some fluid circulating in or around it extracting the radiation absorbed from the sun. The heat losses from such an absorber are large, but there are measures to reduce them, like placing it inside a box with insulation on the backside and a transparent cover serving as a lid to the box. This simple arrangement is known as a single cover flat plate collector /3.1/ (Figure 3- 5). With this collector type the absorbing area is more or less equal to the area of the cover, i.e. the area available for energy collection is more or less the effective area available for absorption. The surface may not be literally flat but often has some relief, in extreme cases even being made of tubes and densely packed fins.

Fig. 3- 5 Schematic representation of a flat plate collector

The fluid circulating in the absorber is mainly water (often with additives for freeze protection). Also other liquids are used, depending on the application and, in particular, on the operating temperature. Absorbers are usually built from copper, aluminum or sometimes even steel. Care must be taken in regard to the compatibility of the metals and the heat transfer media, in particular when more than one metal may be integrated in the collector loop. A very common configuration for the absorber is the fin and tube type shown in Figure 3- 6.

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3  Components of solar thermal systems

Fig. 3- 6 Fin and tube configuration of the absorber

The fin is typically made of aluminum and the tube of copper. Other materials, such as EPDM are used ,in particular for low temperature applications, i.e for swimming pool heating. Covers are usually made of glass, although some transparent plastic materials and other translucent materials have been used, but without great success so far due to the loss of transparency resulting from UV degradation. Back insulation is usually molded polyurethane foam, rock or glass wool. More sophisticated insulation materials may be employed when the operating temperature of the collector gets higher. In an attempt to control heat losses, flat-plate collectors incorporate different technologies; usually they are classified according to the applied heat loss mechanism. Selective and non-selective absorbers Radiative losses are one of the three mechanisms for heat losses in solar collectors. Non-selective coatings present normally significant values of emissivity in the wavelength range corresponding to the typical operation temperature of the absorber. As consequence even with high level of absorptivity in the solar radiation wavelength, they are a bottle neck for the reduction of radiative heat losses. The latter, instead, can be controlled by the use of so-called selective coatings on the absorber. These coatings are designed to have the highest possible absorption in the visible and short wave infrared range, and the lowest possible emissivity in the long wave infrared range corresponding to the collector operating temperatures. The emissivity values, found for the most common commercial coatings (e.g. Alanod /3.12/, Bluetec /3.13/, Tinox a /3.14/, Tinox b /3.15/), are typically between 0.05 and 0.25, with absorption between 0.9 and 0.96). Single cover and double cover; convective barriers Another heat loss mechanism is convection between absorber and transparent cover. One way to reduce the effect of convection is through the use of a double transparent cover with a second glass pane or a transparent Teflon film placed behind the first glass cover. The cover close to the absorber must have high transmissibility and very good heat resistance. An example is represented in Figure 3- 7.

Fig. 3- 7 Schematic representation of a flat-plate collector with minimized heat loss through additional convection barrier (Teflon foil) Source: LNEG

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3.2 Solar thermal collector technologies

With the incorporation of Teflon films, the heat loss coefficient drops typically by 1–1.5 W/m²K (0.18–0.26 Btu/h*ft²*°F) /3.16/. Solutions using two glass covers are also present in the market (see e.g. /3.17/). Transparent honeycomb structures have been developed and applied with good thermal results but with some degradation of optical performance and higher cost /3.18/; /3.16/. The principal problem with this alternative is the stability of the honeycomb under stagnation temperature, i.e. conditions with high radiation and no extraction of heat. Glass honeycombs are an alternative with two main disadvantages, high weight and very high cost. The reduction of heat loss by convection can also be achieved by partial vacuum and use of an inert gas in the absorber/cover chamber /3.17/, a solution requiring special maintenance requirements. Another strategy to improve the thermal performance of flat plates is the production of very large collectors, i.e. collectors with many square meters per unit, in contrast to conventional collectors with a size of about 2 m² (21.528 ft²). This leads to a reduction of the area of the collector frame and thereby also to heat loss reduction through the frame. In addition using thicker back insulation reduces heat loss from the back. Simultaneously, there have been successful attempts at cost reduction by on site assembling of these collectors from their components.

3.2.2 Solar air collector In general, the use of solar air collectors is recommended for all applications where warm or hot air is needed. Typical applications for solar air collectors are industrial processes where heated air in large volume flows is needed and air heating of residential and non-residential buildings (e.g. production halls, offices, seminar rooms). For cooling applications, air collectors can be a particularly interesting option to be combined with cooling technologies which need hot air as their driving heat, e.g. for drying of the desiccant in the case of Desiccant Evaporative Cooling (DEC) technologies. The general advantages of solar air collectors versus liquid collectors are: no freezing problems during winter, no overheating problems in summer, simple system components, no water-leakage problems, lower investment costs per installed capacity (kWth) installed. The general disadvantages of solar air collectors are: no standard heat storage units available on the market, system pressure drop (energy consumption of the fans) has to be considered more carefully than in liquid systems and collector efficiency usually lower in comparison to water collectors (due to lower heat transfer rates from the absorber to the airflow). Due to the low heat transfer coefficient between the absorber and the air, the contact area between absorber and air flow has to be large. This can be achieved by using ribbed absorbers, multi-pass systems or porous absorber microstructures. The general construction of solar air collectors is similar to solar collectors for water heating. The solar collectors which are available on the market can be distinguished into unglazed collectors, flat plate type collectors and evacuated tube collectors.

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3  Components of solar thermal systems

In the last few years, an increasing number of applications using vacuum air collectors has been reported , mainly from the Chinese market /3.19/. Low-cost vacuum collector technology using all glass, evacuated tube, solar collectors has been successfully applied to air collector technology. This type of collector is employed in the building sector for air heating and cooling and also in industrial drying processes. The flat plate type collector is the most widely used air collector and has been studied in detail. It may be divided into 4 basic categories related to the way the airflow removes heat from the absorber. In Figure 3- 8 these four categories are shown.

Fig. 3- 8 Schematic drawing of different types of solar air collectors: a) air flow above the absorber b) air flow under the absorber c) air flow around the absorber d) air flow through the perforated absorber

Type a) is the most simple in construction as the absorber is directly located on the insulation. The main disadvantage is the lower efficiency of the collector as the heated air flow is in direct thermal contact with the front glass cover and therefore the thermal losses are higher. For type b) these losses are significantly reduced as the air flow is under the absorber. To improve the heat flow from the absorber to the air, the absorber is mostly produced with fins. For type c) the heat transfer to the air flow is even better but then again the thermal losses due to contact with front glass cover increase. Type d) typically has the highest heat transfer rate, but on the other hand the pressure drop in the collector is higher and therefore the electric energy to force the air flow through the collector is higher. In deciding on a solar air collector not only has the thermal efficiency to be considered but also the electric energy consumption for ventilating the air through the collector. The construction type b) has been found to have a good overall performance and the commercially available air collector with the highest market share uses this construction. A diagram of the geometry of a solar air collector is shown in Figure 3- 9.

Fig. 3- 9 Schematic cross section of a solar air collector

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3.2 Solar thermal collector technologies

Air collectors using evacuated tubes have been introduced to the market recently. Air flows in parallel through several absorber tubes. An annular gap type absorber tube is used to direct and redirect the air through it. The evacuated tube is fully made of glass and has a selective coating on the inner glass tube inside the vacuum. This collector is able to produce heat at temperatures above 100 °C (212 °F) with sufficient efficiency. Thus the collector seems appropriately designed for the operation of heat driven cooling systems and in particular desiccant cooling systems (see chapter 5). An example of such a collector is shown in Figure 3- 10.

 

Fig. 3- 10 Solar air collector using evacuated tubes Source: Kollektorfabrik, Germany

3.2.3 Evacuated tube collectors Evacuated tube collectors eliminate the convection and conduction heat loss mechanism, by removing the air surrounding the absorber. Stability considerations dictate that these collectors assume the shape of a collection of tubes, usually glass tubes; inside these tubes there is a fin and tube arrangement (Figure 3- 11). Alternatively heat pipes can be used for energy extraction. This solution is already commercially available.

Fig. 3- 11 Evacuated tube collector with flat absorber (right side: indirect flux using heat pipe; left side: direct flux) Source: LNEG

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3  Components of solar thermal systems

The right part of Figure 3- 12 shows evacuated tubes which are constructed with two concentric glass tubes, the inner one selectively coated black, and air evacuated from the space in between. This type of evacuated tube collector is also known as Sydney type evacuated tubular collector.

Fig. 3- 12 Evacuated tube collector: Sydney type (left: direct flux; right: indirect flux using heat pipe) Source: LNEG

Evacuated tubes are generally produced with selective coatings on the absorbers and their typical heat loss coefficient ranges from 1.3 to 2.3 W/m²/K (0.23 to 0.40 Btu/h*ft²*°F). These values were obtained in a survey based on solar keymark data , and correspond to heat losses of complete collectors (i.e. also including manifold losses) /3.10/.

3.2.4 Evacuated flat plate collectors Evacuated flat plate collectors are produced with special frames and covers permitting them to maintain a high vacuum level (10-³Pa) (1.45 x 10-7 psi) despite a flat geometry /3.20/. Their performances are comparable to evacuated tube collectors. Stagnation temperatures can reach more than 300 °C (572 °F). Such a collector does not require a tracking system and can be mounted into a flat roof of e.g. industrial buildings for process heat. Some other commercial products are built with a lower level of vacuum (vacuum to be achieved on the roof after installation) at a lower cost. An example is shown in Figure 3- 13.

Fig. 3- 13 Evacuated flat plate collector. Source: Thermosolar

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3.2 Solar thermal collector technologies

3.2.5 PV-thermal hybrid collectors A PhotoVoltaic/Thermal (PV/T) collector is a combination of photovoltaic cells with a solar thermal collector. It simultaneously converts solar radiation into electricity and heat. An example of a PV/T solar collector can be seen in Figure 3- 14.

Fig. 3- 14 Example of a PV/T solar system. Source: IEA SHC – Task 35

PV/T collectors can generate more energy per unit surface area than side by side photovoltaic panels and solar thermal collectors, at a potentially lower production and installation cost. Moreover, PV/T collectors share the aesthetic advantage of PV. Because of their high efficiency per unit surface area, PV/T is particularly well suited for applications with both heat and power demand and with limited roof space available. Therefore, the potential of PV/T is especially large in the residential market, both collective and individual. There are many ways to combine the different PV and solar thermal technologies in a PV/T collector. Criterias against which PV/T collectors can be classified are: crystalline silicon, amorphous silicon or thin-film PV, liquid or air heat transfer media, flat-plate or concentrating technologies, with or without (transparent) cover, way and degree of building integration etc. So far, most developments have been made based on silicon technologies with liquid or air flat-plate collectors. Further work has been done on concentrating collectors using both liquid and air. Another field of interest by now is building integrated systems for instance for preheating of ventilation air. A general conflict of PV/T-collectors is that the PV efficiency drops with increasing operation temperature. This may be crucial in particular when PV/T-collectors are used in combination with thermally driven cooling systems because they need a certain minimum operation temperature which is always significantly higher than the ambient air temperature. For efficiency characterization of PV/T collectors no specific standard is available. Various authors have presented proposals for the evaluation and characterization of this type of collector /3.21/; /3.22/; /3.23/.

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3  Components of solar thermal systems

3.2.6 Stationary concentrating collectors Another way to decrease the heat losses of a collector is to reduce the absorber area in comparison to the collecting area. Reason for this is that heat losses are proportional to the absorber area while the gained solar energy is proportional to the collecting area. This approach is known as concentration, thus these collectors are called concentrating collectors. This type of collector is characterized by the concentration ratio C, which is the ratio of the collector aperture area Aa to the collector absorber area AA:

C=

Aa AA

Eq. 3- 8

Collectors using a stationary Compound Parabolic Concentrator (CPC) can achieve a small solar radiation concentration reaching a higher efficiency without significantly increasing the production and operation costs compared to flat-plate collectors. The acceptance angle,qa, also characterizes a concentrating collector and is defined as the maximum incidence angle of the beam (direct) solar radiation on the collector aperture that will reach the absorber without tracking; tracking means moving the collector to follow the daily path of the sun. Considering the above definitions of concentration ratio and acceptance angle, it is clear that small acceptance angles are related to high concentration factors and vice versa. For a certain acceptance angle, there is a maximum concentration factor that a collector can achieve. This maximum concentration is imposed by fundamental physical principles. For a two dimensional geometry the maximum concentration Cmax is given by:

C max =

1 sin θ a

Eq. 3- 9

Collectors that achieve this maximum concentration factor are considered “ideal” concentrators. CPC collectors are “ideal” concentrators. They are versatile because they can be designed to accommodate different absorber shapes, for example horizontal and vertical flat absorbers, tubular absorbers etc. For each absorber shape the reflector shape is designed to ensure that all beam (direct) solar radiation reaching the collector aperture within the acceptance angle will reach the absorber, even if it is reflected more than once by the mirror (reflector). Stationary CPC collectors used for the heating of a liquid fluid at low (50–70 °C) (122–158 °F) and medium temperatures (80–110 °C) (176–230 °F) have two orthogonal axes of symmetry and

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are designed with acceptance angles greater than 30º to avoid the need for tracking the sun. As a result they usually have concentration factors lower than 2. Practical design limitations, such as the height of the collector box, can also impose the use of truncated CPC’s. This means that the upper part of the mirror is cut and sunrays with incident angles higher than the acceptance angle can still impinge directly on the absorber, but when they strike the mirror, they are reflected out of the collector. A truncated CPC has a concentration factor which is determined by Eq. 3- 8. An aspect of CPC’s is that they collect beam, diffuse sky and ground-reflected solar radiation. This has to be taken into account when determining the collected solar energy for the different collector trough orientations, typically east-west for forced circulation systems or north-south for thermosiphon systems. The efficiency of stationary CPC collectors is determined using the same standards as for flat plate collectors. Due to its asymmetric optical behavior a quasi-dynamic test method is recommended for thermal performance characterization (see equation 3.6). The thermal performance will depend on the collector’s design and especially on the concentration ratio. Manufacturers of stationary CPC collectors use selective surfaces for the absorbers. As a result these collectors can have lower heat losses than flat-plate selective collectors, depending on the concentration ratio they achieve. A cross-section scheme of a stationary CPC collector is shown in Figure 3- 15; the figure also shows an example of an installation where this type of collector is installed in combination with a desiccant cooling system. Such a collector is designed with concentration factors of about 1.12 for domestic hot water production and about 1.5 for production of heat in the range between 80 °C (176 °F) and 110 °C (230 °F), which can be used to operate thermally driven chillers.

Fig. 3- 15 CPC type collector (top: schematic representation; bottom: installation using this type of collector) Source: AoSol

Fig. 3- 16 CPC type collector (top: schematic representation; bottom: installation using this type of collector) Source: AoSol

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3  Components of solar thermal systems

3.2.7 Solar concentrating tracking collectors As described in the previous section, concentration is a way to reduce collector heat loss due to the reduction of absorber area. Although CPC type collectors are ideal concentrators, for high concentration ratios (C>>2) the CPC should track the sun path. The CPC trough is very long and impractical. For this reason other concentrating devices are used. Their design is based on nonimaging optics. Within this class of collectors, only those that work in a medium temperature range (up to 250 °C) (482 °F) are of interest in designing solar cooling systems. Solar tracking concentrators are classified depending on the way they track the sun movement. This is achieved by one-axis solar tracking and linear focus systems or two axis tracking and point focus systems. Therefore, the first system can track the sun along its elevation or azimuth angle, whereas with the second type the sun rays are always perpendicular to the collector surfaces. Of this second type there are: parabolic dishes, tower plants with heliostats and solar furnaces. For the purpose of this handbook we restrict the description only to linear focus systems because they are especially relevant to solar cooling applications. The parabolic trough The most characteristic single axis-tracking collector is the so called parabolic trough c­ ollector (PTC). It consists of a cylindrical reflective surface of parabolic cross section and a receptor located in the parabola focus. This surface reflects, while it concentrates, incident direct solar radiation over the frame. This makes a fundamental difference between flat plate collectors and concentrating type collectors, as the former ones collect direct and diffuse radiation whereas the later ones only collect beam radiation. A mirror of parabolic cross section has the property of taking all rays incident parallel to the ­parabola symmetry axis to its focus. Its geometry is defined by the rim angle F (Figure 3- 17) and by the concentration ratio:

C=

2a 1 sin Φ = ⋅ 2πr sin θa π

Eq. 3- 10

For the definition of angles and segments see Figure 3- 17:

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3.2 Solar thermal collector technologies

Fig. 3- 17 Geometry of the parabolic trough collector. Source: LNEG

Equation 3- 10 shows that the concentration falls very short of the limit defined in section 3.2.5. For an acceptance angle equal to half the angular aperture of the sun disc (qa~ 4.65 rad*10-³), their limit would be C = 215, and because of the geometrical optics inherent to this design, the maximum concentration is reduced by a factor of (sinF)/p, which means that a parabolic trough having a perfect mirror, can hardly exceed a concentration value of 68 (rim angle F= p /2). In practice the values achieved are more typically between 25 and 35 due to mirror imperfections that enlarge the reflected cone of rays from each point of the reflector, and require an increase in the size of the absorber to make the collection of all the radiation possible. This means that the receptor consists of an absorber tube of an area 25 to 35 times smaller than the aperture area. The fluid to be heated – normally oil or water – is circulated through the absorber piping. The absorber is normally protected by a glass envelope, evacuated in most cases. The concentration results in the loss of most of the diffuse radiation. Thermal losses are reduced because of the inversely proportional relationship with concentration. Moreover, the optical efficiency of the collector is reduced when concentration is increased. This reduction is due to the fact that it is not possible for a system with inherent fabrication errors to collect all the available radiation at its entrance aperture and focus it to an ever smaller area. The optimal concentration will be obtained as a result of minimizing the sum of optical and thermal losses. The Linear Fresnel A Linear Fresnel Reflector (LFR) is an approximation of the parabolic trough by segmented mirrors according to the Fresnel principle. Simple designs of flexibly-bent mirrors and fixed receivers, with the possibility of having secondary concentration and direct steam generation, entail lower investment costs than troughs and the close arrangement of the mirrors requires less land

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3  Components of solar thermal systems

and provides a partially shaded, useful space below, that can be used for agricultural purposes. The principle is shown in Figure 3- 16. Compared to parabolic trough plants, the LFR plants are in general less efficient in converting solar energy to electricity. But for medium temperature applications like industrial processes or cooling applications, the relatively simple and low cost technology involved, makes it very attractive.

Fig. 3- 18 Schematic representation of a Fresnel collector Source: Industrial Solar GmbH, Germany

The main application of this kind of sun tracking collectors has been electricity generation with the working fluid of the collector loop being connected to a conventional power generation process. The overall efficiency can be increased further using the heat at the outlet of the thermomechanical process (e.g. steam turbine) for process heat applications or operation of a thermally driven refrigeration process. Regarding the history of power generation it is interesting to mention the first fields of collectors with a total installed electric capacity of 350 MW installed in California in the 1990s. These systems, denominated SEGS (Solar Energy Generating Systems), were installed by the company LUZ. These systems still supply electricity to the Southern California Edison grid. The experience gained from SEGS is based on present large commercial projects already concretized and in development not only in the Mediterranean area but also along all the earth’s sun-belt. Recent research is focused on developing and on optimizing smaller low cost PTCs for medium temperature applications. The temperature target ranges between 150–250 °C (302–482 °F). The concentration ratio and optical tolerance could be reduced allowing the use of simplified systems and, as a consequence, this has led to cost reduction: cheaper materials, coating and treatments, cheaper production and assembling costs and simpler control systems. In this context several systems are available on the market specifically developed for solar cooling and industrial process heat applications. Work done in Task 33 “Solar heat for industrial processes” of the IEA Solar Heating & Cooling Programme showed the latest developments in this field, collecting not only prototypes but also available commercial products (brochure of /3.17/). Cooling applications with parabolic troughs and LFR technology are presently under evaluation in different places /3.24/.

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The use of tracked concentrating collectors is generally more appropriate at sites with high direct (beam) radiation potential. However, a more detailed analysis is necessary in order to determine the energy production of such a system. Figure 3- 19 shows two examples of concentrating collectors.

Fig. 3- 19 Examples of single-axis tracked concentrating solar thermal collectors. Left: Fresnel collector for hot water preparation in the medium temperature range up to 200 °C. The mirrors are tracked to focus the direct radiation towards the absorber, located above the mirror area. Advantage: low sensitivity to high wind speeds. Source: PSE, Germany Right: Parabolic trough collector, developed by Button Energy, Austria. The collector is designed for steam production and is part of a research project with a solar thermally driven steam jet ejector chiller at AEE-INTEC, Austria.

3.2.8 Summary A broad variety of solar thermal collectors is available and many of them are applicable in solar cooling and air-conditioning systems. However, the appropriate type of the collector depends on the selected cooling technology and on the site conditions, i.e., on the radiation availability. Typical efficiency curves for collectors are displayed in Figure 3- 20 (steady-state efficiency for two different radiation conditions; no dynamic behaviour is reflected in this figure). The same figure also shows typical operation ranges of different thermally driven cooling technologies.

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Fig. 3- 20 Typical efficiency curves for stationary collectors, calculated from parameters related to the aperture area of the collector (not taking into account other collector characteristics as e.g. IAM). For the higher temperature range, the efficiency range of a 1-axis tracked concentrating collector is included as an example (dashed red line) (°F = °C*(9/5) + 32).

FK-ST FK-AR FK-HT VRK-CPC

Flat-plate collector, standard product Flat-plate collector, 1-cover glass, anti-reflective coated Flat-plate collector, 1-cover glass, convection barrier foil, improved insulation Evacuated tube collector, direct mass flow, Sydney type with external CPC-reflector

The curves are drawn for ambient temperature of 25 °C (77 °F) and 800 W/m² (254 Btu/h*ft²) radiation level (left) as well as for 400 W/m² (127 Btu/h*ft²) (right). The figure approximately includes the application range of the most interesting cooling technologies. As the graph represents only steady-state operation conditions and only exemplary sets of efficiency curves, it is not sufficient for a serious appraisal and decision on a specific type of collector in a system to be planned. For the higher temperature range, the efficiency range of a 1-axis tracked concentrating collector is included as an example (dashed red line). Although the efficiency curves of stationary collectors may be theoretically drawn also for higher temperatures, they have been cut in the figure at temperatures > 120 °C (248 °F), since we have too little experience with these collectors at higher temperatures (and thus pressure levels). Source: Fraunhofer ISE

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3.3 Testing and certification of solar thermal collectors

3.3

Testing and certification of solar thermal collectors

In the following methods for testing and certification of solar thermal collectors are presented.

3.3.1 Applicable test standards Standards describing testing methodologies for the assessment of a collectorʼs thermal performance have been developed since the 1970s. In the 80s testing methodologies for the assessment of a collectorʼs reliability and durability were also developed. The research work leading to standards was frequently based on the work of experts within the Solar Heating and Cooling Programme of the International Energy Agency (IEA). This is the case in some of the already completed tasks: IEA – SHC Task 01, “Investigation of the Performance of Solar Heating and Cooling Systems,” 1977–1983 IEA – SHC Task 03, “Performance Testing of Solar Collectors,” 1977–1987 IEA – SHC Task 10, “Solar Materials R&D”, 1985–1991 IEA – SHC Task 27, “Performance of Solar Façade Components”, 2000–2005 IEA – SHC Task 33, “Solar heat for industrial processes”, 2003–2007 IEA – SHC Task 43, “Solar Rating and Certification Procedures, started in 2009 /3.25–3.30/. In the 1990s, a growing solar collector market, especially in Europe, created the need for another instrument for quality assurance – certification of solar collectors. For this purpose, product standards were needed. The work developed within “CEN TC 312 – Thermal solar systems and components” /3.31/, led to the approval of standards for solar thermal collectors and for factory made systems. Table 3- 1 presents the main characteristics of applicable standards for solar thermal collectors.

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3  Components of solar thermal systems

Table 3- 1 Characteristics of applicable standards for solar thermal collectors

The first version of the standard was approved in 2000/2001 and revised in 2006, the date of the present version. In the case of solar thermal collectors, the present version covers mainly stationary collectors (flat plate, evacuated tube collectors and CPC type collectors). The testing methodologies for assessment of thermal performance are also applicable to tracking collectors, as stated in the scope of EN12975: part 2, if the quasi-dynamic test method (/3.6/: part 2; section 6.3) is adopted as a testing methodology. A new period for revision of the standards was initiated in 2009 and is expected to lead to a larger applicability of the standards to new products (collectors) covering mainly the applications in the medium and high temperature range which is of high relevance to solar cooling systems.

3.3.2 Certification schemes In the following section, the situation regarding certification schemes for solar thermal collectors and systems is described for different economic regions.

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3.3.2.1 Situation in Europe In Europe the development of the product standards described in the previous section allowed the development of a “Certification Scheme for Solar Thermal Collectors and Factory Made Systems”. The certification scheme adopted was created by Comité Européen de Normalisation (CEN, European Committee for Standardisation) and is generally called Keymark /3.32/. Within Keymark, specific schemes were developed which are applicable to solar thermal products and called Solar Keymark /3.9/. The Solar Keymark is based on the principal of type testing and factory inspection and observation. The tests are performed by accredited laboratories and the licenses for use of the Solar Keymark are issued by accredited certification bodies. Since the approval of the Solar Keymark scheme was created by CEN in 2003, the number of licenses issued for solar thermal collectors has grown exponentially, /3.11/. Information on testing laboratories, certification bodies, a data base of solar thermal collectors and factory-made systems with Solar Keymark Certification is provided on the website of the European Solar Thermal Industry Federation ESTIF (see /3.33/).

3.3.2.2 Situation in other countries/regions In the US the Solar Rating and Certification Corporation SRCC certifies solar collectors and solar domestic hot water systems based on ISO Standards for collectors and ANSI Standards for systems. Detailed information can be obtained at the SRCC website /3.34/. In China the certification label is the Golden Sun, based on national standards and developed by the China General Certification Center/3.35/. Other countries like Brazil also have developed national certification schemes for solar thermal products. The first steps towards a global certification harmonization for solar thermal collectors and systems are ongoing /3.30/.

3.4

Heat storage

The main purpose of storage in a solar-assisted air-conditioning system is to balance mismatches between solar gains and cooling loads. The most common application is the integration of a hot water buffer tank in the heating cycle of the thermally driven cooling equipment. When a desiccant cooling system is used, this is the most common place to integrate storage into the system. However, in a solar-assisted air-conditioning system that uses thermally driven chillers, there are two obvious places for integrating thermal storage (Figure 3- 21). One option is that the excess solar heat can be stored in the heat storage unit and is made available if the solar heat is not sufficient. The second option is that the excess cooling power of the solar system and the thermally driven chiller is stored in a cold storage unit and is made available if the cold production does not meet the cooling load (see section 4.4.3).

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3  Components of solar thermal systems

Fig. 3- 21 Energy storage typologies in a solar-assisted air-conditioning system

The following sections describe the most common methods of heat storage in thermal energy systems and highlight the techniques that are most relevant in using solar thermal energy for airconditioning.

3.4.1 Hot water stores One of the key elements of a solar-assisted air-conditioning system is the hot water tank. The storage unit fulfils several tasks: –– It delivers sufficient energy to the heat sink (with appropriate mass flow and temperature). –– It decouples the mass flows between heat sources and heat sinks. –– It stores heat from fluctuating heat sources (i.e., solar) from times where excess heat is available for times where too little or no heat is available. –– It extends the operation times for auxiliary heating devices. –– It reduces the needed heating capacity of auxiliary heating devices. –– It stores the heat at the appropriate temperature levels avoiding mixing in order to reduce exergy losses (i.e., stratification). The minimum required hot water storage volume to provide cold to the load in a thermally driven cooling system is shown in Figure 3- 22 in comparison to the respective figure for chilled-water storage.

Fig. 3- 22 Required storage volume (without insulation) for chilled water storage and hot water storage for different values of the COP of a thermally driven chiller as a function of the useful temperature difference. The required storage volume is given in required to provide 1 kWh of cold. 1 m³/kWh = 0.077 USgal/Btu, 1.0 K differential = 1.8 °F differential.

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3.4 Heat storage

Figure 3- 23 illustrates the operating principles of a hot water storage tank with two heating inputs (solar and auxiliary); two alternative designs for coupling the solar collector system are shown, namely an internal heat exchanger and the connection to an external heat exchanger. The temperature of water flowing from the storage tank to the solar collectors should be as low as possible, in order to obtain the highest possible collector efficiency.

Fig. 3- 23 Schematic drawing of a hot water tank storage for solar assisted air-conditioning system. Alternative constructions with either internal (*) or external heat exchanger (**) are shown

The configuration of the piping connections to the different heat sources and heat sinks are described below in order to demonstrate the complexity of such a system. For small collector areas (smaller than 15 m² (161.5 ft²)), an internal heat exchanger located at the bottom of the storage tank and a high-flow strategy for the collector (maximum temperature rise in the collector about 10 °C (18 °F)), are recommended. An example of such a storage tank is shown in Figure 3- 24. If the collector areas are larger, external heat exchangers are preferred, because they can exchange higher thermal capacities with small temperature drops. In the case of the internal heat exchanger, the connecting tube to the collector heat exchanger is mounted at the bottom. The height of the inlet from the collector heat exchanger into the tank varies for different applications. For highflow configurations, this inlet can be positioned in the lower part of the tank in order to slowly heat up the tank contents from the bottom upwards. A high-flow configuration typically leads to a temperature rise of about 5–10 °C (9–18 °F) and resulting flow-rates through the collector are in the range of 30–70 litres/m² (0.012–0.029US gpm/ft²). For low-flow configurations, which lead to a maximum temperature rise in the collector of up to 40 °C (72 °F), the inlet should always be placed at the level with the same temperature as the collector outlet temperature. This can be achieved with special devices, known as stratifiers, as shown in the example in Figure 3- 25. Low-flow systems are often applied for domestic hot water preparation or combined solar systems for hot water and space heating, since a large temperature rise is needed to heat up the cold water coming from the mains to the desired hot water temperature.

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3  Components of solar thermal systems

Fig. 3- 24 Hot water tank with internal heat exchangers for collector areas up to 15 m² (161.46 ft²) Source: Viessmann, Germany

Fig. 3- 25 Stratifying units for hot water storage Source: Solvis, Germany

Although advantageous in solar domestic hot water systems, low-flow design is not recommended in solar-assisted air-conditioning systems. The first reason is that thermally driven cooling equipment generally works at comparatively low temperature differences between inlet and outlet of e.g. 10 °C (18 °F), whereas low-flow regimes produce much larger temperature differences, especially at high radiation levels. Another reason arises from the fact that temperatures up to 100 °C (212 °F) are used in solar-assisted air-conditioning. However, temperatures above 100 °C (212 °F) are not feasible if the tank is operated at atmospheric pressure, to keep investment costs low. Under these conditions, the large temperature differences of up to 40 °C (72 °F) achieved by low-flow operation are unfavorable. In the worst case, the temperature in the tank would not rise above 60 °C (140 °F) by the collector, because temperatures above 100 °C (212 °F) would occur at the collector outlet and the collector pump would be switched off by the control unit in order to prevent boiling. To avoid this, matched-flow systems are often used in solar-assisted air-conditioning, where the mass flow through the solar collector array is varied in such a way that the collector outlet is about 10 °C (18 °F) higher than the required operation temperature of the thermally driven chiller or thermally driven desiccant system (e.g. generator, regeneration heat exchanger). When the whole storage volume has already reached this temperature level, the collector is operated in the high-flow mode and the tank is charged to its maximum allowed temperature (e.g. 95 °C (203 °F)). For matched-flow systems, the inlet from the collector can be positioned close to the top of the tank. The inlet pipe from the auxiliary heater should be located at the top of the tank. To determine the outlet position to the auxiliary heater, the following aspects should be taken into account: –– The tank should always contain enough hot water to meet the heat sink demand. The volume for the auxiliary heater can be calculated from the power of the auxiliary heater and the maximum heat sink demand over a specific time period. –– The auxiliary heater often needs a minimum operating time (especially solid wood burners). The volume between the auxiliary heater inlet and outlet must allow for this minimum operating time.

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3.4 Heat storage

––

––

The outlet to the auxiliary heater should be positioned as high as possible in the storage tank (taking into account the previous aspects) in order to leave the largest possible water volume in the storage tank for the solar collector. If the auxiliary heater has enough power, it can also be positioned behind the storage tank in the supply to the heat sink. This construction provides maximum storage capacity for the solar collector.

Figure 3- 26 illustrates the integration of a collector system with a collector area greater than 15 m² (161.4 ft²) using a matched-flow design (collector inlet close to the top of the tank), auxiliary burner and the connection to the heat sink (solar-assisted air-conditioning system).



Fig. 3- 26 Schematic drawing illustrating the integration of a solar collector to a hot water storage with external heat exchanger. Also the connection to the auxiliary heater and to the heat sink is shown

3.4.2 Storages with phase change materials The thermal energy absorbed by a material when changing its phase at a constant temperature is called “latent heat”. For practical applications, materials that exhibit low volume changes are used, e.g. solid-to-liquid and some special solid-to-solid phase change materials are applicable. During the phase change, a large amount of energy is absorbed without a change in temperature. The commonly used phase change materials for technical applications are paraffin (organic), salt hydrates (inorganic) and fatty acids (organic). For cooling applications, it is also possible to use ice storage; of course, chillers producing cold temperatures below 0 °C (32 °F) are necessary in order to apply this kind of storage strategy. This is not possible using common thermally driven chillers like absorption machines with water-LiBr or adsorption chillers using water as refrigerant. Phase change materials can be incorporated into a thermal storage system either on the cold or hot side of the thermal cooling unit, depending on the melting temperature of the phase change material.

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3  Components of solar thermal systems

Figure 3- 27 shows materials for storage application in the low temperature range (cold side) on the left and materials for heat storage application (hot side) on the right. It is evident from the data that latent heat storage offers a significant advantage if only a small temperature difference is usable, since in these cases the corresponding storage density of water is small.

Fig. 3- 27 Storage density of different phase change materials in comparison to water-ice (cold storage) and pure water (heat storage); only the material related values are shown excluding all equipment associated with storage, such as heat exchangers, insulation etc. Minimum temperatures of –5 °C and 50 °C (23 °F and 122 °F) have been taken as references for cold and hot storage, respectively. The following phase change materials were used in the figure /3.36/

Organic PCM-1: Tetradecane Inorganic PCM-2: Na2SO4 .10 . H2O/KCl/NH4Cl

Organic PCM-3: Nonacosane Inorganic PCM-4: Mg(NO3) . 6 . H2O

°F = °C × (9/5) + 32 1 kWh/m³=12.9 Btu/USgal

In all cases, heat must be transferred between the phase change material and the fluid cycle (­charging, discharging). To optimize the heat transfer processes, different techniques are used in practical applications which include: –– Direct contact between phase change material and heat transfer fluid: this is possible only if the two materials are chemically stable for long periods of direct contact and the solidification does not occur in a uniform block, preventing sufficient heat transfer during subsequent melting. –– Macroscopic-capsules: this is the most frequently used encapsulation method. The most common approach is to use a plastic module, which is chemically neutral with respect to both the phase change material and the heat transfer fluid. The modules typically have a diameter of some centimetres. An example of a capsule is shown in Figure 3- 28. –– Micro-encapsulation: this is a relatively new encapsulation technique in which the phase change material is encapsulated in a small shell of polymer materials with a diameter of some micrometres. A large heat-exchange surface results and the powderlike spheres can be integrated into many construction materials. Up to now this technique is only feasible with materials which are not soluble in water, as is the case for paraffins. There are several products on the market for building applications and textiles, especially in the 18–38 °C (64–100 °F) temperature range, for temperatures higher than 60 °C (140 °F) this technology is still in the development phase.

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3.4 Heat storage

––

Phase change slurries, a mixture of a PCM-particles and a heat transfer fluid, makes it possible to benefit from the higher storage capacity of PCMs while keeping the advantages of a pumpable fluid. There are two types of PCM slurries on the market, both so far only for cold storages: so called Ice-slurries, a water/ice mixture (per definition below zero) and paraffin microcapsules dispersed in water. For both types of storage, capacities of several hundred cubic meters exist. A third slurry type is still under development: paraffin emulsions. Here the paraffin is mixed with water and an emulsifier to get stable droplets, avoiding the cost of encapsulation and leading to better fluid properties. At the moment there are still open questions regarding the long term stability of paraffin emulsion phase change storage.

Fig. 3- 28 Example of a phase change material capsule with a diameter of some centimetres Source: Cristopia, France

The main advantages of phase change storage in comparison to conventional water storage ­techniques are: –– Higher thermal energy storage capacity than sensible energy storage, at least when only small useful temperature differences are to be achieved; this significantly reduces the volume required for a latent heat storage unit compared to a conventional hot water storage unit. –– Relatively constant temperature during charging and discharging; this may result in a considerable increase of the solar collector efficiency, since the collector performance decreases with increasing operating temperatures. Depending on the phase transition temperature the storage losses may also be lower than for water storage because of the more uniform storage temperature. –– Burner cycles for the back-up heat generation unit can be reduced. The main disadvantages of phase change storage are: –– Higher investment cost, in most cases, compared to water storage. –– In many cases, the peak power during discharge is limited. The reason is the growing thickness of the solid layer, which causes increasing heat resistance during discharge. This is the main limit determining the acceptable size for the storage modules. –– Limited experience with long-term operation (after many thousand cycles). –– Risks of loss of stability of the solution and deterioration of the encapsulation material.

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3  Components of solar thermal systems

3.5

Backup heater

There are various possible options to provide back-up in a solar-assisted air-conditioning application. Back-up can be employed on either the “hot” or the “cold” side. Back-up heat may be provided from different heat sources. Back-up heaters are needed to assure that solar-assisted air-conditioning systems always have enough heat available to meet the load. For the cooling process, they are needed especially during days with hot and humid weather and simultaneous clouds, i.e., on days with high cooling loads on which the solar system does not deliver enough heat at the required temperature. Additionally, they can be used for space heating during the heating season, when the heat output from the collectors is not sufficient. In this chapter we refer to gas heaters as back-up heat sources. Gas heaters can be categorized according to different criteria: –– single-stage or capacity-controlled burner; –– atmospheric or ventilated burner; –– high-temperature, low-temperature and condensing burners; –– “catalytic” burner; –– tank-integrated or with separate boiler. Figure 3- 29 shows two examples of gas-fired water heaters. The heating capacity of gas-fired burners can have one or two fixed values (two-stage) or a range of values (typically from 30 % to 100 % nominal capacity). In the latter case, they are also referred to as modulating gas-fired burners. If the heating capacity is fixed, a water storage tank should also be coupled to the gasfired burner in order to avoid cycling operation. Most of today’s small and medium-sized gasfired burners use atmospheric burners. This means that gas and air are mixed in a kind of venturi nozzle before they are ignited. The gas-air mixture burns upwards without a fan. The advantages are low noise emission and simple construction. Ventilated burners are more efficient due to the better mixing of air and gas, but are noisier and technically more complex. The flue gas energy losses can be minimized by lowering the temperature of the flue gas of the burner using a heat sink. When the flue gas temperature drops below 60 °C (140 °F), the water content of the flue gas starts to condense. Below 30 °C (86 °F), most of the water has condensed. Since the condensing heat is also useful heat, the burner efficiency can increase up to approximately 110 %, based on the lower heating value of the fuel. Non-condensing burners reach efficiency values up to 95 %, based on the lower heating value of the fuel. In the event that condensation occurs, the burners must be protected against corrosion. For solar-assisted air-conditioning, the driving temperature is usually too high to achieve gas condensation.

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3.5 Backup heater

Fig. 3- 29 Low temperature two-stage atmospheric burner (right) and modulating gas-condensing catalytic ventilated burner (left) Source: Viessmann, Germany

When the combustion takes place around a special catalyst, the temperature can be lowered. This results in lower NOx emissions. Some manufacturers have included this type of burner in their gas fired boilers. Small gas-fired burners of up to 20 kW (68,303 Btu/h) are sometimes directly mounted in the water storage tank. The boilers for burners with higher heating capacity are installed separately, because the heat transfer rate to the water of the tank is too low and the space available in the tank is too small. In many systems the upper part of the water storage unit for the solar thermal collectors is also used as heat storage for the gas-fired boiler (see Figure 3- 26 in Section 3.1 above). However, the gas-fired burner is assumed to be independent of the water storage tank.

/3.1/ /3.2/ /3.3/ /3.4/ /3.5/

/3.6/ /3.7/ /3.8/ /3.9/

A. Rabl, Active Solar Collectors and Their Applications, Oxford University Press, 1985. J. A. Duffie und W. A. Beckman, Solar engineering of thermal processes, John Wiley and Sons, 2006. EN ISO 9488, Solar Energy. Vocabulary, CEN, 1999. Horta et al 2008 power correction iam steady state B. Perers, “An improved dynamic solar collector test method for determination of non linear optical and thermal characteristics with multiple regression,” Solar Energy, vol. 59, no. 4–6, pp. 163 –178, 1997. EN12975-2:2006, “Thermal Solar Systems and Components – Solar Collector. Part 2: Test Methods,” CEN, 2006. ISO 9806-1:1995, Test methods for solar collectors – Part 1: Thermal performance of glazed liquid heating collectors including pressure drop, 1995. ANSI/ASHRAE, Standard 93 Methods of testing to determine the thermal performance of solar collectors, 2003. Solar Keymark, “Specific CEN Keymark Scheme Rules for Solar Thermal Products, Version 11.04 – December 2009,,” [Online]. Available: http://www.estif.org/ solarkeymark/Links/Internal_links/network/sknwebdoclist/SKN_N0106R3.pdf. [Accessed 24 May 2011].

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/3.10/

/3.11/

/3.12/ /3.13/ /3.14/ /3.15/ /3.16/

/3.17/

/3.18/

/3.19/

/3.20/ /3.21/

/3.22/ /3.23/ /3.24/ /3.25/

/3.26/

84

Solar Keymark, “Solar Keymark Collector Data Base, 2011,” [Online]. Available: http://solarkey.dk/solarkeymarkdata/qCollectorCertificates/ ShowQCollectorCertificatesTable.aspx. [Accessed 24 May 2011]. Solar Keymark, “Solar Keymark Status Update, 2011,” [Online]. Available: http:// www.estif.org/solarkeymark/Links/Internal_links/database/status-updated.pdf. [Accessed 24 May 2011]. ALANOD, [Online]. Available: http://alanod-solar.com/opencms/opencms/ Absorption/Technische_Info.html. [Accessed 16 May 2011]. Bluetec, [Online]. Available: http://www.bluetec-germany.com/fileadmin/user_upload/ pdf/produktinfos/091012_web_techn_doku_eng_reduced2.pdf. [Accessed 16 May 2011]. Tinox, a, [Online]. Available: http://www.almeco-tinox.com/en/products/solar_ absorber/tinox_energy_cu. [Accessed 16 May 2011]. Tinox, b, [Online]. Available: http://www.almeco-tinox.com/en/products/solar_ absorber/tinox_energy_al. [Accessed 16 May 2011]. C. Pereira, M. J. Carvalho, J. F. Mendes and J. Oliveira , “Optical and thermal testing of a new 1.12xCPC solar collector,” Solar Energy Materials and solar cells, no. 37, pp. 175–190, 1995. W. Weiss and M. Rommel, “Process Heat Collectors – State of the Art within Task 33/ IV,” 2008. [Online]. Available: www.iea-shc.org/.../downloads/task33-Process_Heat_ Collectors.pdf. [Accessed 24 May 2011]. M. Rommel and V. Wittwer, “Flat plate collector for process heat with honeycomb cover – an alternative to vaccum tube collectors,” in Proceedings ISES Solar World Congress 1987, Hamburg, 1987. W. Weiss, F. Mauthner (2012). Solar Heat Worldwide Markets and Contribution to the Energy Supply 2010. IEA SHC 2012 Edition [Online]. Available: http://www.iea-shc. org/statistics/SolarHeatWorldwide/index.html. [Accessed 30 May 2012]. H. Hillemans, „NEG commercialisation status at CERN Surface Conditioning for Ultra High Vacuum,“ in Brokering Meeting Daresbury Laboratory, Warrington, 2009. P. Hofmann, P. Dupeyrat, K. Kramer, M. Hermann und G. Stryi-Hipp, „Measurements and benchmark of PV-T-collectors according to EN12975 and development of a standardized measurement procedure,“ in Proc. EuroSun 2010, Graz, Austria, 2010. Y. Tripanagnostopoulos, “Aspects and improvement of hybrid photovoltaic/thermal solar energy systems,” Solar Energy, vol. 81, pp. 1117–1131, 2007. T.T.Chow, “Performance analysis of photovoltaic-thermal collector by explicit dynamic model,” Solar Energy, no. 75, pp. 143–152, 2003. Industrial-Solar, [Online]. Available: http://www.industrial-solar.de/cms/en/company/ references/. [Accessed 24 May 2011]. IEA – SHC Task 01, “Investigation of the Performance of Solar Heating and Cooling Systems, 1977–1983,” [Online]. Available: http://www.iea-shc.org/task01/index.html. [Accessed 2 April 2010]. IEA – SHC Task 03, „Performance Testing of Solar Collectors, 1977–1987,“ [Online]. Available: http://www.iea-shc.org/task03/. [Accessed 2 April 2010].

3.5 Backup heater

/3.27/ /3.28/

/3.29/ /3.30/

/3.31/ /3.32/

/3.33/ /3.34/ /3.35/ /3.36/

IEA – SHC Task 10, “Solar Materials R&D, 1985–1991,” [Online]. Available: http:// www.iea-shc.org/task10/. [Accessed 2 April 2010]. IEA – SHC Task 27, “Performance of Solar Façade Components, 2000 – 2005,” [Online]. Available: http://www.iea-shc.org/task27/index.html. [Accessed 2 April 2010]. IEA – SHC Task 33, „Solar heat for industrial processes, 2003 – 2007,“ [Online]. Available: http://www.iea-shc.org/task33/index.html. [Accessed 2 April 2010]. IEA – SHC Task 43 (2009), TASK 43, „Task 43: Solar Rating and Certification Procedures – Advanced Solar Thermal Testing and Characterization for Certification of Collectors and Systems,“ 2009. [Online]. Available: http://www.iea-shc.org/task43/ index.html. [Accessed 2 April 2010]. CEN TC 312 – Thermal solar systems and components CEN-CENELEC, „Internal regulations, part 4, Certification, The Keymark system,“ [Online]. Available: ftp://ftp.cen.eu/BOSS/Reference_Documents/IR/CEN_CLC/ IR4_E.pdf. [Accessed 22 November 2009]. ESTIF, “Solar Keymark – The Quality Label for Solar Thermal Products in Europe,” [Online]. Available: http://www.estif.org/solarkeymark/. [Accessed 28 March 2010]. [Online] Available: www.solar-rating.org/. [Accessed 30 May 2012]. [Online] Available: www.cgc.org.cn/eng/news_show.asp?id=2. [Accessed 30 May 2012]. Lane G.A. (1986). Phase Change Materials, in Solar Heat Storage: Latent heat Material, Lane G.A. – editor Volume II, CRC-Press, Boca Raton, Florida.

85



Chapter 4

Heat driven cooling technologies: closed cycles Responsible Author:

Paul Kohlenbach, Solem Consulting, Germany Contributing Authors: Francesco Besana, Eurac research, Italy Uli Jakob, Green Chiller Association for Sorption Cooling e.V., Germany Mauricio de Lucia, University of Florence, Italy Harald Moser, Graz University of Technology, Austria Erich Podesser, Podesser Consulting, Austria Lars Reinholdt, Danish Technological Institute DTI, Denmark Hans Martin Henning, Fraunhofer Institute for Solar Energy Systems ISE, Germany Torsten Koller, ITW, University of Stuttgart, Germany Alexander Eichhorn, ITW, University of Stuttgart, Germany

Nomenclature/Units Symbols kW W kWh MWh COP Rt or ton

Indices Kilowatt Watt Kilowatt-hour Megawatt-hour Coefficient of performance tons of refrigeration

el th p r h

electric thermal peak refrigeration heat

This section deals with the most common types of thermally driven chillers, namely absorption chillers and adsorption chillers, and in particular, ones which are feasible for coupling with a solar thermal energy source. Desiccant cooling technology is another solution for solar assisted air-conditioning and is treated in a separate chapter. Furthermore, conventional vapour compression chillers are introduced since they may be used as a cold back-up source in a solar air conditioning system and they serve as a reference for comparison between solar-assisted and conventional systems. Conventional air-conditioning and refrigeration technologies employ a reverse vapour cycle (Clausius-Rankine) process to remove heat from a given application. Figure 4- 1 shows this ­refrigeration cycle based on cyclic compression and expansion of the refrigerant vapour to re-

87

4  Heat driven cooling technologies: closed cycles

move heat from a lower and reject to a higher temperature level. In this cycle, refrigerant is ­evaporated (1), compressed (2), condensed (3), expanded (4) and evaporated again (1), forming a closed cycle and providing continuous heat removal.

Fig. 4- 1 Conventional closed refrigeration cycle

Any refrigeration machine consumes energy to transfer heat from a source at a low temperature to a sink at a higher temperature. In the case of air-conditioning the heat extracted from the low temperature source is the useful cooling, i.e., the heat removed from the conditioned space is thereby producing the cooling effect. In the vast majority of air-conditioning applications, the intermediate temperature heat sink is the external environment and the heat is rejected to the external air. The driving energy is heat in the case of a thermally driven process and is mechanical energy in the case of a conventional refrigeration machine. In most cases the mechanical energy is delivered by a motor, as is in the case for building air-conditioning and refrigeration. The compression of the refrigerant vapour requires significant mechanical work and is usually done by an electric motor coupled to the compressor. A different approach is to use mostly heat for the compression of the refrigerant vapour, thus saving electrical energy, Figure 4- 2.

Fig. 4- 2 Thermally driven closed refrigeration cycle

88



Here, the mechanical compressor is replaced by a thermal compressor that uses heat and a smaller amount of mechanical work to raise the refrigerant pressure. Both refrigeration cycles described above are heat pumps, however with different temperature levels. As a result of the first law of thermodynamics, the flux of heat rejected at the intermediate temperature level, Qintermediate, is equal to the sum of the heat flux extracted from the low-temperature heat source, Qlow, and the driving power of the process, Pdrive, i.e.,

Q& int ermediate = Q& low + Pdrive Eq. 4- 1

For an electrically driven chiller, the driving power is the electricity input to the motor, Pel. In the case of thermally driven chillers, the driving energy flux, Pdrive, is a heat flux at a high temperature level, Qhigh. Figure 4- 3 shows the principle for the example of a thermally driven chiller.

Fig. 4- 3 Schematic diagram of energy flows in a thermally   driven refrigeration cycle

A key figure to characterize the energy performance of a refrigeration machine is the Coefficient of Performance, COP. For thermally driven air-conditioning systems, the COPthermal, which indicates the required heat input for the cold production, can be defined as follows:

COPthermal =

Q& low Heat flux extracted at low temperature level = Q& high Driving heat flux supplied to cooling equipment

Eq. 4- 2

The COPthermal varies with the equipment operation conditions, i.e., the three temperature levels, the percentage of load etc.; therefore COPthermal-values of different systems are only comparable if the same operation conditions are considered. For a conventional, electrically driven vapour compression chiller, the COPconv is defined as the required electricity input for production of cooling energy:

COPconv =

Q& low Heat flux extracted at low temperature level = Pel Electrical power supplied to the chiller

Eq. 4- 3

89

4  Heat driven cooling technologies: closed cycles

The following chapter explains the details and different technical applications of a thermally driven refrigeration process.

4.1

Principles of absorption and adsorption cooling

Commercially available heat driven cooling cycles use two different physical principles, absorption and adsorption. Both principles have in common the fact that one substance is taken up by another substance and that this process is reversible. Absorption is defined as a process where a liquid or gaseous material is taken up by the bulk volume of another liquid or solid material, e.g. a ʻspongeʼ that is taking up water. On the other hand, adsorption is a process that occurs when a liquid or gaseous material accumulates on the surface of another solid material, forming a molecular film. An example for adsorption is activated carbon as a filter for impurities in water. Both processes have in common the fact that heat is released when the sorption takes place. To reverse the process and release the absorbed or adsorbed material again, heat input is needed. The application of both processes in heat driven chillers is common and discussed in the following chapters.

4.1.1 Absorption chillers Absorption chillers contain two different substances, a refrigerant and a liquid absorbent. Their functionality is based on the different boiling points for the refrigerant and the absorbent, allowing the separation of the two using external heat input. The most common application is a single-stage chiller which implements the thermal cycle shown in Figure 4- 2. A more efficient cycle is the double-stage cycle which employs a dual compression process of the refrigerant solution. Both cycles are further discussed in chapter 4.1.2. The most commonly used pairs of refrigerant and absorbent in commercial absorption chillers are given in Table 4- 1.

Refrigerant

Absorbent

Water

Lithium bromide (LiBr)

Water

Lithium chloride (LiCl)

Ammonia (NH3)

Water

Tab. 4- 1 Common refrigerant/absorbent pairs in commercial absorption chillers

LiBr and LiCl absorption chillers are mostly used for air-conditioning applications with chilled water temperatures above 6 °C (43 °F). This temperature limit is set to avoid freezing of the refrigerant (water) in the evaporator. The ammonia-water chillers are mostly used in refrigeration systems designed for chilled water temperatures below the freezing point. Here, ammonia is the refrigerant which has a lower boiling point than water. Recent developments in absorption

90

4.1 Principles of absorption and adsorption cooling

r­ efrigeration have involved efforts on small capacity chillers (160 °C (>320 °F)) depending on the technology used (see chapter 4) and the actual operation conditions. As an example, with Tg = 90 °C (194 °F) = 363,15 K, Te = 9 °C (48.2 °F) = 280,15 K and Tc = 28 °C (82.4 °F) = 301,15 K, results (COP)ideal = 2,53. As mentioned in Chapter 4, also real sorption chillers show a dependence of the COPthermal on the operative conditions and particularly on Te,Tg and Tc. For example, a comparison of the COPthermal for the ideal case and the various types of sorption chillers, for a specific chilled water temperature (Te=9 °C (48.2 °F)) and cooling water temperature (Tc = 28 °C (82.4 °F)), is given in Figure 3. The picture shows that adsorption machines permit a relatively low driving temperature (55–80 °C (131–176 °F)), which is definitely interesting for solar applications. And the achieved COP values are in the range of 0.5 to 0.65. Single effect absorption chillers have generally higher COP values (0.6 to 0.75) but require slightly higher Tg (from 75–80 °C (167–176 °F)). Double-effect absorption chillers may achieve a better performance (COP @ 1.2) but, with driving temperatures of 120–160 °C (248–320 °F), they require high temperature collectors (e.g. CPC, tracking systems, etc.) for solar cooling applications.

Fig. 7- 3 Typical COP curves of sorption chillers, °F = °C × (9/5) + 32

180

7.2 Performance of thermally driven chillers

In general, since (1–Tc/Tg ) < 1, it is COPideal < EERideal, i.e. the ideal energy efficiency of a thermally driven chiller is always lower than that of a compression chiller for any given user temperature, Te, and heat sink temperature, Tc. However, such a comparison is not properly stated because the two definitions refer to energy of different qualities, namely heat (Qg) for the sorption machine and work (W) for the compression machine. A more rational basis for comparison is the Primary Energy Ratio (PER), defined as the ratio of the delivered useful cold (Qe) to the Primary Energy (PE) consumption to drive the machine, within a certain interval of time. The latter is normally used referring to the whole cooling season. t1

∫ Q&

e

PER =

⋅ dt

t0

PE to ,t1

=

Qe PE

to ,t1

Eq. 7- 7

t0 and t1 are the begining and the end of the cooling season, respectively. Further the PER expressions will always b refer to the whole cooling season if not otherwise specified. The PE in the currently accepted definition is “energy that has not been subjected to any conversion or transformation process” (e.g., fossil or nuclear) (EN 15603:2008 /7.3/). Thus for a compression chiller and a fuel fired thermally driven chiller it is possible to state the following formulas: Compression chiller

PER =

Qe Qe W = ⋅ = SEER ⋅ ε el , grid PE W PE

Eq. 7- 8

Thermally driven chiller

PER =

Qe Q Qg = e ⋅ = [COPthermal ]season ⋅ ε fuel ⋅ η heat source PE Qg PE

Eq. 7- 9

181

7  Energy and economic figures for solar cooling

Here, eel,grid is the efficiency of the electrical network system which depends on the electric structure of each country (for EU the average value of eel,grid is about 40%), while hheat source is the efficiency of the heat source (for instance a gas boiler) and efuel is the ratio between the energy content of the fuel and the primary energy needed for its production and transportation to the place of consumption. As will be seen later, using fossil fuel to provide the driving heat to a thermally driven chiller is not a reasonable approach from a primary energy point of view; the example is used here only to describe the formulas in a consistent way.

7.3

Energy performance of solar driven cooling systems

A solar thermal collector system of a solar cooling installation will, of course, also be used for space heating or DHW production, when required. However, in the following analysis only the energy performance of the cooling application is studied. In solar cooling applications the aim for the solar section is to provide the thermal energy necessary to power the thermally driven chiller or to regenerate the desiccant material in the case of a DEC system. The particular focus in this section is on the analysis of the energy performance of the chain from solar radiation down to delivery of cooling to the load.

7.3.1 Fractional PE savings For technological and economic viability reasons, in a solar heating and cooling plant the solar components are usually not designed to cover the whole user need, but only a fraction of it. The fraction of the total load which is covered by solar energy is referred to as the solar fraction (SF), which is usually expressed as a percentage. Two options can be used for covering the part of the cooling load exceeding the production of the solar assisted system: back-up on cold side, i.e. use of a secondary cold source, normally a vapour compression chiller, or back-up on the hot side, i.e., additional heat source which drives the thermally driven chiller. As will be shown further, the choice of one of the two options has a significant impact on the overall energy performance of the system. In case of a back-up on the hot side, the annual useful solar heat for the cooling application, Qg,sol, is given by

Qg ,sol = Qg ,tot − Qg ,back −up Eq. 7- 10

where Qg,tot is the total annual heat required for cooling, i.e., for operation of the thermally driven ­ bviously, cooling system and Qg,back-up is the annual heat from a secondary heat source (back-up). O

182

7.3 Energy performance of solar driven cooling systems

this formula only makes sense if a back-up heat source is employed to operate the thermally driven cooling process. Otherwise the cooling back-up is provided by a secondary cold source and thus the back-up heat, Qback-up, is zero. Further the use of a second heat source (as back-up) is analyzed. A simplified scheme of this solution is shown in Figure 7- 4.



Fig. 7- 4 Scheme of a solar thermally driven cooling system with a fossil fuelled backup heat source (TDC = thermally driven chiller)

The solar fraction for cooling, SFcool, this system design is then given by

SFcool = 1 −

Qg ,back −up Qg ,tot

Eq. 7- 11

In the previous sections the primary energy ratios, PER, were defined for electrically and thermally driven cooling machines (section 7.2). However, when looking at the energy performance of entire installations, all the auxiliary energy consumers such as pumps, fans, control units etc. have to be taken into consideration. Thus the total primary energy consumed to cover the cooling demand in a system following the scheme of Figure 7- 4 is composed of the primary energy demand of the fossil fueled back-up heat source, PEback-up, and all components that consume electricity:

183

7  Energy and economic figures for solar cooling

PE sol = PEback −up + PEel =

Qg ,back −up

ε fuel ⋅ η boiler

+

Eel

ε el , grid

Eq. 7- 12 3

where Eel and Qback-up are the consumption of electricity and heat from fossil fuels (if present), respectively. ePE,fuel denotes the primary energy efficiency of the fuel when used at the site of application and hboiler refers to the boiler efficiency. The primary energy factor, ePE,fuel for natural gas is less than 1 because of grid losses and the embodied primary energy for keeping the pressure of the natural gas network. In order to assess the advantages of a solar-assisted cooling plant in terms of energy savings it is necessary to compare the system’s performance with that of a conventional system. To this aim it is suitable to introduce the “Reference Design” concept, referred to as that cooling plant able to accomplish the same task as the solar cooling system, but with no use of solar energy. Usually the reference design is a conventional air-conditioning system based on an electric chiller with a given (SEER), which refers to the average EER of the entire cooling season. The primary energy consumption of such reference system, PEref, when completely based on electricity operation, is given by:

PEref =

Qe SEER ⋅ ε el , grid

Eq. 7- 13

The distance in primary energy consumption between the reference and the solar assisted system is a crucial design parameter and is measured by the “Fractional PE savings”, fPE,sav defined as:

f PE , sav = 1 −

PEsol PEref

Eq. 7- 14

where PEsol is the primary energy consumption of the solar cooling plant and PEref is that of the reference design operating under the same boundary conditions.

3 In systems using an electricity driven backup the consumption of the latter will be included in Eel and Qg,backup results zero.

184

7.3 Energy performance of solar driven cooling systems

7.3.2 Primary energy sensitivity analysis of solar cooling systems The energy performance of solar assisted cooling systems is crucial, as primary energy saving is the main reason for their application. Therefore the primary energy of a system according to Figure 4 will be analyzed in more detail and the sensitivity of the key parameters on the overall energy performance will be investigated. The main primary energy consuming subsystems are the backup heat system (e.g. gas boiler), the solar subsystem, the thermally driven chiller itself and the heat rejection subsystem (e.g. cooling tower). The primary energy demand of the fossil fueled backup heat source is given by:

PEbackup =

Qe ⋅ (1 − SFcool ) [COPthermal ]season ⋅η boiler ⋅ ε fuel

Eq. 7- 15

where Qe, denotes the seasonal cold production, [COPthermal]season the seasonal thermal COP of the thermally driven chiller, hboiler the energy efficiency of the backup heat source (e.g., gas boiler) and efuel the primary energy efficiency of the fuel used to operate the backup heat source. The electricity consumption of the back-up heat source is considered negligible. The components of the solar assisted cooling system which mainly consume electricity are the pumps operating in the solar collector field, the thermally driven cooling machine itself (e.g. refrigerant pump, control unit) and the heat rejection device (cooling tower). The primary energy of the electricity consumption (mainly pumps) of the solar subsystem is given by:

PEcoll =

Qe ⋅ f el ,sol ⋅ SFcool

[COPthermal ]season ⋅ ε el , grid

Eq. 7- 16

where fel,solar is the electric efficiency factor of solar heat generation and is defined as the fraction of the seasonal electricity consumption to operate the solar system, Eel,sol, and the seasonal solar gains used as driving energy of the cooling system, Qg,sol:

f el ,sol =

Eel ,sol Qg ,sol

Eq. 7- 17

185

7  Energy and economic figures for solar cooling

The primary energy of the electricity consumption of the thermally driven chiller also includes the electricity to operate the pump of the driving heat cycle and is defined by:

PETDC =

Qe ⋅ f el ,TDC

ε el , grid

Eq. 7- 18

Here, fel,TDC is the electric efficiency factor of the thermally driven chiller and is given by the fraction of the seasonal electricity consumption to operate the thermally driven chiller and the produced cold.

f el ,TDC =

Eel ,TDC Qe

Eq. 7- 19

Finally the equation for the electricity consumption of the heat rejection subsystem including the cooling cycle pump is given by:

PEreject

  1  Qe ⋅ f el ,reject ⋅ 1 +  [ ] COP thermal season   =

ε el , grid

Eq. 7- 20

Here, fel,reject is the electric efficiency factor of the heat rejection device (e.g. cooling tower) and is given by the fraction of the seasonal electricity consumption to operate the heat rejection device and the rejected heat. The rejected heat, Qreject, is given by the sum of the driving heat, Qheat, and the produced cold, Qe:

Qreject = Qg ,tot + Qe = Qg ,tot ⋅ (1 + [COPthermal ]season ) Eq. 7- 21

The summation of the equations for the different subsystems leads to a formula for the primary energy consumption of the solar cooling system, PEsol:

PE sol = Eq. 7- 22

186

Qe [COPth ]season

 1 − SFcool [COPth ]season ⋅ f el ,TDC f el ,reject ⋅ (1 + [COPth ]season )  SF ⋅ f + cool el ,sol + + ⋅ η  ε el , grid ε el , grid ε el , grid  backup ⋅ ε fuel 

7.3 Energy performance of solar driven cooling systems

In the following the sensitivity analysis was based on the equation for the fractional primary energy savings, fPE,sav according to equation fPE,sav, while the primary energy consumption attributable to the solar cooling system was calculated using the equation above. The reference values for the various parameters occurring in the formulas are given in Table 7 -2. The values used here are based on experiences made in solar thermally driven cooling systems which are so well designed as to achieve the expected operation performance. It has to be noticed that many realized installations do not achieve these values due to shortcomings in design, assembly, installation and operation.

Parameter

Meaning

Standard value

efuel

primary energy efficiency of fuel at application site

0.95

hboiler

energy efficiency of back-up heat source (gas boiler)

0.9

SFcool

solar fraction for cooling (i.e. for driving heat of the TDC)

0.9

eel,grid

primary energy efficiency of electric grid

0.4

[SEER]

seasonal EER of the reference cooling system (here: air cooled compression chiller)

3.0

fel,sol

electric efficiency factor of the solar subsystem, kWhel/kWhth

0.02

fel,TDC

electric efficiency factor of the TDC subsystem, kWhel/kWhcold

0.02

fel,reject

electric efficiency factor of the heat rejection subsystem, kWhel/kWhreject

0.03

Tab. 7- 2 Set of standard values for the various parameters in the sensitivity analysis

Figure 7- 5 shows that the solar fraction for cooling, i.e. the solar fraction of the generator heat to drive the thermally driven chiller has a strong influence on the fractional primary energy ­savings of solar cooling systems. In particular for systems with a low seasonal COP of the thermally driven chiller, the effect is significant and a too low solar fraction can even lead to a primary energy consumption higher than for the conventional reference. The reason is that a significant amount of cooling is produced by converting primary energy of a fossil fuel into cold with a low COP of the thermally driven cooling device and thus with a low overall conversion rate.

187

7  Energy and economic figures for solar cooling

Fig. 7- 5 Fractional primary energy savings as a function of the solar fraction for cooling for different values of the seasonal COP of the thermally driven chiller

Another important influence is that of the efficiency of the heat rejection system, as can be seen from Figure 7- 6. The amount of electricity needed for heat rejection strongly depends on the heat rejection technology used (see section 4.3.1). The seasonal amount of electricity for heat rejection can be kept low by using highly efficient components (pumps, fans) and employing speed controlled components which can be operated at reduced energy consumption under part load conditions.

Fig. 7- 6 fractional primary energy savings as function of the electric efficiency factor of the heat rejection subsystem for different values of the seasonal COP of the thermally driven chiller

188

7.3 Energy performance of solar driven cooling systems

Finally the fractional primary energy saving of course depends on the PER of the conventional reference system. Small room air-conditioners typically have SEER-values in the range of 2–2.5 or even below, while large centralized compression water chillers using wet cooling towers can achieve SEER-values up to 5 or even higher. The influence of this parameter on the fractional primary energy saving is depicted in Figure 7- 7.

Fig. 7- 7 Fractional primary energy savings as a function of the seasonal EER of the reference chiller for different values of the seasonal thermal COP of the thermally driven chiller

7.3.3 Other useful energy performance parameters A possible comparison with SEER of conventional cooling systems: the solar cooling COPelectric The overall electricity consumption of a solar assisted air-conditioning system, e.g. calculated or measured for an entire cooling season ranging from t0 to t1, can be put into relation to the overall amount of produced useful cooling for the same period of time. This figure is called electric COP, COPelectric, of the solar assisted cooling system: t1

∫ Q&

e

COPelectric =

to t1

∫P

el

=

Qe Eel

to

Eq. 7- 23

189

7  Energy and economic figures for solar cooling

Although this figure does not provide information on the primary energy saving of a solar cooling system it is very useful to compare different system designs and solar assisted cooling systems to conventional reference systems. It has to be noted that the figure includes the electricity consumption of all components such as pumps, control units, electric valves etc.. Therefore, for an adequate comparison with a conventional reference, also for the conventional reference the electricity consumption of these entire auxiliary components have to be taken into account. The COPelectric is used in chapter 10, chapter 11 and chapter 12 to assess installed systems based on measured electricity consumption. Best values of realized solar cooling installations achieve values of the COPelectric up to approximately 8 and target for future installations is to achieve values of 10 and higher. The Fractional primary energy savings A second parameter that already has been defined in section 7.3.1 (Equation 7- 14) and shall be highlighted here again is the fractional primary saving, fPE,sav, defined as:

f PE , sav = 1 −

PEsol PEref

Eq. 7- 24

where PEsol is the primary energy consumption of the solar cooling plant and PEref is that of the reference design operating under the same boundary conditions. This figure can also be used to assess the overall advantage of a solar assisted cooling system in terms of primary energy based on measured results. For this purpose the overall electricity consumption and the overall fuel consumption of a solar cooling system – or a solar heating and cooling system, respectively – have to be measured. Both amounts of energy have to be converted into primary energy using the valid conversion factors (see above); this yields the primary energy consumption of the solar cooling (and heating) system, PEsol. In order to assess the primary energy which would have been consumed by the conventional reference system, PEref, the measured overall outputs of the solar (heating and) cooling system, i.e. the measured useful cooling (and heating) have to be converted as well into primary energy. For this purpose the conversion factors for assumed conventional reference equipment (compression chiller, gas boiler) are employed in order to calculate the electricity and fuel consumption which would have been used by a conventional system. These values are then also converted into primary energy using the valid conversion factors. fPE,sav is calculated for some of the examples which are described in chapters 10 to 12, where enough reliable measurement data for a sufficiently long period of time were available. Good working systems achieve primary energy savings of 50 % and more compared to conventionally designed reference systems. 4 1m³/h=0.59cfm, 1 kW=0.28 ton, 1m³=264 gal, 1 m²=10.76 ft²

190

7.4 Environmental impact analysis

7.4

Environmental impact analysis

Environmental benefits of solar cooling technologies, especially in terms of primary energy savings and CO2 emission reduction, can be easily perceived throughout this book. However, in the engineering system analysis, not only process energy and its direct environmental burden has to be considered, but also the energy related to every single step of the process, from the search for raw materials and mining activities to shaping and assembling the components, up to delivering and installing the system as a whole, not to mention the materials for spare parts and maintenance, as well as the activities for disposal and/or recycling. Each process impacts on the environment in a different way and measure, for instance by depleting energy and material resources and releasing potentially polluting elements, with related consequences to human health and the natural environment. In literature, such an analysis is commonly referred to as Life Cycle Impacts Assessment (LCIA), and it is thoroughly described and regulated by the international ISO standards of the 14040 series. According to this procedure, the effect of a process on the environment must be quantified by means of appropriate impact indicators, such as Global Warming Potential (GWP), Acidification Potential (AP), Ozone Depletion Potential (ODP), and others. Following the ISO standards, the first stage of a LCIA, is the “Inventory Analysis”, i.e. the collection of data concerning the relevant inputs (raw materials and energy consumption) and outputs (emissions and wastes) related to the production, use and disposal of a product. Then follows the “Impact Assessment”, where mathematical and/or empirical models, based on physics, natural science or economics, are used to evaluate the contribution of every output of the Inventory Analysis to the impact categories. Different criteria may be used to judge and/or to optimize the design from an environmental point of view. Usually the optimum design arises from a multi-criteria analysis, which includes of course the economic issues. Going further with this topic is beyond the scope of the chapter. Yet a lot of scientific effort is being focused on this issue in recent times and, to the author’s best knowledge, the undergoing research and the literature existing so far (e.g. /7.4/–/7.8/) seem to prove that solar cooling technologies are promising not only in terms of energy savings but also in the more general environmental perspective.

7.5

Economic figures of solar cooling systems

The determination of the economic performance of a solar assisted air-conditioning system is another important feature which must be considered in order to assess the suitability of using this technology instead of conventional air-conditioning systems. The assessment of the economic performance must take into account the initial investment for all components, design, installation and commissioning, the annual costs for operation and maintenance as well as the final disposal

191

7  Energy and economic figures for solar cooling

cost after time of use. This section outlines the most important figures enabling a comparison among energy systems from the economic point of view. It includes an example to demonstrate the impact of different design decisions on the energy and cost performance. First cost – Capital investment Firstly, the initial investments have to be assessed, in order to calculate the economic viability of a solar cooling system. A detailed list of materials, controls, space and structural requirements, services and installation activities has to be set up to provide a complete overview of all cost items occurring as the initial cost. Cost records of previous installations of comparable design may be helpful as well as quotations submitted by manufacturers and contractors. For solar cooling installations the equipment costs are normally dominated by the cost of the solar thermal collector field and the thermally driven chiller, including heat rejection or the desiccant air handling unit. The typical breakdown of initial costs for two examples of solar cooling systems is shown in Figure 7- 8. The figure refers to a desiccant cooling system and an absorption cooling system both with flat plate solar collectors. The data derive from the research activity developed within the ROCOCO Project /7.9/ and were calculated from a detailed design and cost estimation for solar-assisted cooling systems to be installed in different types of building (hospitals, hotels, offices, trade centres and residential buildings) and in different climates (Southern, Central and Northern Europe and non-European aridhot and humid-hot climates). This study indicated that the cost fraction of the solar collector field including supporting structure ranges between 20 and 34 % of the total initial cost.

Fig. 7- 8 Example of composition of first cost for solar cooling systems (7)

For all major components specific costs, i.e., cost per unit of size or capacity, decrease with the installed size (m² for the collector field; kW for chillers; m³/h for desiccant air handling units, etc.).

192

7.5 Economic figures of solar cooling systems

Figure 7- 9 presents specific cost curves which are based on experiences from installations that were analysed during the work of Task 38 “Solar Air-Conditioning and Refrigeration” of the IEA Solar Heating and Cooling Programme. It has to be noted, that significant deviations from these cost curves may occur in a particular project depending on the chosen manufacturers, installer and other site related boundary conditions. In addition, most of the documented installations used to build the curves were demonstration projects rarely in pure commercial and competitive conditions with optimum pricing. These curves indicate a typical average value and show the typical specific cost decrease with growing component size or capacity. Moreover, for the solar collectors the cost value includes supporting structure and installation cost; piping and other hydraulic component costs (pumps, valves, heat exchangers, storage, etc.) are not included. For all other components the cost represents pure component cost without cost of the hydraulic circuit and other related components.

Fig. 7- 9 Typical cost curves (logarithmic x-axis) for key components of solar cooling systems4

193

7  Energy and economic figures for solar cooling

The overall cost of existing solar cooling installations, Ccost,specific, ranges between 2000 € up to 10.000 € per kW of cooling capacity. Similar figures emerge from the ROCOCO project activitiesalthough this project has been carried out earlier (2005–2007) /7.9/. This figure includes all initial cost items mentioned above, i.e. also the back-up components for both heating and cooling. As for the single components, the specific overall initial cost typically reduces with the system size. Several other parameters linked to the boundary conditions have an important impact on the initial cost, i.e., site dependent cost conditions, chosen components, particular system design, complexity of installation of the solar field, etc. Systems installed in existing buildings present higher overall specific costs compared to similar installations in new buildings. Another result of the ROCOCO study indicated that overall initial costs of a solar cooling installation are up to 50 % higher than the conventional reference systems in the case of an air cooling system (solar assisted desiccant air handling unit versus conventional air handling unit with vapour compression chiller for cooling) and 2–5 times higher than a conventional reference in the case of chilled water systems. Annual and Total Annual costs Annual costs, also referred to as operating and maintenance costs (O & M costs) are due to fuel, electricity, water, maintenance parts and services. Energy consumption is a large part of these operating costs; its incidence largely depends on the efficiency of the system, but also on the local price of gas and electricity. Electricity rates are usually more complex than those of gas or water. For instance the price may be different according to the season and the time of the day, or can increase when a consumption threshold is passed /7.6/. It has also to be taken into consideration that the price of energy will not be stable but can change during the operation lifetime of a system. For these reasons the energy costs must be assessed with care, usually on an annual basis. Further hidden costs, like insurance and property taxes, must not be forgotten. In some cases even income taxes might occur, when the owner of the energy system derives an income from the system operation. Once the annual operating costs, Cannual, and the initial cost, Cinitial, for both a solar-assisted cooling system (indicated with the subscript “sol”) and a conventional reference system (indicated with “ref”) are assessed, the most straight forward economic performance figure can be defined, namely the simple Payback Time (PT) (in years):

PT = Eq. 7- 25

194

Cinitial,sol − Cinitial,ref C annual,sol − C annual,ref

7.5 Economic figures of solar cooling systems

The PT is meaningful only if the more expensive system (the solar-assisted system, in this context) is also the one with the lowest annual operating costs. In this case, the use of the solar-assisted system may be regarded as economically convenient if PT is lower than the system technical lifetime (which normally is considered as 20–25 years). Actually, PT can be used only for a first basic analysis, but it may be misleading for a final decision about an investment. As the system’s cost evaluation requires comparison of cash flows at various points in time, methods are needed to account for the value of money over time. Various methods are used in order to provide a complete and correct total cost picture of a given investment during the lifetime of the system. Three major methods which are typically applied for cost assessment of energy systems are: –– The levelled cost of energy (LCoE) is typically used to provide a figure which expresses the total cost per unit of energy taking all cost items over the whole lifetime of the system into account. This figure is typically used to compare the cost of electricity production for different technologies and designs. It is less useful to compare systems that deliver energy in different forms; such as heating and cooling in the case of solar cooling systems. –– The calculation of the net present value (NPV) of an investment is a widely used method in discounted cash flow analysis taking into account the time value of money to appraise long-term projects. The method is typically employed if a series of future cash flows, both incoming and outgoing, have to be taken into account. Thus it allows to normalize cash flows at a defined time and provides a decision making tool for complex money transactions. –– Another method to be considered for the time value of money is the annuity method in which the initial investment (or initial cost) is spread over the life-time of the system (e.g., n years) in constant annual quantities. This method is typically applied to compare different alternatives which differ in initial cost and annual cost and where a comparison over the whole lifetime shall be made. Therefore this method is used further in this section. The constant annual quantities derived by the application of the annuity method are called annuities, A, and are obtained according to the relation:

A = Cinitial ⋅

ieff 1 − (1 + ieff ) n

Eq. 7- 26

where ieff is the effective rate of return, which can be estimated as the difference between the average interest rate offered by the bank and the inflation rate. By adding the annuity determined through Equation 7- 26 and the annual operating costs, one obtains the total (or overall) annual costs, Cannual,total, which allow direct comparison amongst different solutions:

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7  Energy and economic figures for solar cooling

C annual ,total = C annual + A Eq. 7- 27

The total annual cost is a comprehensive number which allows comparison of different solutions regarding their overall cost, since it considers the overall lifetime of the installation. However, in the previous definition disposal cost has been neglected since it is difficult to assess. Combined energy-cost performance During the design of a system, different alternative solutions should be compared, and also compared to a conventional reference, i.e. a standard solution using conventional energy sources such as electricity or natural gas. In most cases the energy performance calculation and assessment of total annual cost do not lead to the same optimum, i.e. the system design that leads to the lowest value of energy consumption does not necessarily lead to the lowest value of total annual cost. In order to combine cost and energy performance of solar cooling systems, a further parameter can be defined, namely the cost of saved primary energy, CPE,sav. It quantifies the additional costs per unit of primary energy savings with respect to a conventional reference system, and is given by:

C PE ,saved =

∆Cannual,total € [ ] kWhPE ∆PE

Eq. 7- 28

where DPE describes the annual primary energy savings of the solar assisted system in comparison to the reference,

∆PE = PE ref − PE sol Eq. 7- 29

and ΔCannual,total the difference of overall annual costs of the solar assisted system compared to the reference:

∆Cannual ,total = Cannual,total,sol − Cannual,total,ref Eq. 7- 30

This parameter leads only to reasonable results if the overall annual cost of the solar-assisted system is higher than that of the reference system. Otherwise the fraction does not make sense since not a cost per benefit is calculated, but a relation between two benefits (cost saving and energy saving, respectively) is computed.

196

7.5 Economic figures of solar cooling systems

An alternative combined cost-performance figure is the summation of the differences in cost and primary energy of the solar assisted system and the reference. However, in order to be able to do this summation both figures have to be converted to the same unit. This can for instance be done by multiplying each unit of saved primary energy with a corresponding cost value, fC,PE (€/ kWhPE). As a result the overall cost-performance indicator, FPE,cost, is defined as:

FPE ,cos t = f C,PE ⋅ ∆PE − ∆C annual ,total Eq. 7- 31

Example: a hotel in Madrid and Malta In the following a virtual design example is presented in order to illustrate the different performance figures and how they can be used to drive a decision in sizing the key system’s components. The general approach of this virtual study is shown in Figure 7- 10. Based on meteorological data (hourly values of a typical year), hourly values of heating and cooling demand were calculated based on dynamic building simulation. Also the demand for domestic hot water was computed because the solar system can also provide heat for hot water preparation. As a result an hourly file of meteorological and load data is computed which serves as an input for the system simulation. For the design of the solar assisted heating and cooling system a dynamic system simulation programme (such as TRNSYS) can be employed. The comparison of different design options can be implemented through parameter variation of the size of key system components such as the solar thermal collector field size, the heat buffer storage size and the thermally driven chiller capacity. Comparison with a conventional reference system allows calculation of savings in primary energy and differences in total annual cost.

Fig. 7- 10 Design approach used for the example of a hotel building

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7  Energy and economic figures for solar cooling

The example used here is a virtual hotel building located in Madrid with a low energy standard and with an overall floor area of 3050 m² (32,831 ft²); for dynamic building simulation the building was divided in 4 major zones (guest rooms, restaurant, kitchen, lobby including corridors etc.) and for each zone the load profiles were calculated. The key summary values of the annual energy demand and peak consumption for heating, cooling and domestic hot water of the entire building are summarized in Table 7- 3.

Peak demand

Annual demand

Full load hours

kW

W/m²

kWh/a

kWh/m²a

h/a

Domestic hot water

108.1

35.46

155439

51.0

1437

Heating

127.1

41.68

163428

53.6

1286

Cooling

114.2

37.44

77469

25.5

680

Heat for cooling (with assumed constant COPthermal = 0.63)

181.3

59.43

123252

40.4

680

Heat for heating and cooling

181.3

59.43

286680

94.0

1582

Heat for heating, cooling and domestic hot water

235.3

77.16

442118

145.0

1879

Tab. 7- 3 Key values of annual energy demand and peak consumption of the example building. 1 kW = 0.28 ton, 1 m² = 10.76 ft², 1 kWh = 3,412.14 BTU

In the parameter study different system combinations were compared: –– The collector size was varied between 100 m² (1,076 ft²) and 500 m² (5,382 ft²). –– The size of the thermally driven chiller was varied between a nominal capacity of 0 kW and 40 kW (11.4 ton). The size 0 kW corresponds to a solar thermal system which is only used for heating and domestic hot water preparation, but not for cooling. The thermally driven chiller is always covering part of the cooling load only and is backed up by a vapour compression chiller. –– The specific heat buffer storage size was varied between 30 litres per m² (0.74 gal/ft²) of collector and 80 litres per m² (1.96 gal/ft²) of collector. In the result figures shown below the best value of the storage size in terms of energy saving for each parameter set of collector area and thermally driven chiller size was always used. The operation strategy for the solar thermal system was assumed as follows: solar heat is used to cover heating loads – if there are any – first, since a low temperature heating system was assumed. Second priority is to cover domestic hot water demand and third priority is to operate the thermally driven chiller.

198

7.5 Economic figures of solar cooling systems

The cost curves for all key components shown above (Figure 7- 9) were used. In addition the values shown in Table 7- 4 for energy prices, other cost related parameters and primary energy conversion factors were used in the calculations. It should be noted that the results presented here originate from a theoretical simulation study. Cost values in a real project may be different due to local conditions. In particular higher cost for planning and installation may occur in case that companies are involved with little or no experience in solar heating and cooling technology.

Planning, installation and maintenance costs

Energy cost

Other parameters

Planning HVAC + solar thermal

% of sum of all components first cost

20.0%

Installation HVAC + solar thermal

% of sum of all components first cost

30.0%

Maintenance

% of total first cost p.a.

1.5%

Electricity

€/kWh

0.15

Peak electricity cost

€/kWpeak

50.00

Fuel

€/kWh

0.07

Increase rate of electricity cost

% p.a.

3%

Increase rate of fuel cost

% p.a.

3%

Lifetime

A

20

Interest rate

%

5.0%

PE factor electricity

kWhPE/kWhel

2.7

PE factor fuel

kWhPE/kWhfuel

1.1

Tab. 7- 4 assumptions for cost calculation and primary energy conversion. (1 kWh = 3,412.14 BTU)

In the following the energy performance and cost results of the study are presented. Figure 7- 11 shows the saved primary energy of the solar assisted systems in comparison to the conventional reference system. Here the primary energy saving is expressed as a fractional saving fPE,sav as it has been defined in section 7.3.

f PE =

PEref − PE sol PEref

Eq. 7- 32

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7  Energy and economic figures for solar cooling

Fig. 7- 11 fractional primary energy saving in % compared to the reference system as a function of the collector size for different values of the size of the thermally driven chiller, 1m² = 10.76 ft²

The fractional primary energy savings increase with the collector area. However, the curves show a typical saturation pattern which is due to an increasing number of hours in which not all of the available solar heat can be consumed from larger collector areas. The larger the collector area the larger is the energy saving effect of the use of a thermally driven chiller for covering part of the cooling load. As systems with a solar collector field with 500 m² (5,382 ft²) and a thermally driven chiller with 40 kW (11.4 ton) lead to additional primary energy savings of about 11 %. Since the energy load of a hotel is characterized by a large sanitary hot water demand the main energy saving effect of the solar collector system is due to covering the hot water load. The initial cost of all investigated systems is shown in Figure 7- 12. The initial cost lies in the range of 1500 € per kW of total cooling capacity for the smallest collector area (100 m²/1,076 ft²) and in the range of € 3100 per kW of total cooling capacity for the largest collector area (500 m²/5,382 ft²). A system without a thermally driven chiller and related equipment such as a cooling tower has an initial cost which is about 500 to 600 €/kW lower than for a system with a thermally driven chiller with a capacity of 40 kW (11.4 ton).

200

7.5 Economic figures of solar cooling systems

Fig. 7- 12 initial cost of the reference system and of all solar assisted systems as a function of the collector size for different values of the size of the thermally driven chiller, 1 kW = 0.28 ton, 1m² = 10.76 ft²

The simple payback time as defined in Equation 7- 25 for the different solar assisted systems is shown in Figure 7- 13. Based on this simple calculation all systems would amortize within their assumed lifetime (20 years). The lowest simple payback time – below 8 years – is achieved with the system with the smallest collector area (100 m²/1,076 ft²) and without a thermally driven chiller.

Fig. 7- 13 simple payback time (see equation 7- 25) of all solar assisted systems as a function of the collector size for different values of the size of the thermally driven chiller, 1 kW = 0.28 ton, 1m² = 10.76 ft²

201

7  Energy and economic figures for solar cooling

The total annual cost is shown in Figure 7- 14. In this Figure the total cost is shown as a fraction of the total annual cost of the particular solar assisted system and the conventional reference. Thus, a value of 100 % means that the total annual cost of the particular solar assisted system is equal to that of the reference system. As Figure 7- 14 indicates the lowest value of the total annual cost is achieved with a solar collector size of 200 m² (2,152 ft²) and a system that does not use thermally driven cooling. But we also see that most systems that use a larger collector area and a thermally driven chiller lead to lower total annual costs compared to the reference solution. A comparison between Figures 7- 14 and 7- 13 clearly shows that the simple payback time may be misleading as it results in another “best” system and in an underestimation of the real cost. The reason is that it does not include interest and neglects the time value of money.

Fig. 7- 14 total annual cost of all solar assisted systems as a function of the collector size for different values of the size of the thermally driven chiller for the hotel example in Madrid, 1 kW = 0.28 ton, 1m² = 10.76 ft²

Figure 7- 15 shows the results of a simulation for exactly the same hotel building but located in Malta. The results indicate that under the load conditions in Malta a larger solar collector area leads to minimal total annual cost values. Systems with a collector area of 300 m² (3,228 ft²) and a size of the thermally driven chiller in the range of 15 to 30 kW (4.2 to 8.3 ton) lead to a total annual cost of about 94 % of that of the conventional reference.

202

7.5 Economic figures of solar cooling systems

Fig. 7- 15 total annual cost of all solar assisted systems as a function of the collector size for different values of the size of the thermally driven chiller for the hotel example in Malta, 1 kW = 0.28 ton, 1m² = 10.76 ft²

For the final decision on a certain system design the two main objectives have to be considered: primary energy saving and cost. However, looking at the performance of the various solar assisted systems regarding primary energy saving and cost, leads to different results. The best system in terms of primary energy saving for the hotel example in Madrid is the one with the largest collector area (500 m², 5,380 ft²) and the largest capacity of the thermally driven chiller (40 kW, 11.4 ton) (see Figure 7- 11). On the other hand in Malta, the best system in terms of total annual cost is a system with a collector area of 200 m² (2,152 ft²) and no thermally driven chiller (see Figure 7- 14). Since many solar assisted systems lead to a lower total annual cost than the reference cost of saved primary energy, CPE,saved, is not an appropriate figure for combined energy-cost performance. Therefore the cost-performance indicator, FPE,cost, is used to compare the different solar assisted systems. The cost value of saved primary energy, fC,PE is assumed as 0.05 €/kWhPE. This value can be considered as the money equivalent of the ecological benefit of saved primary energy. Results of FPE,cost for the example hotel in Madrid are shown in Figure 7- 16 and results for the example hotel in Malta in Figure 7- 17. Based on this analysis the optimum system for the hotel example in Madrid has a collector area of 400 m² (4,304 ft²) and a thermally driven chiller with a capacity of 20 kW (5.7 ton). However, the maximum is not very pronounced and also somewhat smaller and larger systems lead to similar results. Nevertheless, this system seems to be a good compromise between cost and primary energy saving. The corresponding best results for the same hotel located in Malta are a collector area of 450 m² (4,842 ft²) and a size of the thermally driven chiller in the range of 30 to 45 kW (8.4 to 12.6 ton).

203

7  Energy and economic figures for solar cooling

Of course, the method described here is not the only one possible for the comparison of different solutions in terms of energy saving and cost. But it provides a consistent approach to translate results of a design study into a single overall performance figure which can be helpful in the selection of the sizing of the key parameters of a solar assisted heating and cooling system.

Fig. 7- 16 combined cost-performance, FPE,cost, of all solar assisted systems as a function of the collector size for different values of the size of the thermally driven chiller for the hotel example in Madrid, 1 kW = 0.28 ton, 1m² = 10.76 ft²

Fig. 7- 17 combined cost-performance, FPE,cost, of all solar assisted systems as a function of the collector size for different values of the size of the thermally driven chiller for the hotel example in Malta, 1 kW = 0.28 ton, 1m² = 10.76 ft²

204

7.5 Economic figures of solar cooling systems

/7.1/ /7.2/

/7.3/ /7.4/ /7.5/

/7.6/

/7.7/ /7.8/ /7.9/

EN 14511: 2007 : Air conditioners, liquid chilling packages and heat pumps with electrically driven compressors for space heating and cooling – Terms and definitions CEN/TS 14825:2008 : Air conditioners, liquid chilling packages and heat pumps with electrically driven compressors, fpr space heating and cooling – Testing and ratings at part load conditions. EN 15603:2008: Energy performance of buildings – Overall energy use and definition of energy ratings. August 2008. Ardente F., Beccali M., Cellura M., Lo Brano V.: Life cycle assessment of a solar thermal collector. Renewable energy, Vol 30, 2005, pag.1031–1054 Marletta L. Evola G., Sicurella F. : Energy and environmental performance of solar assisted desiccant units in a life cycle perspective. Proceed. Intern. Conference CLIMAMED, Genoa, Sept. 2007. Beccali M.: Life Cycle Assessment of solar cooling systems. A technical report of Subtask D, Dec. 2010, IEA Task 38. [15] Kalogiru S. A.: Environmental benefits of domestic solar energy systems, Energy Conversion and Management, 2004, 45, pp. 3075–3092 Kalogiru S. A.: Environmental benefits of domestic solar energy systems, Energy Conversion and Management, 2004, 45, pp. 3075–3092 Kalogiru S. A.: Thermal performance economic and environmental life cycle analysis of thermosiphon solar water heaters, Solar energy, 2009, 83, pp. 39–48 ROCOCO: Reduction Of COsts of solar COoling systems, Final Report 2008, Anita Preisler coordinator, Sixth Framework Programme.

205



Chapter 8

Overall system design, sizing and design tools Responsible Authors:

Daniel Mugnier & Romain Siré, TECSOL, France Contributing Author: Mario Motta, Politecnico di Milano, Italy

In the first planning phase of a solar assisted air conditioning system, the type of building to be equipped with the air-conditioning system and the use of the building should be analysed in detail. Attention should be paid to architectural and technical measures capable of reducing the cooling load, e.g., use of external shading, night ventilation in combination with the thermal inertia of the building and energy efficiency measures capable of reducing internal loads. Furthermore an annual load file should be prepared, containing at least hourly values of cooling and heating loads, with details on humidification and dehumidification requirements. As shown in chapter 2 of this book, the load structure of a considered building depends on the physical properties of the building and the thermal and solar gains from the use of the building, i.e. the frequency of occupation, occupation density and on the additional technical equipment in the rooms. The determination of a representative time series of loads is necessary since the correlation between solar radiation power and heating/cooling loads determines the energy demand from auxiliary sources as well as the utilisation of the solar thermal sub-system. Including these effects into the annual energy balance allows an assessment of primary energy savings, to choose appropriate sizes of the core components and thus to obtain the economic figures of the planned installation. The solar-assisted air-conditioning design process is a crucial step towards the development of a successful solar cooling project. During this process many questions need to be answered for the optimal use of the solar cooling production in the building. Issues to be considered include; technical features of the system, investment and running costs for the solar system, integration of the solar system into the main production and distribution system, primary energy savings, advantages and risks of the solution. This chapter aims to present the main steps in the design process and provides an overview of some necessary and existing tools. Different design methods are presented starting with the simplest ­qualitative approach, requiring minimum basic information and ending with references to more complex design tools and simulation programs.

207

8  Overall system design, sizing and design tools

As shown in Figure 8- 1, different methods are applied at successive phases of the project’s lifetime, requiring inputs and giving outputs with appropriate levels of detail. A short description of the method is presented: 1. Suitability analysis of a targeted building for a defined solar air-conditioning application. In the preliminary phase of a project, an assessment of the project’s feasibility can be carried out. This can, for instance, be done through the use of a “Check list” concept presented in the following sections. The applied method is qualitative and ranks the proposed project against a benchmark derived from the experience gained during the last decade in solar cooling 2. Selection of the most appropriate pairs of thermally driven cooling equipment and solar thermal collectors for the selected air-conditioning application. This selection can be carried out with qualitative methods, such as the decision scheme. The latter presented in the following sections, can help to approach the most appropriate pair, based on technical considerations (e.g., generation temperature required by the thermally driven technology selected, typology of distribution system, etc.) 3. Sizing of the solar collector field and other major system components with regard to energy and cost performance. This task, as explained above, requires an estimation of the load profiles to be used as input in detailed computer simulations. Differing form conventional HVAC systems, solar assisted air-conditioning plants are designed for a target solar fraction (SF) (see chapter 7). Therefore simulation output will include the annual energy balance of the considered application, which can be expressed by different performance figures (e.g., primary energy consumption) 4. Cost analysis: Based on the above mentioned energy performance figures, the last step of the design process corresponds to the costs calculation (investment, operation, maintenance, etc.) and savings (avoided fuel, etc.)

Fig. 8- 1 Summarized presentation of the predesign methodology developed within the IEA-SHC Task38

Within the IEA-SHC Task 38, a pre-design methodology has been developed. In the following this overall approach from a feasibility study to the detailed calculation for design purposes is described in more detail.

208

8.1 Suitability analysis of a targeted building for a ­defined solar air-conditioning application

8.1

Suitability analysis of a targeted building for a ­defined solar air-conditioning application

8.1.1 Presentation and objectives of the check-list A check list based on the feedback of European solar cooling experiences has been developed in the framework of the IEA-SHC Task 38. This method aims to assess, at the pre-design phase, whether a particular project seems realizable or not by giving a score between –20 and +20, and a conclusion linked to the score obtained. Moreover, the method warns the designer if issues related to crucial parameters occur. The development of this check list method is based on the experience carried out during the previous ten years of solar cooling project implementation. The crucial set of requirements is met by the vast majority of the successful solar cooling projects. They range from technical to economic and organizational issues. Some of the aforementioned characteristics are compulsory and others have less importance. Therefore the check list is a synthetic method capable of summarizing these requirements and the assessment of the project against each point in a single document. The result of the analysis is a score which provides an indication of the project’s suitability.

Fig. 8- 2 Interface of the check list tool

209

8  Overall system design, sizing and design tools

8.1.1.1 Check-list’s user guide Two different versions of the check list are available. The first version is available online /8.1/. The other one is a Microsoft Excel file /8.2/. The check list, see Figure 8- 3, is divided into 7 sections (7 single spreadsheets named “Home”, “Project”, “Technical Feasibility”, “Economic Feasibility”, “Operational Feasibility”, “Result” and “Calculation”): the first is an informative spreadsheet summarizing key information on the method. The sections from 2 to 5 require the user to enter detailed information on the particular building and other boundary conditions. Another section contains the calculation and finally the user can check the result and the conclusion in the “Result” section. Next to each question, a list of pre-defined answers is available (multiple choice). The user has to pick an answer for each question, or he has to leave the cell empty if the answer is unknown.

Fig. 8- 3 Example of the available pre-defined answers

The different questions are related to four issues: –– The “Project” section allows the user to fill in general information about the project, such as the project name, its location and a short description about it. –– In the “Technical feasibility” section the user is asked about technical information related to the building and its loads, such as climate, the space available for the collectors and for the technical equipment, the distribution system used, the correlation between solar energy and building loads, etc. –– In the “Economical feasibility” section the user is asked for economical information such as the cost of electricity and water, the financial conditions and the motivation of the building owner, potential financial grants, etc. –– In the “Organizational feasibility” section the user is asked information about operation, maintenance and monitoring issues. –– The last section, “Result”, shows the score of the project and the conclusion. A system of warnings is also implemented to detect if at any step some of the entered inputs correspond to situations which could cause the project to fail.

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8.1 Suitability analysis of a targeted building for a ­defined solar air-conditioning application

Example application of the Check list method: the INES installation in Chambery (Bourget du Lac, France) The application of the check list method will be presented through a real example. The solar cooling installation, object of the analysis was installed in 2009 for a research centre building (INES) close to Chambery (in Bourget du Lac, France).



Fig. 8- 4 Photo of the technical room of the Chambery installation

The system is composed of a 4.5 kW (1.3 ton) absorption chiller and a 30 m² (323 ft²) flat plate solar collector field, using glycol-water as the heat transfer fluid. The condenser circuit dissipates the medium temperature heat by means of a horizontal geothermal heat exchanger. The chiller, on the cold production side, is connected to a distribution loop supplying several fan-coils. A hot auxiliary backup is installed to warm up the water in the hot storage when the solar energy is not sufficient. The system is also used for space heating in winter (direct heating) Evaluation process The check list was filled in to assess the different parameters of the project: The “project” spreadsheet was filled in with the general information of the project, the answers given by the user here do not affect the final result, but they are useful to describe the general background. The second spreadsheet “Technical feasibility” was filled in according to Figure 8- 5. This sheet deals with specific and technical parameters about the targeted building and its loads.

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Fig. 8- 5 “Technical feasibility” spreadsheet of the checklist filled in for the Chambery installation

For the first question about the climate of the location, the two answers which lead to the highest score in this question are “Mediterranean climate” because of the good solar potential and the dry weather, and “Mountainous climate” because of the relatively long heating period and its good solar potential. As a consequence, the Chambery installation located in the Alps achieved the maximum possible score for this item. The next two questions are about the space available for the solar collector field and for the technical equipment. The Chambery installation is located in a large research building, so there is a lot of free space which can be allocated to the solar cooling installation. The maximum possible score is also reached for these two questions. The next two questions refer to the existing heating and cooling devices in the building. In the case of the Chambery project, there are no other heating or cooling devices in the building. As a consequence, the solar project has also to cover the costs of the distribution and of the backup system. The maximum score would have been attributed if the solar system was only added to an existing system. The next question deals with the passive and bio-climatic solutions around the solar project. This question reflects the fact that passive ways of cooling or heating a building have to be evaluated before a decision is made on the installation (solar driven active systems).

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The last question in the “Building” section asks for the experience of the installer and of the engineering office. Good experience for these two entities is decisive for the success of a project. It was the case for the Chambery example as some of the main solar cooling actors in France were involved in the design. The first three questions of the “Loads” section in the spreadsheet about “Technical feasibility” deal with the correlation between the solar irradiation and the loads; focusing on daily time patterns in the first question and on annual time patterns in the second question. For the Chambery installation, the yearly profile is favourable as there are no closing periods and cooling and heating loads occur all year long. However, because of the good thermal quality of the building, only part of the load is related to the sun on a daily level. The last question deals with the desired level of air-conditioning in summer. If a guaranteed cooling of the building is required (and not just provision for increased indoor comfort) a backup is compulsory because the gains of a solar system correspond to the fluctuating solar radiation and thus a system that relies completely on solar energy can not guarantee a fixed temperature in the building at all times. In the Chambery installation a backup has been installed so the installation is able to cool the building to a specified temperature. The third spreadsheet “Economical feasibility” for the example is shown in Figure 8- 6. This sheet is divided into two sections: the first is about energy costs and the second about the ­building owner features.

Fig. 8- 6 “Economical feasibility” spreadsheet of the checklist filled in for the Chambery installation

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The first section of the sheet aims at estimating whether or not the installation will be financially affordable once the system is working. The solar system will be more viable if the substituted energy is expensive and if water is cheap (or if the solar installation is not using water because of the use of a dry air cooler or geothermal heat exchangers for heat rejection). The Chambery installation does not use any fresh water and the price of the substituted energy is low in an average range; overall the operation cost saving is not very significant and thus does not play a main role in the financial affordability. The second section of the “Economical feasibility” spreadsheet focuses on the building owner. Indeed the motivation and the financial possibility of the building owner are crucial to achieve a successful project. The first and the last question of this section deal with the financial capacity of the building owner: the building owner has to finance the investments at the beginning of the project, and he has to have sufficient financial stability in order to wait for the subsidies which normally will be granted after the beginning of the project. This is the case in the example shown here as the research institute is covering the first cost of the Chambery installation. The two next questions consider the motivation of the building owner. This is also a very crucial point because the building owner should be the driving force of the project otherwise the chance of success of the installation will be significantly lower. In the case of the Chambery installation, the building owner is a research institute specialized in solar energy. Thus the building owner considers the solar cooling installation as a good way to promote the research institute. As a consequence, the building owner was very motivated and ready to take risks for the project, because the marketing impact is very important. Finally, the subject of subsidies has to be tackled in this economical section of the checklist. Considering the extra initial costs of a solar cooling and heating system in comparison to a conventional one, an important amount of subsidies have to be granted to make the project financially affordable for the building owner. This was the case for the Chambery project which was cofinanced by the European commission in the framework of a R&D project. Finally the spreadsheet “Organisational feasibility” has to be filled (see Figure 8- 7).

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Fig. 8- 7 “Organisational feasibility“ spreadsheet of the checklist filled in for the Chambery installation

Basically, the sheet is divided into two parts, the first one dealing with the operation and maintenance while the second with the monitoring. For the operation and maintenance topic, the Chambery installation is implemented in a research institute. As a consequence skilled staff is always present on site which can look after the solar system whenever a problem is noticed. This project gets the maximum points for these two questions. The monitoring is also very important to calculate installation performance, and to check if the installation is working properly. As a consequence a continuous monitoring of data should be implemented to design a successful installation and assure high performance during its operation. This is the case for the Chambery installation. Result After filling in these four previous spreadsheets, the “result” spreadsheet can be checked to obtain the final score of the project (see Figure 8- 8).

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Fig. 8- 8 “Result” spreadsheet of the checklist filled in for the Chambery installation

This project obtained 16.8 points out of 20 which is a very high score. As a result a solar cooling installation is recommended based on the results of the checklist. This result does not guarantee that in the further project development all conditions for a successful project will be achieved but it provides a helpful first hint about the potential feasibility.

8.1.2 Selection of the appropriate system technology: the SAC ­decision scheme The simplified decision scheme presented here has been developed within Task 25 of the IEASHC and updated within the most recent Task 38 IEA-SHC. It is meant as a tool to support the pre-selection of air-conditioning technologies, which can be used in combination with solar thermal systems. The following analysis deals with solar air-conditioning systems designed for comfort cooling in buildings; this analysis does not cover other applications such as industrial refrigeration systems. The thermal load of a building is characterised by the impact of climatic factors and building related factors. Depending on the particular climatic conditions, this load can be covered by particular solar driven cooling technologies (ab-/adsorption, DEC systems). The first attempt to assess which technology best fits a dedicated building can be guided by technical considerations.

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The goal of the presented decision scheme is to guide the decision for a certain technical solution for a given situation, defined by climatic conditions, building and occupation related factors as described above. Each depicted solution represents a technical solution for the use of solar thermal energy for building air-conditioning.

8.1.2.1 The decision scheme The decision scheme presented in Figure 8- 9 has been developed on the basic assumption that both indoor temperature and humidity have to be controlled. The starting point is the calculation of cooling loads based on the design case. In general, air-conditioning systems fall into one of four major categories: all-air, air-water, water and refrigerant-based systems. Each has their own unique technical and economic advantages in some situations. Depending on the cooling loads and the constraints linked to the typology of the end-user, either an all-air system, a water system or hybrid air/water systems can be employed for the extraction of heat and humidity from the building; there are no purely refrigerant based systems available which can be operated based on thermal driving energy.

Fig. 8- 9 Basic scheme for decision guidance

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The starting point of the decision scheme is linked to the basic technical decision whether or not the minimum air change rate (which is the minimum air volume flow for building ventilation capable of complying with the quality requirements for indoor air) is sufficient to also cover the cooling loads (sensible and latent). This will typically be the case in rooms/buildings with a requirement for high ventilation rates, such as lecture rooms. However, a supply/return air system (which is a mechanical system providing the building with fresh air from outside while removing exhaust air from the building) makes sense only in a tight building, since otherwise the leakage through the building shell is too high. In cases of supply/return air systems both thermally driven technologies are applicable, i.e., desiccant systems as well as thermally driven chillers. In all other cases only thermally driven chillers can be used in order to employ solar thermal energy as the driving energy source. In the application of desiccant techniques (see chapter 5 for more information) in extreme climates, i.e., climatic conditions with high humidity values for the ambient air, special configurations of the desiccant cycle are necessary in order to be able to employ this technology. More items of the design which are not covered by this method are as follows: –– Necessity of a backup system for the cold production or permitting solar autonomous operation of the solar assisted air conditioning system; –– Flexibility in comfort conditions, e.g. permitting certain deviations from the desired air states; –– Economical issues; –– Availability of water for humidification of supply air or for cooling towers; –– Comfort habits for room installations: fan coils have the lowest investment cost, but allow dehumidification only when connected to a drainage system; chilled ceilings and other gravity cooling systems require a high investment cost, but provide higher comfort. The type of thermally driven chiller applied is not indicated here. A basic technical scheme of a system which contains both open desiccant cycles and closed cycle water chillers is shown in Figure 8- 10. Different options of backup are shown in Figure 8- 10, namely back-up on the heat side by other heat sources (e.g. gas burner, connection to a district heating network, co-generation plant, etc.) and a back-up compression chiller. To provide cooling in the room several solutions are possible: a fan-coil system which is used in summer and winter, a radiative cooling system such as chilled ceilings or a ventilation system providing fresh air which is cooled and dehumidified.

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Fig. 8- 10 Scheme of a complete system including desiccant technique and thermally driven water chiller

8.1.2.2 Solutions for different conditions In the following all possible paths in the scheme shown in Figure 8- 9 are briefly described and the technical solutions are presented. Thermally driven chiller supplying a chilled water network When the installation of a centralized air handling unit is not feasible or desired, the only technical solution to use solar thermal energy for building air-conditioning is to use a thermally driven chiller which supplies chilled water to a chilled water network. An example might be an office building from the building stock which does not have the available space for installation of an air duct system. Independent of the climate a low temperature of the chilled water (approx. 6–9°C, 43–46°F) is required in order to allow air dehumidification. A technical solution for such a system is shown in Figure 8- 11. This system uses solar collectors and other heat sources to run a thermally driven chiller. Cold water is used to supply a fan-coil system. Solar gains in winter can be used for heating using the same indoor units.

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Fig. 8- 11 Pure water-based system using a thermally driven chiller

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Thermally driven chiller supplying a chilled water network and a air handling unit It is important to assess whether or not it is possible to cover cooling loads with the air flow needed to comply with air quality requirements in a given building. This building whould have then a specific planned and installed central ventilation system. If this is not the case another system to purge sensible loads has to be installed alongside the air handling unit. A further decision depends on the building tightness: in a very tight building a supply and return air system can be installed and ventilation heat recovery becomes possible. However, in a building which is not sufficiently tight the installation of a supply/return air system is problematic since either external air is sucked into the building (internal pressure lower than external) or is lost through the building shell (internal pressure higher than external). In such a case an air handling unit to provide only fresh air would be installed. Fresh air is cooled and dehumidified and sensible loads not covered by the fresh air must be covered by other means. An example might be a chilled ceiling system. The path for such a configuration and an example of a technical realization are shown in Figure 8- 12. The example shown consists of a solar collector field and a back-up heat source to run a thermally driven chiller. The chiller supplies chilled water to the supply air handling unit and to a chilled ceiling. Dehumidification is realized in the air handling unit delivering supply air. A control valve controls the inlet temperature to the chilled ceiling in order to avoid condensation. When the air flow needed to comply with air quality requirements is sufficient to also cover the sensible loads this results in a similar system solution but without a chilled ceiling. Desiccant cooling system and chilled water system for temperate climates In a tight building which is equipped with a centralized ventilation system, application of a desiccant cooling system is possible. However, the exact design of the desiccant system depends on the climatic conditions: in mild conditions, i.e. a temperate climate, a standard desiccant cooling cycle will be sufficient to provide enough dehumidification in all cases. In a more humid climate the standard desiccant cooling cycle has to be adjusted in a way to cope with the higher humidity of outdoor air. In the path and configuration shown in Figure 8- 13 the required ventilation rates for air quality are not high enough to cover all sensible loads and an additional chilled water system must be installed. The example shown in Figure 8- 13 is composed of a solar collector field and other heat sources as back-up to run a desiccant cooling system. The desiccant system shown here uses evaporative cooling in both supply and return air. Since the required ventilation rates for air quality are not sufficient to cover cooling loads an additional vapour compression chiller is used to operate a chilled ceiling system.

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Fig. 8- 12 Thermally driven chiller that provides chilled water to a supply air system and to a chilled water system (e.g. chilled ceilings)

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Fig. 8- 13 Desiccant system and an additional conventional water chiller to supply a chilled water system (e.g. chilled ceilings)

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Desiccant cooling system and chilled water system for extreme climate In the case of a climate with a high humidity of the ambient air (denoted extreme climate) desiccant technology can be used, but another configuration of the desiccant system is necessary in order to cope with the high humidity of external air. A possible realization of a system corresponding to the path is shown in Figure 8- 14. Conventional air handling unit with supply/return air and additional chilled water system An alternative to the application of a solar operated desiccant system is to use a solar thermal driven water chiller and to operate a conventional supply/return air handling unit in cases of tight buildings with centralized ventilation systems. If the air change rate needed to comply with air quality requirements is not high enough to cover all sensible loads again an additional system to purge sensible loads is necessary. In this case, which is shown in Figure 8- 15 no distinction has to be made between moderate and extreme climatic conditions, since in both cases the same air handling unit can be used because dehumidification is realized through cooling the supply air below the dew-point. The example system in Figure 8- 15 is composed of a solar collector field and other heat sources as a backup which are used to operate a thermally driven chiller. The chiller supplies cold water to the air handling unit and the chilled ceiling. Dehumidification is realized in the air handling unit by cooling the supply air below the dew-point. The air handling unit is equipped with an evaporative cooler in the return air stream which in combination with the heat recovery wheel allows pre-cooling of the fresh air. A control valve controls the inlet temperature to the chilled ceiling in order to avoid condensation. Desiccant air handling unit in temperate climate In the case of a tight building with an air exchange rate needed to comply with air quality requirements which is sufficient to cover the sensible loads, a pure air system is possible. This might be the case in a well designed building in which energy saving equipment is used, a highly efficient shading device is implemented, artificial lighting is minimized by using day-lighting concepts and a high thermal mass leads to reduced temperature peaks. Other means like night ventilation support the reduction of peak cooling loads. Another example of such a building is a seminar room with a high occupation rate; in such a room the required fresh air amount is rather high due to the high occupation and may be high enough to purge sensible loads completely. In the case of a temperate climate with relatively low humidity of the ambient air a ventilation cycle desiccant cooling system is a possible solution to use solar thermal energy for building airconditioning. The corresponding path and a technical solution are shown in Figure 8- 16. The system in Figure 8- 16 is composed of a solar collector field and other heat sources as backup which are used to operate a desiccant cooling system with evaporative cooling in supply and return air. No other system to provide cooling is necessary. Solar energy can be used in winter for the pre-heating of air and for supplying heat to the radiator heating system.

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Fig. 8- 14 Desiccant system designed for humid climatic conditions and an additional conventional water chiller to supply a chilled water system (e.g. chilled ceilings)

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Fig. 8- 15 Supply/return air system using a conventional air handling unit and an additional chilled water system, both operated with a thermally driven chiller

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Fig. 8- 16 Desiccant air handling unit in a pure full air system (moderate climate)

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Desiccant air handling unit in extreme climates As already outlined in the “Desiccant cooling system and chilled water system for extreme climates” section, the “normal” desiccant cycle is not applicable under conditions of high humidity of the ambient air. In such cases another design of the desiccant system is required. A scheme of a system for the corresponding path in the design scheme is shown in Figure 8- 17. The example in Figure 8- 17 consists of a solar collector field and other heat sources as back-up to operate a desiccant cooling system, which is adapted to climates with high values of humidity of the external air (denoted “extreme climate”). The desiccant system shown in Figure 8- 17 differs from the one shown in Figure 8- 16 since two cooling coils which are connected to the water chiller are installed on the fresh air side. Such a configuration is a promising concept in climates with very high humidity ratios of the ambient air. The first cooling coil serves a first dehumidification and cooling. Since the dehumidification starts at a high humidity level, the required cold water temperature is relatively high (approx. 15–18 °C (59–64 °F)). Final dehumidification to achieve the desired supply air humidity is realized with the sorption wheel. The second cooling coil is used for controlling the supply air temperature. Pure full air systems with a conventional air handling unit An alternative to the application of a solar operated desiccant system is to use a solar thermally driven water chiller and to operate a conventional supply/return air handling unit in cases of tight buildings with a centralized ventilation system. Since in the case shown in Figure 8- 18 the air exchange rate needed to comply with air quality requirements is sufficient to cover the sensible cooling loads (see description of section “Desiccant air handling unit in temperate climates”), no additional cooling devices in the room are necessary. A possible configuration is also shown in this Figure. No distinction has to be made for moderate and extreme climates. The system shown in Figure 8- 18 consists of a solar collector field and other heat sources as backup which are used to operate a thermally driven chiller. Cold water is used to operate the conventional air handling unit. The air handling unit is equipped with an evaporative cooler for the return air which in combination with the heat recovery wheel allows pre-cooling of the fresh air.

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Fig. 8- 17 Desiccant air handling unit in a pure, full air system (extreme climate)

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Fig. 8- 18 Full air system using a conventional air handling unit and a thermally driven chiller

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8.1.3 Selection of the proper type of solar collectors for the selected air-conditioning system and thermally driven cooling equipment 8.1.3.1 Introduction The solar thermal collector system is one of the most essential sub-systems, because the efficiency of the whole system depends on its design and efficiency. Several different types of solar thermal collector can be used in solar cooling and air-conditioning systems. In the following the combination of solar thermal collectors to thermally driven cooling technologies is discussed in greater detail. Table 8- 1 provides a general overview on the most important heat driven cooling technologies based on sorption processes and their needed driving temperature range. In the final line of the table the suitable solar collector technologies that can be employed are shown. Figure 8- 19 shows the required driving temperature level as a function of the useful temperature lift for different closed cycle sorptive cooling technologies (single-effect, double-effect, triple-effect). Different examples of system configurations for small, medium, high and very high temperature lifts are given. This figure is based on a fundamental thermodynamic analysis and does not represent particular market available thermally driven chillers.

Tab. 8- 1 Overview of market available thermally driven cooling systems based on sorption technology and the corresponding driving temperature range and the corresponding suitable solar collector technology, °F = (9/5) °C + 32

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Fig. 8- 19 Required driving temperature for single-effect, double-effect and triple-effect thermally driven chillers as function of the useful temperature lift, °F = (9/5) °C + 32

8.1.3.2 Solar collectors for solar cooling application A general overview on solar collectors is given in chapter 3 of this handbook. A first indication on the suitability of a certain type of solar thermal collector for a given thermally driven cooling system is achieved by assessing the collector efficiency in the operation temperature range of this thermally driven cooling technology. This is demonstrated here for the example of flat plate collectors and evacuated tube collectors. A general overview of flat plate solar collectors is given in section 3.2.1 of this handbook. Depending on the application, flat plate solar thermal collectors are installed on a simple supporting structure to provide optimum tilt and orientation, but they may also be integrated into a sloped roof, which is advantageous from an architectural point of view. Most solar thermal collectors used in temperate climates have a selective coating, which has a high absorbance in the visible range of the solar spectrum but a low emissivity in the infrared range. The influence of different coatings is shown in Figure 8- 20. The grey shaded area represents the typical operation range of an absorption chiller.

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Fig. 8- 20 Typical efficiency curves for different types of flat-plate collectors; the grey area denotes the typical operation range of an absorption chiller (x = (Tcollector – Tambient) / Solar irradiance, for a given irradiance value)

Due to their performance characteristics, selective flat-plate collectors can be used effectively in combination with desiccant cooling systems, adsorption chillers and possibly single-effect absorption chillers. However, only high-quality solar thermal collectors with a selective absorber coating are suitable because of the typically high required driving temperatures for this kind of cooling technique. A general overview of evacuated tube collectors is given in section 3.2.3 of this handbook. Depending on the application, evacuated tube type solar thermal collectors are installed on a supporting structure to provide optimum tilt and orientation, but they may also be mounted on a flat roof and each single tube oriented to an optimal angle. A real roof integration of this type of solar collector is not possible in the large majority of cases and products. Typical efficiency curves of evacuated tube collectors are shown in Figure 8- 21. The grey shaded area indicates a typical operation range of an absorption chiller.

Fig. 8- 21 Efficiency curves for typical evacuated tube collectors; the grey area denotes the typical operation range of an absorption chiller (x = (Tcollector – Tambient)/Solar irradiance, for a given irradiance value)

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8.2

System sizing

8.2.1 Guidelines Different guidelines have been established to help in the sizing of a solar air-conditioning system. The following expression gives an indication of the necessary collector area to be installed for a solar air-conditioning system:

Aspec =

1 G ⋅η coll ,design ⋅ COPdesign

Eq. 8- 1

with Aspec (m²/kWcold) G (kW/m²) η coll, design (–) COPdesign (–)

= specific collector area per installed kW thermally driven chilling capacity = irradiance at collector surface at design condition = collector efficiency at design condition (driving temperature of the thermally driven chiller) = thermal COP of chiller at design conditions

As an example, for an irradiance of 800 W/m², an efficiency of the collector of 50 % and a thermal COP of 0.65 of the thermally driven chiller at design conditions the specific collector is calculated to be 3.85 m² per kWcold (145.8 ft²/ton). The advantage of this simple method is that it allows a quick assessment of the required collector area. However, this method is not sufficient for a serious design of the collector size since it neglects the influence of the variation of radiation on the collector during day and year. Furthermore, any information on the specific site and load is neglected and the method neglects part load conditions in the cooling load and the part load behaviour of the thermally driven cooling equipment. In /8.3/, other simple sizing methods are introduced. The method shown above considers only the power produced by the collector under nominal conditions. However, solar thermal collectors do not produce a constant power, but a variable power which depends of the variation of the irradiance incident on the collectors due to variation of radiation conditions. Therefore, to determine the energy produced by a solar thermal collector, information on climatic data is needed. With hourly values of irradiance incident on the collector and the knowledge of the collector efficiency curve and collector incidence angle modifier it is possible to calculate the maximum energy produced by the collector at a fixed working temperature. References /8.4//8.5/ give the necessary methodology for this calculation. From this methodology the annual gross energy produced by the solar thermal collectors is obtained, Qgross, which can be expressed in kWh/m² (1 m² = 10.76 ft², 1 kWh = 3,412.14 BTU).

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The annual cost of the heat produced by the solar thermal system – Solar heat cost, Costheat – can then be calculated by:

Cost heat =

Cost annual Qgross

Eq. 8- 2

where Costannual is the annualized cost of the solar collector and is calculated by

Cost annual = Cost spec f annuity Eq. 8- 3

fannuity is the annuity factor that takes the interest rate of the invest­ment and the lifetime of the collector system into account and Costspec is the collector specific cost, based on information of current solar thermal systems installed, i.e., the collector cost per collector area. Advantages of this method are: –– It allows a good comparison of different solar collectors using their parameters and the radiation data of a specific site. –– The maximum possible heat production of a specific solar collector for a given site (annual meteorological data file) and a given constant operation temperature (e.g. nominal driving temperature of the particular thermally driven chiller) is determined. Disadvantages of this method are: –– Any information about the load profile is neglected. –– The method neglects part load conditions of the cooling load and part load behaviour of the thermally driven cooling equipment. –– For the calculation of Qgross some software tools, usually for the design of solar thermal sys­tems for hot water preparation, can be used.

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8.2.2 Simple pre-design tools Relatively simple pre-design tools are available for free download, which have been developed in the framework of different European Projects. Table 8- 2 provides reference to two of these tools; this list does not claim for completeness.

Software

Reference / Source

SACE – Solar Cooling Light Computer Tool

Reference /8.6/ http://www.solair-project.eu/218.0.html

ODIRSOL

Reference /8.7/ Software available on demand by writing to TECSOL at [email protected]

Tab. 8- 2 List of example of simple pre-design software tools

A simple description of each of these tools is given here focusing mainly on their capabilities and limits of application. SACE – Solar Cooling Light Computer Tool The SACE Solar Cooling Light Computer Tool is a fast and easy-to-handle computer tool for studying the feasibility of using solar thermal energy for building air conditioning. It has been developed in the framework of the EU project SACE Solar Air Conditioning in Europe. This tool can be downloaded from the SOLAIR European project website: http://www.solair-project.eu/218.0.html General Scheme of use The tool has the function of carrying out a draft feasibility study in order to investigate the application of solar cooling for a given load file and a given solar collector. The tool uses a combined meteo-load-file as input (produced with TRNSYS or any other similar program) and a configuration file for the definition of the system. Figure 8- 22 shows a picture of the main user interface of this tool. Short description of the method The purpose of this method is to provide a draft assessment of a reasonable value of collector area and storage size for a given building in a given climate (meteo-load system) and for a given solar collector. The SACE tool gives a result as a table and an example is shown in Table 8- 3. These results are valid for the profile of an office building, located in Madrid. Meteorological data were produced using the Metenorm software /8.3/.

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Fig. 8- 22 Main window of the SACE – Solar Cooling Light Computer Tool

Tab. 8- 3 Example of a SACE result file: Solar fractions for heating, cooling and heating+cooling are calculated for different collector areas and storage sizes based on the simulation results.

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Based on the meteo-load-file and configuration of the system, the annual solar fraction for heating and cooling is calculated based on an hour-by-hour comparison of needed heat for a thermal driven cooling and available solar heat. A parametric study for different collector areas, expressed as specific collector area AA (m² of collector per m² of conditioned room), is automatically carried out. In order to assess the effect of storage, different sizes of a system integrated energy storage, expressed in h of peak cooling load, are considered. The storage is not modelled as a technical component, but simply as a buffer which enables the use of excess solar heat (or cold) which is stored in one hour to be used in later hours in which the solar gains are not sufficient to match the load. A storage size of 1 hour means, that the peak cooling load of the entire year can be covered by the full storage for one hour without any other energy input into the system. In a real system, storage can be integrated either on the hot side (storage of excess solar heat), on the cold side (storage of excess solar produced cold) or on the load side, e.g. using the building thermal mass. The method allows for a fast assessment of achievable solar fractions for heating and cooling as a function of the collector and storage size. The results have to be considered as upper limits of potential solar fractions since no thermal losses within the systems are considered. ODIRSOL – Solar Assisted cooling Software The ODIRSOL software has been developed in partnership between CSTB and TECSOL. It aims to be a decision tool for designers and planners. The tool is based on dynamic simulations with TRNSYS, in order to provide a technical and economical assessment of a detailed solar cooling project using single effect absorption chillers. Simulations cover simple configurations with hot and/or cold back up and/or hot and/or cold storage. An online “help” is available for most of the steps, as well as examples of projects of several sizes. The hourly results of a yearly load simulation of the building to be cooled and heated have to be provided by the user. All the data used in the program are presented in a data base included in the software. The content of the database is mostly coming from French available components and products collected by TECSOL during the period 2005–2007. Meteo database is made of stations with long term average French data.

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Content and method

Fig. 8- 23 Main user interface of the ODIRSOL program

The first step consists in selecting the hydraulic configuration; 4 choices are available, with or without storage. Then the user has to provide hourly meteorological and cooling load data under the required format from a TRNSYS PREBID study, or coming from any other dynamic building simulation which produces hourly load data. The user has also to provide some basic geographical data. From these data, the software will automatically pre-size all components of the installation. The pre-sizing method is based on simple ratios, guidelines and on a database of commercially available products. In the main screen, the user has the possibility of modifying the specifications of each component, to define a specific one or to choose other components in the database (see Figure 8- 24).

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Fig. 8- 24 Main screen of the ODIRSOL software

When the overall equipment has been chosen, the user starts the simulation. The TRNSYS core runs a simulation based on 30 minutes time steps. The component models mainly belong to a standard library and a CSTB supplemental library. The absorption chiller is a standard model modified by TECSOL using operating curves provided by manufacturers. At the end of the simulation, the software provides a written report including energy balance across all loops and financial assessment of the solar cooling project. The user can also look at temperature profiles at different positions in the system (mainly component input and output temperatures). Monthly data files are also produced in plain text format. Status The final ODIRSOL software is freeware but is, at the date of writing this book, delivered in a Beta test format. The program can be downloaded from the internet at (http://software.cstb.fr/soft/ download.asp?page_id=us!Odirsol#). All configurations have been tested in detail and provide realistic results. However, some are not validated in comparison with real installations because of a lack of reference monitoring results. The software is now available only in French and the database need to be enlarged to cover recent commercially available products.

240

8.2 System sizing

8.2.3 Detailed simulation tools Detailed simulation tools have the ability to model the building and the HVAC system either separately or combined. Table 8- 4 lists the software packages that are most commonly used for transient simulation of solar heating and cooling systems; the list is not exhaustive. Software

Reference / Source

TRNSYS

Reference /8.8/ www.sel.me.wisc.edu/trnsys/

TRANSOL 3.0

Reference /8.9/ http://www.aiguasol.coop

EnergyPlus

Reference /8.10/ www.eere.energy.gov/buildings/energyplus/

PolySun

Reference /8.11/ http://www.velasolaris.com

INSEL

Reference /8.12/ http://www.inseldi.com/index.php?id=21&L=1

Tab. 8- 4 List of detailed simulation software tools; the list does not claim for completeness

Some of these simulation tools put their main focus on the HVAC system and have an additional possibility of modelling the building and/or generating load files (heating and cooling). In this case we consider them as “System oriented”. Others put their main focus on the detailed modelling of the building energy needs and in addition provide the active components of HVAC Systems. In this case we consider them as “Building oriented”. A short description of these software tools is given in the next sections. System orientated software packages TRNSYS TRNSYS is a very well know software tool in the solar energy community. It was developed in the 1970s at the University of Wisconsin. Subroutines describing the components of solar systems, hydraulic components and HVAC are called TYPES. TRNSYS already includes TYPES for many solar heating system components and also for HVAC system components. It is mostly open source and allows for the inclusion of new components. TRNSYS also has a special editor for definition of the building characteristics which allows the calculation of heating and cooling loads. The source code of the building model type is not open. Many examples of its use for design of solar assisted air conditioning systems are available; some are described in /8.13/.

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8  Overall system design, sizing and design tools

A graphical interface is available which allows creating the connections between the different components in hydraulic circuits. An example of the graphical user interface is shown in Figure 8- 25.

Fig. 8- 25 The TRNSYS program interface shown for an example of an absorption solar cooling system

TRANSOL 3.0 – Solar thermal systems simulation software TRANSOL is designed to be an easy-to-handle simulation tool for solar thermal systems. It has been developed jointly by the French Scientific and Building Technology Centre (CSTB) and the Spanish engineering company Aiguasol. Currently the tool includes more than 40 different solar thermal configurations programmed with TRNSYS. These configurations allow the simulation of solar cooling systems, and it is also possible to simulate the system with heating loads and domestic hot water loads. A demonstration version of the software is downloadable via the websites: http://www.aiguasol.coop or http://software.cstb.fr. TRANSOL has been developed with the aim of bringing the detailed dynamic simulation tools to the design process of solar thermal systems, providing the accuracy and the rich features of the more detailed models, but without the need for time and knowledge in developing the systems in a fully featured simulation platform like TRNSYS. TRANSOL combines a simple user interface with detailed TRNSYS models containing the Meteonorm 6.0 meteorological database on a worldwide level.

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8.2 System sizing

To facilitate the task, it includes a step by step wizard that guides the user sequentially through each screen of the system (climate, solar collector, collection …) and carries out a pre-dimensioning of elements based on data previously entered by the user. In addition, to ensure a correct set of parameters during the simulation, some correlations are performed to resize the components each time the user modifies any of the system characteristics (for example, if the user changes the number of collectors, TRANSOL, by default, updates the flow of the primary and secondary pumps, the pipe diameter of the circuits, the volume of the storage, etc). This facilitates the pre-sizing, since it is not necessary to update all system parameters to ensure a consistent simulation. The main user interface of the TRANSOL software is shown in Figure 8- 26.

 Fig. 8- 26 Main window of the TRANSOL 3.0 tool

The list below outlines some specifications of the TRANSOL program: –– It permits calculation of the cooling and heating loads with the TRNSYS building model (type 56, developed by Transsolar [Transsolar]), or by entering the loads already calculated by another program using an external file.

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–– –– –– –– –– –– –– –– ––

Systems can be defined that either provide sensible cooling only or sensible and latent cooling. It allows parametric studies of a previously defined system, and allows comparison of the results. It contains a large and expandable component database. Shading can be taken into consideration. It includes a collector model with inertia. It includes a bi-dimensional characterisation of the incidence angle modifier IAM and provides specific models for flat plate collectors and evacuated tube collectors. Customised reports displaying results possible. It permits reporting of hourly values on an annual basis for the temperatures, flow rates, most important powers, working times of pumps and auxiliary systems. Available languages are: Spanish, Catalan, French, English and Italian

INSEL® – Integrated Simulation Environment Language This software is described as an “integrated environment and a graphical language for the creation of simulation applications”. It uses graphical symbols that are interconnected by mouse operations and which can represent mathematical functions or real components of different systems, e.g. solar thermal collectors. The present version of INSEL already has a toolbox for Solar Thermal evaluations available that includes collectors for the heating of liquid fluids such as flat plate collectors and evacuated tube collectors, it considers solar air collectors, storage tanks and also models for solar thermal cooling plants, like desiccant and evaporative cooling systems as well as absorption cycles. This software also has a user-programmable environment in which other component models can be built and integrated. Programming languages like FORTRAN and C/C++ are supported. An example of the user interface is displayed in Figure 8- 27. A learning edition is available for free download from the internet (http://www.inseldi.com/index.php?id=21&L=1).

244

8.2 System sizing

Fig. 8- 27 The INSEL program interface

PolySun Developed by Vela Solaris, PolySun (http://www.velasolaris.com) includes different groups of software models for solar thermal, photovoltaic, geothermal, and solar cooling systems. The “PolySun Solar cooling” tool can be combined with “PolySun Solar thermal” and “PolySun Heat Pumps”. The software is designed to calculate the total energy consumption, as well as the energy consumption of each single element of a solar cooling system. In addition, it contains features for performance optimization, for performing comparisons with conventional cooling and heating system solutions and for optimising the operation strategy. The user can run a simulation by using an example configuration available in the program or by building its own configuration. A large catalogue of materials with their properties is available in the software for each component of the system; the user can also add devices directly to the library. An example of the user interface of Polysun is shown in Figure 8- 28.

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Fig. 8- 28 The PolySun software interface

Building orientated software package Energy Plus EnergyPlus is described as “an energy analysis and thermal load simulation program”. It is based on a user’s description of a building from the perspective of the building’s physical make-up, associated mechanical systems, etc. EnergyPlus calculates the heating and cooling loads necessary to maintain thermal control set-points. Some of the main characteristics of this software, selected from the detailed description in /8.10/, are: –– Integrated, simultaneous solution where the building response and the primary and secondary systems are tightly coupled (iteration performed when necessary). –– Sub-hourly, user-definable time steps for the interaction between the thermal zones and the environment; variable time steps for interactions between the thermal zones and the HVAC systems (automatically varied to ensure solution stability). –– ASCII text based weather, input, and output files that include hourly or sub-hourly environmental conditions.

246

8.2 System sizing

––

––

Heat balance based solution technique for building thermal loads that allows for simultaneous calculation of radiant and convective effects at both, the interior and exterior surface during each time step. Loop based configurable HVAC systems (conventional and radiant) that allow users to model typical systems and slightly modified systems without recompiling the program source code.

An example of one of the available user interfaces of EnergyPlus is displayed in Figure 8- 29.

Fig. 8- 29 Example of a user interface of the EnergyPlus software

EnergyPlus is distributed as freeware and can be downloaded from the internet (www.eere.energy. gov/buildings/energyplus/). It has already incorporated several HVAC system components as well as solar thermal collectors. The physical models of the components are described in detail in the reference (Energy Plus, 2005). It will be possible to include new models. Guidelines for this are given as well in the reference document. /8.14/ includes reference to other simulation tools.

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/8.1/ /8.2/ /8.3/ /8.4/

/8.5/

/8.6/

/8.7/ /8.8/

/8.9/ /8.10/

/8.11/ /8.12/ /8.13/ /8.14/

248

Tecsol, “http://www.tecsol.fr/checklist/,” [Online]. Socol&Tecsol, 2011. [Online]. Available: http://www.tecsol.fr/RafrSol2/downloads/ IEA-SHC-T38_TECSOL_Check-List_SolarCooling.xls. H.-M. Henning, Solar-Assisted Air-Conditioning in Buildings – A Handbook for Planners, 2 ed., Springer Wien/New York, 2008. P. Horta, M. J. Carvalho and M. Collares Pereira, “Long-term performance calculations based on steady-state efficiency test results: Analysis of optical effects affecting beam, diffuse and reflected radiation,” Solar Energy, vol. 82, pp. 1076–1082, 2008. P. Horta, M. J. Carvalho and S. Fischer, Solar thermal collector yield – experimental validation of calculations based on steady-state and quasi-dynamic test methodologies, EUROSUN 2008, 2008. Fraunhofer Institute for Solar Energy Systems ISE, “SACE Solar Cooling Light Computer Tool – Guidelines for Use,” 31 August 2003. [Online]. Available: http://www.solair-project.eu/218.0.html. A. Le Denn und D. Mugnier, „SOLAIR Guidelines,“ 2009. [Online]. Available: www.solair-project.eu. Solar Energy Laboratory at Univ. of Winsconsin, Madison, USA, “TRNSYS – Transient system simulation environment,” 2010. [Online]. Available: http://sel.me.wisc.edu/trnsys/. M. Proville und I. Garruchaga, „SOLAIR Guidelines, Spanish version,“ 2009. [Online]. Available: www.solair-project.eu. US Department of Energy, Energy Efficiency and Renewable Energy, “Energy Plus Manual – version 1.2.2,” 2005. [Online]. Available: www.eere.energy.gov/buildings/energyplus/. A. Le Denn und M. Gutierrez, „MeGaPICS deliverable L11 – State of the art of the existing calculating tools,“ 2010. Doppelintegral GbR, “INSEL – INtegrated Simulation Environment Language,” 2003–2006. [Online]. Available: www.inseldi.com. E. Wiemken et al, “Design and planning support for solar assisted air-conditioning guidelines and tools,” Eurosun, 2004. M.J. Carvalho, “WP 4.5: SOLAR COOLING: Contribution to a future development of CTSS method applicable to solar assisted air conditioning systems (or solar cooling systems)”, 2007 [Online]. Available: www.swt-technologie.de/WP4_D2.5-cooling.pdf

Chapter 9

Solar thermal system design Responsible Author:

Wolfgang Streicher, University of Innsbruck, Austria

9.1

Field configuration parallel/series, high/low-flow

9.1.1 General characteristic of high/low-flow systems In most cases a collector field connects several single collectors. These collectors can be either connected in series or in parallel. Combinations of these two forms are also used. To assure turbulent flow and therefore a high heat transfer rate in each collector tube, the specific volume flow through each collector col (given in litres per hour and per m² of collector) should be kept above a certain rate. On the other hand the volume flow should not be too high in order to avoid unnecessary pressure drops and therefore high electricity demand for the circulation pumps. A single collector itself consists either of several parallel collector tubes (as shown in the modules of Figure 9‑ 1) or of one tube that meanders over the whole module. Connecting all collectors in parallel yields a low temperature increase ΔTcol = ΔTtot. The volume . flow through the whole collector field Vtot is the sum of the volume flows through each collector . ­Vcol. Therefore this layout is called High-Flow (see right side of Figure 9‑ 1). . By connecting the collectors in series the total volume flow through the whole collector field Vtot . is the same as the volume flow through each collector Vcol. As the total volume flow in a series connection is lower than the total volume flow in a parallel connection, this layout is called LowFlow. The temperature rise in the collector field ΔTtot is naturally higher than in the High-Flow layout, as the temperature increases of each collector (ΔTcol) are summed. If the solar radiation is high enough the Low-Flow layout supplies hot water quickly even if the hot water storage is cold. The disadvantage is a higher thermal loss to the environment, which is due to the higher temperature difference in the rear collectors to the ambient. The pump electricity demand Pel de. creases due to the lower level of total volume flow in the collector loop Vtot, but increases due to the higher pressure drop Δptot in the collectors. This also depends on the pump efficiency ηel at the operating point.

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9  Solar thermal system design

Pel =

V&tot ⋅ Atot ⋅ ∆ptot η el ⋅ 1000 ⋅ 3600

Eq. 9‑ 1

. with Vtot A tot

total specific volume flow in total collector area in

[l/(m²col,tot·h)] [m² (1 m² = 10.764 ft²)]

These relations are illustrated in Figure 9‑ 1. The high pressure losses of collectors in series could be partly compensated by using enlarged diameters in the collector loop piping. If possible, the pressure drop in the collector itself could be reduced. Obviously different pumps have to be selected for High-Flow and Low-Flow systems. Collectors connected in series have a more regular flow through the collector area due to a higher driving pressure drop over the field. The hydraulic layout has to be adjusted to the total volume flow. Low-Flow systems normally use stratification units in the tank in order to prevent mixing the hot temperature from the collector fluid with colder temperature at the inlet of the store (see section 9.3). An equal flow resistance in all parallel lines should be achieved to obtain the same volume flow in all parallel pipes. This can be partly achieved by a “Tichelmann” arrangement, where the added length of the inlet and outlet pipes is the same for all parallel collectors (Streicher, 2008).

Fig. 9‑ 1 Hydraulics for the collector field for Low-Flow serial (left) and High-Flow parallel (right) systems /9.1/, 1 l/(m²h) = 0.047 GPM/ft², DT(°F)=DT(°C)x(9/5)

250

9.1 Field configuration parallel/series, high/low-flow

An alternative or an addition to the Tichelmann approach is to use balance valves which permit the compensation of the pressure drop difference between loops, actively controlling the balance of the system after a long operation time (dust inside the pipes) and above all have a possibility of measuring the pressure drop through the valves. Finally, for large systems with significant collector piping diameter, the cost of balance valves is similar (or even cheaper) to the avoided piping that would be needed for a Tichelmann approach.

9.1.2 Heat needs of solar cooling systems Single effect thermally driven chillers need an inlet temperature between approximately 60 to 95 °C (140–203 °F) and a temperature difference between flow and return of typically (6–10 K, 11–18 °F), because of a low useful temperature difference in the generator of the cooling machine. Therefore the collector layout for cooling should be High-Flow operation at a high temperature level.

9.1.3 Heat needs of domestic hot water and space heating preparation Domestic hot water (DHW) production requires a temperature lift from around 10 °C, 50 °F (cold water inlet) up to 45–60 °C, 113–140 °F (hot water outlet). This is a temperature lift of 35 to 50 K, 63–90 °F. Solar thermal systems for the DHW production can be designed therefore as Low-Flow systems. The advantage is that the necessary temperature lift for domestic hot water preparation can be reached in one single step in the solar collector if the solar radiation is high enough. Therefore the auxiliary heater has to be switched on less often. Such a design requires proper stratification in the tank in order to profit from the achieved large temperature lift and avoid mixing of the hot water coming from the collector with cold water in the storage. However, if the solar system is used only for preheating with a low solar fraction even during the summer, High-Flow systems with higher collector efficiency due to lower collector temperature can also be chosen. The same is true for systems oversized for the summer load. The concept is such that the maximum temperature is achieved in the heat store every day, no matter which flow system is chosen. Solar assisted space heating (SH) is usually operated in low temperature heating systems in order to achieve a high efficiency in the solar collectors. Therefore the needed temperature lift is between 5 and 15 K (9 and 27 °F). On the other hand the return temperature from the heating system is not lower than 20 °C (68 °F). Compared with the requirements for DHW production a High-Flow system should be chosen for space heating systems. Many existing combined solar systems (solar assisted DHW and SH production), also known as solar combi-systems, work with the Low-Flow principle in the collector loop. This strategy

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o­ utperforms the High-Flow systems in most cases. Obviously the hydraulics of the systems have to take Low-Flow into account and minimize temperature losses by mixing valves or incorrect inlet positions at the heat store (see section 9.3).

9.1.4 Possible layouts and control strategies for collector fields for solar cooling systems with DHW and SH production (solar combi-plus-systems) The combination of a conventional solar system for DHW and SH production with a thermally driven chiller may cause problems, because the temperature difference between the flow at the supply and return of the hot side of the chiller is typically small (6–10 K (11–18 °F)) and the return temperature of the desorber of the cooling machine to the store is around 60–80 °C (140–176 °F). The collector inlet temperature thus ranges from 70–90 °C (158–194 °F) by taking into account a temperature loss of the heat exchanger and the heat storage between solar loop and desorber. Operating the collector in Low-Flow operation would induce a very high collector outlet temperature of up to 130 °C (266 °F) in normal operation. This means the collector loop has to be operated above 3 bar (43.5 PSI) absolute pressure at the collector and above 4 bar (58 PSI) in the basement (assuming 10 m (33 ft) height difference between storage and collector) in order to avoid evaporation in the collector. Additionally the efficiency of the collector is not optimal at such high temperatures as well as the antifreeze (mostly propylene-glycol) and the corrosion inhibitors tend to degrade faster at these operating conditions. Therefore High-Flow operation is the desired flow regime during cooling operation which, however, is in contradiction with the desired Low-Flow for operation in domestic hot water and heating mode. One option to overcome this dilemma is to switch the collector loop operation between Low-Flow in winter and High-Flow in summer. Turbulent flow for a good heat transfer between absorber and collector fluid and acceptable pressure drops to assure a unique flow through all parallel collector pipes has to be achieved for both operational modes. In principle there are four possibilities to switch between High- and Low-Flow: The collector hydraulics stay the same in summer and winter, but the volume flow through the collector varies by switching the pump speed or the type of pump. Two cases are possible –– the collector loop is defined for Low-Flow and a high pressure drop in the collector loop occurs during High-Flow operation, that results in a high electricity demand of the collector pump, –– OR the collector loop is designed for High-Flow and during Low-Flow operation the flow regime in the collector is laminar, which reduces the collector efficiency by a few percentage points and may additionally result in non-uniform flow in the parallel collector pipes.

252

9.1 Field configuration parallel/series, high/low-flow

Following the example shown in Figure 9‑ 1 High-Flow has about four times the volume flow of Low-Flow and the pressure drop is about 16 times (4²) higher in High-Flow than in Low-Flow (according to Bernoulli’s law with Δp ~ (ρv²)/2), where r denotes the fluid density and v the fluid velocity. The electricity demand is theoretically up to 64 times higher (Pel = (Δp)/ η ~ v³). By switching between High- and Low-Flow, two different pumps, for High-Flow and Low-Flow mode, have to be chosen (high Δp and high volume flow versus low Δp at low volume flow). If only one pump is used, mostly one of the operating points (either High- or Low-Flow) have a corresponding lower pump efficiency. The second possibility is to drive the collector constantly in between High- and Low-Flow. This assures equal flow in the collectors but gives slightly less solar yield during both the heating and cooling season. A third option would be to drive the collector field only in High-Flow and adjust the hydraulic system for DHW and SH, which provides less useful solar gains in winter but makes the plant cheaper. Another possibility is to change the hydraulics of the solar collector for cooling and non-cooling operation. This can be done by installing switching valves, which connect more collectors in series for Low-Flow operation and more collectors in parallel for High-Flow operation. Probably again two pumps are required: one for a high volume flow and a low temperature difference and one for low volume flow and a high pressure difference. Beside these options a compromise is also possible by installing some of the collectors in parallel and the remaining in series. Such a system then operates with a medium temperature lift and a medium volume flow. For the example shown in Figure 9‑ 2 a design with 2 collectors in parallel and 2 in series is shown. Then the operation range of a variable speed pump would be sufficient to change the volume flow e.g. by a factor of 2 depending on the actual operation regime (cooling, heating, DHW).

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Fig. 9‑ 2 Hydraulics for the collector field for Medium Flow systems, 1 l/ (m²h) = 0.047 GPM/ft², DT(°F) = DT(°C) x(9/5)

If only space heating and solar cooling is required for a system (e.g. office buildings with low DHW demand) both demands have only a small temperature difference between flow and return but at different temperature levels. However, both demands normally do not occur at the same time. In such cases the solar system can be operated in High-Flow.

9.2

Stagnation of solar plants

9.2.1 Stagnation in collector fields Most common collector circuit layouts are subject to periods of stagnation when the collector pump is switched off due to the store being fully charged. This is more common in solar combisystems than in solar domestic hot water systems, because the collector area is normally oversized for the summer DHW load. A thermally driven cooling machine used as heat sink in summer reduces the stagnation time of the collector, because of a higher summer load. Nevertheless, stagnation may occur in spring and autumn, where the solar radiation is high, but no space heating or cooling demand occurs and in off-times due to other reasons (e.g. weekends in office buildings, holidays in educational buildings, etc.).

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9.2 Stagnation of solar plants

Other possible reasons leading to stagnation include: –– Electricity supply failure –– Some parallel collector circuit pipes are not in operation due to a blockage by air in pipes –– Partial evaporation may occur due to high temperatures in the collector field for solar cooling and uneven flow distribution in parallel collector modules. Modules with lower volume flow may reach the evaporation temperature in the collector outlet. The produced steam leads to an even higher pressure drop in the respective collector module, because of its far higher volume compared to the liquid heat carrier. Therefore the volume flow becomes even lower and the evaporation temperature is reached earlier in the collector module producing even more steam. In this way this collector module is “blocked” by the produced steam. In many systems the pressure in the collector circuit is kept below 3 bars. During stagnation the collector fluid evaporates and is forced down into the expansion vessel. Stagnation temperatures depend highly on the type of collector and the fluid in the collector circuit as shown in Table 9- 1.

Saturated steam pressure Stagnation temperature

Water

Water / Glycol (70% / 30%)

°C (°F)

bar (PSI)

bar (PSI)

Swimming pool collector (no cover)

90 (194)

0.7 (10.2)

0.6

Non selective absorber (single glass cover)

120–150 (248–302)

2.0–4.8 (29–69.6)

1.7 (4.2)

Selective absorber (single glass cover)

160–200 (320–392)

6.2–15.6 (89.9–226.2)

5.4–14.1 (78.3–204.5)

Evacuated tube- and evacuated flat plate collector

250–300 (482–572)

40–86 (580–1247)

> 25 (> 363)

Tab. 9- 1 Stagnation temperatures for different types of collectors

The occurrence of collector stagnation can be divided into several phases: 1. Heating of the collector fluid to the boiling temperature 2. Evaporation of the collector fluid 3. Superheating of steam in collector 4. Condensation of steam in pipes after the collector is near thermal equilibrium 5. Re-condensation of fluid in the collector, when the solar radiation decreases and the temperature in the collector drops below the boiling point of the fluid again Figure 9‑ 3 shows two possible hydraulic schemes which indicate how to deal with the evaporation of collector fluid during stagnation.

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Case 1: In the left scheme of Figure 9‑ 3 the increasing vapour pressure forces all liquid out of the collector as the increased pressure pushes the liquid down equally on both sides. Therefore no vapour occurs in the connecting pipes from the collector to the store. Case 2: If the pressure is equal on both sides all liquid in the “U” has to be evaporated because it is “trapped” in the collector (right scheme of Figure 9- 3 [case 2]). The steam is forced into the pipes and has to be condensed in the heat exchanger to the heat sink. Very high temperatures occur in the whole collector circuit. There is an increased degradation of the propylene glycol and the corrosion inhibitors, because of the high temperature in the collector. Therefore case 1 is recommended for the collector layout.



Fig. 9‑ 3 Hydraulic collector flow scheme pressing steam out of the collector during evaporation (case 1 left and case 2 right) /9.2/.

A similar situation as in case 2 is shown on the right side of Figure 9‑ 4. The collector loop and the elements in the pump group (including the expansion vessel, a check valve, pressure gauges of the hydraulic flow scheme of the collector circuit) must allow the liquid to be drained from both sides of the collector to the expansion device (Figure 9‑ 5). Again the liquid can be drained out of the collector by the evaporated steam. Special attention has to be paid to the position of the check valve. It will block the backflow from the collector into the expansion vessel, if positioned in the wrong place (Figure 9‑ 4, left). If fluid is blocked in the collector, more steam will be produced and a water hammer may occur.

Fig. 9‑ 4: Inappropriately placed check valve is preventing the fluid from the collector to empty on both sides (left) and an example of a large collector field with good emptying behavior (right) /9.2/.

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9.2 Stagnation of solar plants

Collector stagnation may be reduced or even avoided in consideration of the following: Additional heat sink (swimming pool, boiler as heat exchanger to chimney, etc.) –– Layout only for summer demand (small plants but stagnation possible during i.e. resident’s vacations) –– Cool down period during night –– Drain back systems. This technology is particularly suitable for solar cooling systems because it represents a passive solution to avoid evaporation of collector fluid even during electricity failure /9.4/. This technology is also useful in protecting the collector against freezing risks. A detailed know-how and management of the installation process of the solar collector field and the collector loop is needed.

9.2.2 Implications of stagnation on the solar pump group Several manufacturers deliver complete pump groups (Figure 9‑ 5), which also include the expansion vessel, a check valve, pressure gauges and combined thermometer-manometers, isolation or closing valves, filling valves, emptying valves and safety valves.



Fig. 9‑ 5 Elements and positioning of the elements in a pump group /9.1/

The pump group serves the following functions: –– Filling and emptying the loop with propylene glycol/water mixture. –– Release of collector fluid through the safety valve if necessary. –– Check valve, which protects against unwanted night-time backward circulation due to buoyancy effects and allows draining/expansion of the collector fluid from both sides into the expansion vessel in case of stagnation. –– Possibility of changing the pump without releasing the whole collector fluid (two isolation or closing valves on both sides of the pump). –– Placement of the safety valve and the expansion vessel in such a way that it is not possible to disconnect them from the collector loop.

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––

Ensure the possibility of volume expansion and collector emptying in stagnation without release of fluid by means of the expansion vessel

In the example shown on the left of Figure 9‑ 5 the expansion vessel also protects the pump effectively from low pressures at the suction side that may cause cavitation and destroy the pump by neutralizing any pressure changes on this side.

9.3

Stratification and necessary hot water storage tank volume

9.3.1 Heat input from solar collectors to the heat stores Heat losses between the solar collector and the heat storage have to be reduced to a minimum. Temperature loss can occur either with too little heat transfer area or with temperature loss by mixing. Different heat input options should be used for High-Flow and Low-Flow options in order to avoid mixing: –– An internal heat exchanger at the bottom of the store slowly heats the complete storage volume. A High-Flow solution for the collector is suitable for such an arrangement, because there is mostly only a low temperature gradient around the heat exchanger in the store. A high temperature lift is not appropriate in this case. –– Another High-Flow solution with some stratification heats the upper part of the store with one internal heat exchanger in the upper part of the store and subsequently heats the bottom of the store (if solar energy is still available) with a second internal heat exchanger. So both heat exchangers are operated one after the other (see Figure 9- 7, left). –– If both internal heat exchangers are connected in series a higher temperature gradient occurs in the store around the heat exchangers and therefore a low flow operation of the collector field should be chosen (Figure 9- 6 a). –– Using stratification devices, which naturally have outflow height with the same temperature in the store as the inlet fluid from the collector loop, a high temperature lift in the collector can be transferred to storage without mixing. This can be either done by an internal heat exchanger connected to a stratification device and operated in countercurrent flow to the cold storage water (Figure 9- 6 b) or with an external (Figure 9- 6 c) or internal (Figure 9- 6 d) stratification unit.

258

9.3 Stratification and necessary hot water storage tank volume

Fig. 9- 6 Different variants for solar heat input into a solar combi-store (heat store with SH and DHW demand) with Low-Flow system. Variants a and c are both possible with one or more inlet heights each /9.3/.

High-Flow installations with collector areas below 15 m² (49.2 ft²) are mostly combined with internal heat exchangers in the storage tank, because the available volume in the tank is large enough for the required heat exchanger areas. For larger fields with High-Flow, external heat exchangers with fixed inlet positions are used because the area needed for heat transfer with low temperature loss from collector fluid to storage fluid cannot be incorporated in the store. There can be two or more internal heat exchangers (Figure 9- 7, left) for small collector areas or pairs of inlets located at different heights in the storage tank (Figure 9- 7, right) for large collector areas to allow stratification even with High-Flow systems.

Fig. 9- 7 Heat input with High-Flow collector layout and two internal heat exchangers (left) or external stratification with two inlet/outlet pairs (right) /9.1/

Low-Flow systems with a small collector field may use the internal heat exchanger connected to a stratification device according to Figure 9- 6 b. Larger collector fields driven in the Low-Flow regime could be attached to the store with an external (Figure 9- 6 c) or internal (Figure 9- 6 d)

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stratification unit. The external stratification unit is used for large collector fields, where the three way valves are cheaper than the stratification units. Two inlet positions are normally sufficient. The added value of a third can be insignificant.

9.3.2 Heat input from solar collectors into the heat store for solar combi-systems with solar cooling Various systems for the heat transfer from the solar loop to the heat storage for High-Flow and Low-Flow systems exist. Solar combi-systems with solar cooling operate in High-Flow mode only. The heat input solutions according to Figure 9- 7 (left for smaller collector areas, right for large collector fields) may be the appropriate solution.

9.3.3 Necessary volumes in the tank for solar combi-systems without cooling Figure 9- 8 shows the principle of a buffer store for a solar combi-system for domestic hot water (DHW) and space heating (SH) with two energy inputs (solar and auxiliary). In the following some layout aspects from the piping connections to the different heat sources and heat sinks are described in order to show the complexity of such a system. The demand volume for DHW can be either used via an internal spiral heat exchanger (Figure 9- 9 a), by an integrated DHW tank (Figure 9- 9 b) or by a continuous flow heater (heat exchanger) (Figure 9- 9 c). The two methods need a preheater at the bottom of the store.

Fig. 9- 8 Zones for a hot water store of a solar combi-system, charge volumes for solar collectors and auxiliary, demand volumes for DHW and SH /9.5/

DHW requires the highest temperatures and therefore the reserved demand volume is placed at the top of the tank. The SH demand volume is situated under this, as for low temperature heating systems the required temperature is lower. The lowest part should be reserved for the preheating of cold water from the mains of the DHW production.

260

9.3 Stratification and necessary hot water storage tank volume

Fig. 9- 9 Different variants for DHW preparation from a solar combi-store /9.3/

The charge volume for the auxiliary boiler should always be large enough to guarantee the satisfaction of the user demand for DHW and space heating. This means that on the one hand, the lower the auxiliary power, the bigger this volume should be and on the other hand, the bigger the hot water demand the bigger the volume should be. As a worst case for the power demand of the users, the filling of a bath tub with hot water can be used, with a required power of about 26 kW (88,795 BTHU/h) (see calculation example below). This power is normally not available from the burner. Therefore the volume for the auxiliary boiler must be at least big enough to satisfy a hot bath tub and in addition some spare volume for the space heating production. On the other hand the charge volume of the solar collectors should be as large as possible in order to increase the solar energy use. Another aspect is the type of auxiliary burner. If it is a non-regulated burner (e.g. log wood burner) the volume should be large enough to store the energy of the burner filled with fuel. Example: Power demand DHW Volume of bath tub:

Vtub = 0,15  m³ (150 l, 40 gal)  

Filling time:

t = 0,2 h (12 minutes)

Specific heat capacity, water:

cpwater = 1,16 kWh/m³, K – 8.3 BTU/gal, °F

Cold water temperature :

To = 10° C (50° F)

Hot water temperature:

T1 = 40° C (104° F)

Volume-flow: Heat capactiy for DHW:

Vtub = Vtub / t = 0,75 m³/h (198.15 GPH) tub = tub

× cpwater × (T1–To) = 26 kW

According to this the required volumes are defined as follows: Normally the return pipe from the domestic hot water preparation is the coldest return. It should always enter at the bottom of the store. From here to the return line of the space heating system, the preheating zone for domestic hot water can be defined. If the space heating is an ambient temperature controlled system, medium level temperatures are needed for the space heating demand. Therefore the demand volume for

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9  Solar thermal system design

space heating should be above the preheating zone for domestic hot water, but below the demand volume of domestic hot water which normally is the warmest part. The auxiliary heating should cover the DHW hot zone plus some volume for space heating.

9.3.4 Storage volume for solar combi-systems with solar cooling Adding the cooling function to the store means in principle adding an additional demand volume at the top of the tank, because solar cooling needs the highest temperature compared to the DHW and SH functions. If the store size of the solar combi-system is relatively large, it may be necessary to reduce the total heated tank volume in summer to ensure that the cooling operation can start relatively early in the day without having to heat the entire storage tank to high temperatures first. This can be done by including an extra inlet of the solar collector at mid height of the storage tank that is used in summer (see Figure 9- 7). Of course this outlet can also be used in winter to reach temperatures for space heating operation faster. This second inlet can be an additional integrated heat exchanger (Figure 9- 7, left) or an additional outlet to the external heat exchanger with a three way valve (Figure 9- 7, right). A separate store for the solar cooling function can be added to the system. It must be decided whether solar cooling or domestic hot water has priority in summer. For office buildings, where the DHW demand is low and often produced by decentralized small electrically driven water heaters, the demand volumes in the store are reduced to one useful volume. During the heating season this volume is heated to SH temperatures and in summer it is kept at the solar cooling level. The charge volume for the auxiliary boiler should be kept as small as the auxiliary boiler allows (see above).

9.3.5 Stratification Stratification of the storage tank means that the various temperature levels (lower temperature in the lower part and higher temperature in the higher part) are not mixed during the operation of the store. This is very important, if a wide range of temperatures occurs in the tank at the same time (eg. DHW, SH and solar cooling application). DHW has a fresh water inlet temperature of 10 °C (50 °F) and an outlet temperature of 45–60 °C (113–140 °F). 90 °C (194 °F) are needed for the desorber of the cooling machine. Mixing would destroy the useful water temperature layers, which would increase the required auxiliary heat and thus would reduce the solar contribution. If the temperature difference in the store is small (e.g. only space heating with low temperature heating systems and/or solar cooling) with little difference between flow and return temperature

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9.4 Other components of the solar loop for solar cooling systems

of the demand, good stratification is far less important. This is also true, if the DHW demand is only minor compared to the other heat loads.

9.4

Other components of the solar loop for solar cooling systems

All components must stand the higher operating temperatures of the solar cooling system compared to the DHW and space heating system (especially pumps and valves). Temperature losses should be avoided with even greater care relative to DHW and space heating systems, since the operating temperature is close to the evaporation temperature. Therefore the heat exchangers should be sized for 3–5 K (5.4–11.3 °F) temperature loss rather than 5–10 K (9–18 °F) as in space heating and DHW systems. The operating pressure of the collector fluid should be normally higher in solar cooling systems (4 bars (58 PSI) rather than 2 bars (29 PSI)) in order to avoid evaporation during normal operation. Only for very well hydraulically balanced systems can the pressure be lower. The insulation of the pipes should be thicker than for space heating and DHW systems, to avoid higher thermal losses due to the higher operating temperatures.

/9.1/ /9.2/

/9.3/

/9.4/ /9.5/

Streicher, W. (2008): Solar Thermal Heating Systems, lecture book, Institute of Thermal Engineering, Graz University of Technology Streicher, W. (2001), Minimizing the Risk of Water Hammer and Other Problems at the Beginning of Stagnation of Solar Thermal Plants – a Theoretical Approach, Solar Energy, Volume 69(Suppl), Number 1–6, pp 187–196. Haller, M. (2010): Combined solar and pellet heating systems – Improvement of energy efficiency by advanced heat storage techniques, hydraulics, and control, PhD Thesis, Graz University of Technology, Institute of Thermal Engineering. Casals, L. (2010): L’autovidange, une solution contre les surchauffes, issue 88, Vecteur gaz – Juin–Juillet 2010 Bales, C., Drück, H., Hadorn, J., C., Streicher, W. (2005), Advanced Storage Concepts for Solar Houses and Low Energy Buildings – IEA-SHC TASK 32, International Solar Energy Society (ISES) Conference 2005, p. 323–345

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Chapter 10

Pre-engineered systems: built examples and experiences Responsible Author:

Dagmar Jähnig, University Kassel, Germany Contributing Authors: Anita Preisler, AIT, Austria Martin Helm, Bayerisches Zentrum für Angewandte Energieforschung e.V., Germany Daniel Mugnier & Romain Siré, TECSOL, France Daniel Neyer, TU Graz, Austria Alexander Thür, AEE – Institute of Sustainable Technologies, Austria Tomas Nuñez, Fraunhofer Institute for Solar Energy System ISE, Germany

Small-scale solar heating and cooling systems with cooling capacities up to 20 kW (5.7 ton) have become increasingly popular in recent years. One reason for this is that a number of small-scale thermally driven chillers have entered the market. Such systems are installed in small residential applications as well as in small office buildings and other commercial applications. Especially for small capacity plants, it is important that systems are as pre-engineered as possible by the manufacturer, reducing planning efforts and thereby the overall cost. But preengineering for typical applications and loads can make sense for larger plants too. This chapter concerns only systems including thermally driven chillers, not desiccant evaporative cooling (DEC) systems. The reason is that only very few small capacity DEC systems exist. In the past these systems have always been custom-made. Because of their complexity they can hardly be installed by a traditional installer which means it is very likely that they will remain custom-made at least within the near future. A general description of pre-engineered systems has already been given in section 6.1. A number of companies, at least in Europe, offer cooling kits where all the necessary components for a solar heating and cooling system are sold in one package. However, sales numbers of these systems are still relatively small and they are, in many cases, not truly pre-engineered but certain parts, like some important hydraulic components (pumps), are adapted to fit the specific application. This chapter shows some examples of existing small-scale solar heating and cooling systems. These examples are among the best working systems that have been monitored in the framework of IEA-SHC Task 38. The chapter also reports on experiences in order to give recommendations to persons planning to install such a technology as well as to system suppliers designing

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p­ re-engineered systems. Further recommendations on how to avoid problems during planning, installation, operation and maintenance, are given. The experiences reported here have been collected among participants of IEA-SHC Task 38 who have designed (only in some of the cases), operated and monitored small-scale systems. Furthermore the experiences of non-professional users gained through a survey on 18 small-scale solar heating and cooling systems are presented.

10.1 What can be expected from a pre-engineered system? Small-scale solar heating and cooling systems can be bought from specialized companies as complete packages, these include the entire solar thermal plant with hydraulics, heat storage tank, absorption or adsorption chiller, heat rejection system, possibly cold storage tank and a control system of the whole installation. The degree of pre-fabrication varies from supplier to supplier. In some cases, all pumps are pre-defined depending on the flow rate and an assumed pressure drop taking into account distances between components. In other cases these components need to be sized specifically for each installed system. These pre-engineered packages are then connected to heat, cold and domestic hot water distribution systems. As already mentioned the degree of pre-fabrication varies from supplier to supplier. It is recommended to choose a package that is as pre-engineered as possible. In this way, it is ensured that all components work together well. An important aspect is that the controller is included in the package and the control strategies necessary to ensure proper operation of all system components have been pre-defined by the manufacturer. Most manufacturers offer pre-defined control strategies depending on the exact configuration of the system e.g., the temperature level of the chilled water circuit. Not all packages on the market include a control system which manages, besides the solar thermal plant and the chiller, the heat and cold distribution systems. In these cases, a separate controller is needed. However, the overall control concept of the building often has a big influence on the performance of the entire system. As an example, the cold distribution system should take into account whether cooling from the thermally driven chiller is available or not. That shows that the design matching of the controller for cold production and the controller of the heat and cold distribution is an important issue. Therefore packages that include the control algorithms also for the distribution system are recommended. The packages available on the market present different cooling capacities. The number of smallscale thermally driven chillers on the market is not large; therefore often the design process starts from the chiller choice. The rest of the system is then pre-engineered to fit the chiller size. For choosing the correct package size for the given application, an assessment of the cooling load of

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10.2 Built examples

the building has to be carried out. The system suppliers help the customer in the planning process or recommend the service of a planning office to identify the cooling load of the building. The suppliers of such systems typically have affiliated installers that carry out the installation work. Most likely the commissioning is carried out by the supplier company itself to ensure proper installation and operation of the plant. Many companies offer monitoring and optimization processes of the system. Since solar air-conditioning is a relatively new technology, this should systematically be used in order to avoid possible malfunctioning, minimize energy consumption and therefore operating costs of the system. Maintenance is normally carried out by the system suppliers themselves or by an affiliated company typically at least once a year. However, some packages include a wet cooling tower as the heat rejection component which requires particular maintenance: emptying of the water before the winter and refilling in the spring, a water treatment management process (compulsory in some countries). With a pre-engineered system, it should not be necessary for the end-users to conduct any maintenance operations on their own. In the survey mentioned, most end-users have reported that the operation of the system is convenient and simple. The desired room temperature can easily be modified on the control terminal.

10.2 Built examples Example 1: 7.5 kW (2.1 ton) – adsorption chiller in laboratory building in southern France (Perpignan) General Description The targeted building selected for the solar cooling application is the CNRS PROMES research centre office. It is dedicated to research work and the offices in the technical area TECNOSUD of Perpignan located in the Languedoc Roussillon area (South of France). The plant serves only a small part of a large building (administration office). The general orientation of the building is 45° towards west (30° collector tilt angle) and the collector field is oriented in the same direction of the roof. The building was created in 2000 and is of a good quality energy level. A photograph of the solar collector and the adsorption chiller is shown in Figure 10- 1 and a sketch of the hydraulic scheme (without multi-split system that works as back-up) is shown in Figure 10- 2. The system is based on a 7.5 kW (2.1 ton) adsorption chiller coupled with 24 m² (258 ft²) double glazed flat plate collectors. The system produces cold energy independently, in parallel to a general multi split compression system (i.e. solar cooling system covers the base load and the backup adapts the power to fit the load).

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The distribution system for the solar cooling system is an independent chilled/hot water network using 3 fan-coils working at 14/18 °C (57/64 °F) supply/return temperature level, in cooling mode. The heat rejection system consists of a dry cooler assisted by a spring water spraying device, only used in the case of very hot days (over 30 °C, 86 °F, ambient temperature). In operation since 

2008

Air-conditioned area 

180 m² (1938 ft²) (on 2 levels)

System used for space heating?

Yes

System used for DHW preparation?

No

Fig. 10- 1 Double-glazed flat plate collectors and adsorption chiller installed at a laboratory/office building in Perpignan, France Source: TECSOL

Central air-conditioning unit Nominal capacity

7.5 kWcold (2.1 ton)

Type of closed system

Adsorption (Silica gel-water)

Chilled water application

Fan coils

Dehumidification

No

Heat rejection system

Dry cooling tower with optional spring water spraying

Solar thermal Collector type

Double-glazed flat plate collectors

Collector area

25 m² (269 ft²) absorber area

Tilt angle, orientation

30° tilt, 45° west

Collector fluid

Water (drain-back system)

Typical operation temperature

75 °C (167 °F) driving temperature for chiller operation

Configuration Heat storage

0.3 m³ (793 gal) water

Cold storage

0.3 m³ (793 gal) water

Auxiliary heater

el. heat pump (separated system, multisplit)

Auxiliary chiller

el. heat pump (separated system, multisplit)

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10.2 Built examples

solar collector

heat exchanger Drain Back

heat storage

backup

hot side

B

G A

main cooler

heat exchanger

auxiliary cooler

heat exchanger

heat rejection C A

load

distribution

cold storage

backup

cold side

A

E T

B

solar collector

heat exchanger

heat storage

backup

hot side

G

Drain Back

main cooler

heat exchanger

auxiliary cooler

heat exchanger

heat rejection C A

distribution

cold storage

backup

cold side

E

T

M

load

T

Fig. 10- 2 Hydraulic scheme of solar cooling and heating system in Perpignan, cooling mode (top) and heating mode (bottom) (multi-split system, that works as back-up, is not shown)

Monitoring Results – Cooling operation A summary of the monitoring results are presented in Table 10- 1. The system has been working properly for more than 1.5 years in cooling and heating mode. The total electrical COP reached an average of 4.6 in the summer months. In this period almost 39% of primary energy was saved compared to a conventional compression cooling system that supplies the same amount of cooling. The building owner is satisfied with the solar cooling and heating system.

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10  Pre-engineered systems: built examples and experiences

2009:

May

June

July

August

Septemper

Average

Total electrical COP

5.5

4.7

4.6

4.0

4.4

4.6

Primary energy savings

49.3%

40.3%

38.8%

30.2%

35.8%

38.9%

Tab. 10- 1 Monitoring results adsorption chiller in laboratory building

Figure 10- 3 and Figure 10- 4 show a typical summer day operation. The system provides cooling from about 10:30 a.m. to 5 p.m. It has to be noted that the temperature level evolution at the generator side during the operation period closely follows the solar irradiation while the condenser temperature level remains quite stable. The consequence of this behaviour is good stability of the chilling capacity over the working period: this stability is mainly due to the fact that the adsorption chiller used is not very sensitive to temperature variation between 70 and 80 °C (158 and 176 °F) in the generator loop (the heat exchange being nearly the same).

Fig. 10- 3 Temperatures for the 19th July 2009, °F = °C × (9/5) +32



Fig. 10- 4 Power for the 19th July 2009, 1 kW = 3415 BTU/h



270

10.2 Built examples

Example 2: 10 kW (2.8 ton) absorption chiller in office building in Garching, Germany General Description This solar heating and cooling system was erected at the office building of ZAE Bayern in Garching, Germany in 2007. The system is used to cool and heat 400 m² (4306 ft²) of office area and to provide domestic hot water. It consists of 57.4 m² (618 ft²) of flat plate collectors and a compact water/LiBr absorption chiller with 10 kW (2.8 ton) nominal capacity. There is a wood pellet boiler for backup during the heating season and a water well as backup for cooling purposes. A special feature of this system is the use of a latent heat storage system. By integration of a latent heat store into the heat rejection system of the absorption chiller, a part of the reject heat of the chiller can be buffered during the operation of the solar cooling system, allowing for lower heat rejection temperatures during peak load operation of the chiller. The stored reject heat then can be discharged using a dry cooler during off-peak operation or night time when more favorable ambient conditions, i.e. lower ambient temperatures or lower electricity tariff, are available. According to its melting temperature in the range of 28–29 °C (82–84 °F) the salt hydrate calcium chloride hexahydrate (CaCl2•6H2O) was chosen as the phase change material (PCM), providing a heat capacity of about 150 J∙g-1 (64.6 BTU/lb) or 240 kJ∙litre-1 (863.3 BTU/gal) between 22 °C and 36 °C (72–97 °F). During the heating season the latent heat storage balances heat generation by the solar thermal system, other heat sources and the supply to the consumer. Thus, a low operating temperature of the solar thermal system is accomplished yielding efficient operation with optimum solar gains. The chilled water is distributed within the building by means of chilled ceiling panels. Figure 10- 5. shows a schematic diagram of the absorption chiller and a photograph of the latent heat storage tanks. A sketch of the hydraulic scheme is shown in Figure 10- 6.

Fig. 10- 5 Absorption chiller and PCM storage tanks of the solar cooling and heating installation at ZAE Bayern in Garching, Germany.

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10  Pre-engineered systems: built examples and experiences

In operation since 

2007

Air-conditioned area 

400 m² (4306 ft²)

System used for space heating?

Yes

System used for DHW preparation?

Yes

Central air-conditioning unit Nominal capacity

10 kW cold (base load) (2.8 ton)

Type of closed system

Absorption (Water-LiBr)

Chilled water application

Ceiling panel

Dehumidification

No

Heat rejection system

Dry cooler supported by a latent heat storage

Solar thermal Collector type

Flat plate

Collector area

57.4 m² (618 ft²)

Tilt angle, orientation

40°, south +10° west

Collector fluid

Water-glycol

Typical operation temperature

92 °C (198 °F) driving temperature for chiller operation

Configuration Heat storage

2 × 1 m³ (264 gal) water tank (in series) and 1.6 m³ (423 gal) latent heat store

Cold storage

None

Auxiliary heater

Pellet boiler

Auxiliary chiller

Water well

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10.2 Built examples

solar collector

heat exchanger

heat storage 45C°

C

T

backup

hot side

B

G

T

A

E

main cooler

heat exchanger

auxiliary cooler

T

heat exchanger

heat rejection C

C PCM A

T

load

D

E

D

distribution

cold storage

backup

cold side

A

E T

solar collector

T

B

heat exchanger

heat storage

backup

hot side

45C°

T

G

T

T

main cooler

heat exchanger

auxiliary cooler

heat exchanger

heat rejection C

PCM A

T

load

distribution

cold storage

backup

cold side

E T

T

Fig. 10- 6 Hydraulic scheme of the solar cooling and heating installation at ZAE Bayern in Garching, Germany, for cooling mode (top) and heating mode (bottom)

Monitoring Results – Cooling Operation The system has been operated and monitored for three complete years. Table 10.2. shows a summary of the results for the summer of 2009. The total electrical COP of the solar cooling system was 6.6. Almost 60% of primary energy has been saved compared to a conventional compression cooling system that supplies the same amount of cooling.

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10  Pre-engineered systems: built examples and experiences

June 2009

July 2009

August 2009

Average

Total electrical COP*

5.4

7.0

7.2

6.6

Primary energy savings

49.4%

60.0%

61.2%

56.9%

*

only for the solar cooling part, not water well backup

Tab. 10- 2: Monitoring results absorption chiller in office building

Figure 10- 7 shows the capacity of the chiller and the heat rejection loop during a typical summer day. The absorption chiller provides about 11 kW (3.13 ton) chilled water by means of 14.5 kW (49520.1 BTU/h) driving heat from 10 a.m. to 5 p.m. Until 9 p.m. the chiller operation is continued with surplus driving heat from the buffer tank, stored during the day. The chiller’s reject heat is primarily dissipated by the dry cooler. During hot ambient conditions (10 a.m. to 8 p.m.) the latent heat storage tank supports the dry cooler and ensures a maximum heat rejection temperature of 33.2 °C (92 °F). The discharging of the latent heat storage tank is controlled with regard to the actual heat content (in this case from 0 to 6 a.m.).

Absorption chiller

COP thermal COP electrical

Dry cooler and latent heat store

0.76 7.86

Tmax ambient Tmax cooling water

32.8 °C (91 °F) 33.2 °C (92 °F)

Fig. 10- 7 Measured data of the chiller and the reject heat loop during a typical summer day

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10.2 Built examples

Example 3: 17.5 kW (5 ton) absorption chiller in an office building in Graz, Austria, 1 kW=3415 BTU/h General Description The office building was renovated in 2004 and the solar heating and cooling system was installed in 2008. The office façade has a south and west orientation. In order to reduce the solar gains external shading devices are installed at each glazing. Because of internal gains and ventilation through the windows, active cooling is indispensable. The solar heating and cooling equipment is installed in a so-called “Cool Cabin” placed in front of the building. The solar collectors are installed on the roof of the cabin. The hybrid cooling tower is placed on the flat roof of the office building. The cooling of the office rooms is realized using ceiling cooling elements. A photograph of the main façade of the office building and the cooling cabin covered by the collector field as well as a view of the cooling cabin is shown in Figure 10- 8. A sketch of the hydraulic scheme is shown in Figure 10- 9.

Fig. 10- 8 “Cool cabin” with solar thermal collectors on the roof and view inside the cabin Sources: Thomas Weissensteiner (left), SOLID (right))

The cooling system uses a 17.5 kW (5 ton) water – lithium bromide absorption chiller and high temperature flat plate collectors (equipped with an additional Teflon foil to reduce heat losses). In winter the solar collectors operate in assistance to the conventional space heating. In summer and winter the solar generated heat is stored in a buffer store and all energy needed is taken out of the buffer store. The backup heat from the district heating is not stored in the tank but directly transferred to the space heating system. A special application is the direct usage of the hybrid cooling tower for free cooling via the chilled ceilings. There is no back-up system for cooling in summer.

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10  Pre-engineered systems: built examples and experiences

In operation since 

2008

Air-conditioned area 

430 m² (4629 ft²)

System used for space heating?

Yes

System used for DHW preparation?

No

Central air-conditioning unit Nominal capacity

17.5 kW cold (5 ton)

Type of closed system

Absorption (Water-LiBr)

Chilled water application

Chilled ceilings

Dehumidification

No

Heat rejection system

Closed circuit cooling tower, hybrid

Solar thermal Collector type

Double glazed flat plate collectors

Collector area

60 m² (646 ft²) gross area

Tilt angle, orientation

11°, south

Collector fluid

water-glycol

Typical operation temperature

88 °C (190 °F) driving temperature for chiller operation

Configuration Heat storage

2 m³ (528 gal) water

Cold storage

0.2 m³ (52.9 gal) water

Auxiliary heater

District heat for space heating in winter

Auxiliary chiller

none

276

10.2 Built examples

solar collector

heat exchanger T

heat storage

hot side

backup B

T

G A

main cooler

heat exchanger

auxiliary cooler

heat exchanger

heat rejection C T

load

distribution

cold storage

backup

A

cold side

A

E T

B

solar collector

heat exchanger T

heat storage

hot side

backup

T

G

main cooler

heat exchanger

auxiliary cooler

heat exchanger

heat rejection C T

load

distribution

cold storage

backup

A

cold side

E T

Fig. 10- 9 Hydraulic scheme of solar heating and cooling system in an office building in Graz, Austria for cooling mode (top) and heating mode (bottom)

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10  Pre-engineered systems: built examples and experiences



Fig. 10- 10 Measurement curves for a typical sunny day, °F=°C × (9/5) +32, 1 ton = 3.517 kW

The daily performance of a typical sunny day is shown in Figure 10- 10. The chiller starts its operation at 11.30 a.m. after the storage reaches 80 °C (176 °F). The driving temperature increases slightly during operation and the average temperature is about 77 °C (171 °F). The chilled water temperature decreases from 20 °C (68 °F) to 8 °C (46 °F) while the heat rejection temperature is kept at 29 °C (84 °F). Within these conditions a decreasing driving heat and cold production is visible. An average generation power of 20.7 kW (5.9 ton) and an average cooling power of 14 kW (4 ton) leads to an average thermal COP of 0.68; these are average values for this particular day. Example 4: 19 kW (1.2 ton) absorption chiller in an office building in Gleisdorf, Austria General Description The office building of the local utility in Gleisdorf was equipped with a solar heating and cooling system in June 2010. Air ventilation is done via a manual window opening and all windows are equipped with external shading devices. Cold water is generated by an ammonia-water absorption chiller and is distributed via chilled ceilings. Solar thermal energy is produced by a 64 m² (689 ft²) collector field. During the cooling period waste heat from a combined heat and power plant (CHP) driven with vegetable oil serves as the heat back-up. Within the heating season three CHP units, a condensing natural gas boiler and a heat pump serve as the back-up; in addition to the office building also a local district heating system has to be supplied with heat. The generator heat for the absorption chiller is taken out of the high temperature tank, which is a 2 m³ (528 gal) tank part of 10 m³ (2642 gal) total storage. The cold water generated by the chiller is transported directly to the cold distribution system without any cold water storage in order to avoid an extra cold water pump and thereby reducing electricity consumption. A dynamic cooling power control configuration is implemented for this chiller. The mass flow values and relevant temperature values of the particular hydraulic loops of the chiller (generator and heat rejection loop) are varied in a way that a variable control of the produced cooling power is possible. With that control strategy the actual required cooling energy demand of the building can be generated and an unsteady operation performance of the chiller can be avoided.

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10.2 Built examples

Heat rejection of the absorption cooling process is realized with an open, wet cooling tower combined with an electrolytic water preparation device with the advantage of reducing electricity consumption. Figure 10- 11 shows the entrance of the office building and the cooling tower located here and the ammonia-water absorption chiller located in the technical room in the basement. A sketch of the hydraulic scheme is shown in Figure 10- 12.

 In operation since 

June 2010

Air-conditioned area 

1,000 m² (10764 ft²)

System used for space heating?

Yes

System used for DHW preparation?

Yes

Fig. 10- 11 Cooling tower in front of the building (left), absorption chiller (right) Source: AEE INTEC

Central air-conditioning unit Nominal capacity

19 kWcold (5.4 ton)

Type of closed system

Absorption (Ammonia-water)

Chilled water application

Chilled ceilings

Dehumidification

No

Heat rejection system

Wet cooling tower

Solar thermal Collector type

Double glazed high temperature flat plate collector

Collector area

64 m² (689 ft²)

Tilt angle, orientation

40°, 34° east

Collector fluid

Water-glycol

Typical operation temperature

70–90 °C (158–194 °F)

Configuration Heat storage

10 m³ (2642 gal) water

Cold storage

None

Auxiliary heater

3xCHP units (35 kWth + 18 kWth + 16 kWth) (119,531 BTU/h + 61,473 BTU/h + 54,643 BTU/h)

Auxiliary chiller

Split units in individual rooms like the control room for the local electrical grid

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solar collector

heat storage

heat exchanger

backup

hot side

B

T

G

CHP

T

G

T

A

main cooler

heat exchanger

auxiliary cooler

heat rejection

heat exchanger

M T

C T

load

A

distribution

cold storage

cold side

backup

A

E T

solar collector

T

B

heat exchanger

heat storage

backup

T

G

CHP

hot side

T

G

T

main cooler

heat exchanger

auxiliary cooler

heat exchanger

heat rejection

M T

C T

load

A

distribution

cold storage

backup

cold side

E T

T

Fig. 10- 12 Hydraulic scheme of solar heating and cooling system in an office building in Gleisdorf, Austria for cooling mode (top) and heating mode (bottom)

Monitoring Results – Cooling Operation The monitoring data of July and August 2010 showed a good thermal coefficient of performance (average 0.54) as well as a good electrical coefficient of performance (average: 5.5).

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10.3 Experiences An end-user survey for small scale solar cooling plants was carried out to get an overview of the quality of currently installed plants as well as experiences in system configuration, planning, installation, operation and maintenance issues. In total, 18 interviews were conducted in 6 countries (Spain, Austria, Germany, Switzerland, France and Italy) and on 8 application types (office buildings, single family houses, kindergarten, training centre, sport centre, retirement home, winery, mixed use). No statistical conclusions can be taken from this survey because of the small sample size but their results give an indication of the ongoing trends on quality issues. However, it has to be noted that the documented systems may in some way have been pre-selected by the concerned actors (installers, system providers) in order to show mostly good experiences.

Fig. 10- 13 Opinion of plant owners on solar cooling plant quality level Source: end-user survey Task 38

Although the plant quality level was seen as good and very good in most cases (see Figure 10- 13), a number of issues have come up that show that most of the installed small-scale systems are not yet truly pre-engineered and have required significant effort for the planning, installation, operation and maintenance than should be expected from a pre-engineered system. In the sections below, these issues which characterize truly pre-engineered systems are described.

10.3.1 Installation issues The idea of pre-engineered systems is not only that the planning effort should be reduced but also that installation can be done by less specialized installing companies. The system should be pre-engineered and pre-assembled as far as possible so that it should be possible to undertake the installation within a few days at a maximum. There should be no need to employ several different companies for each part of the system (e.g. solar thermal system, chiller, controller, heat and cold distribution). However, the end-user survey showed that even for small-scale solar cooling plants the installation duration differs quite a lot. The long installation periods were mainly caused by the difficult coordination when several companies were involved or because of delays on component delivery. When pre-assembled systems were directly delivered on site, delays were avoided. In some of the cases covered by the end-user survey, a high number of workers were employed for the installation; this should be avoided in the case of the installation of pre-engineered systems in order to keep the coordination effort on site low.

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10.3.2 Commissioning Commissioning is an important part of the installation process. A first static commissioning must be done reviewing all the components. Dynamic commissioning must follow and it important in order to check whether operation of the system in different working conditions is not leading to disfunctional behavior. However, the end-user survey showed that in several cases the commissioning was of very low quality or did not take place at all. This can lead to a poor operating performances and a long time and cost consuming adjustment phase until the plant is fully operational according to the planned values. It cannot be expected that end-users of small-scale solar cooling plants – who may be single family house owners or owners of small commercial enterprises without any experience in solar cooling plants – make major adjustments to the plant. This experience shows that high quality commissioning is important if the operation performance of the whole system at the end of installation is required. In addition, a detailed operating handbook for the end-user should be part of each pre-engineered system.

10.3.3 Maintenance issues Maintenance of a pre-engineered system should be as simple as possible. Therefore, simple system designs and simple control strategies are recommended. The maintenance should serve two main purposes: –– Ensuring the functionality of the solar cooling plant over the lifetime –– Keep the work load for end-users of the plant as low as possible Preferably the company responsible for the maintenance should be the system supplier/provider himself or the installer (because they know the installation and are often closely located to the site of the system). If that is not possible, a company should be chosen which has experience with this kind of system. Table 10.3 shows a list of typical maintenance items of a pre-engineered system: How often?

Maintenance item

Once a year

Chiller: evacuation Whole system: • pressure • temperatures • pumps • probes checking (water, glycol) • elimination of air from the pipes

Tab 10- 3: Typical maintenance items of a pre-engineered system

In addition, it may be necessary to manually switch from cooling to heating operation and vice versa. This leads to a greater simplicity of the system design. In addition, it can be seen as an advantage if the users can decide themselves whether they want the system to operate in cooling or heating mode. However, for some applications it is better to have a fully automatic controller.

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This can be useful in buildings where the building users are not responsible for the operation of the building themselves (e.g. office buildings). Experiences from small-scale solar cooling systems show that it is recommended to do the cleaning of cooling towers automatically rather than manually. This can be done by automatic frequent elutriation (bottom cleaning) of the wet cooling towers and backwashing of the filters (twice a day). In addition one should pay attention not to install heat rejection devices too close to trees, since leaves may fall on the coils and cause malfunctioning of the fans.

10.3.4 Control issues In a pre-engineered system, a central system controller which manages all the components of a solar cooling plant (chiller, solar thermal system, heat rejection, pumps) is necessary. The advantages of a single (centralized) system controller are: –– avoiding communication problems between two or more controllers – especially if they are from different manufacturers –– achieving an optimized operation of the whole solar cooling plant. This can be achieved much more easily with one central controller. Communication connections (RTC, ADSL, etc..) from the local system controller to the system supplier also showed several advantages for the end-user and the system supplier. The system supplier has the possibility to monitor the main variables of the plant to ensure a continuous and efficient operation. The system supplier however can use the data to further improve the system concept. Some system suppliers offer the service of providing their clients with edited monitoring data via internet, in return for getting the data from the plant.

10.4 Recommendations for system suppliers System performance of 10 small-scale solar heating and cooling systems was monitored in detail during the years 2008 to 2010. One parameter that illustrates the energy performance of the system during the summer months is the electrical COP of the overall system (see chapter 7 for more details). This figure compares the produced cold energy to the necessary electricity consumption of all components of the system such as pumps, fans, controllers, boilers and chillers. In order to make use of this figure for comparison with conventional systems, the electricity consumption considered does not include the energy used by pumps for heat and cold distribution or domestic hot water preparation, since these hydraulic cycles are normally also needed for a conventionally designed system. Figure 10- 14 shows this total electrical COP for the summer months for the 10 monitored systems.

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Fig. 10- 14 Total electrical COP of various monitored small-scale solar cooling systems during the summer months of 2008 to 2010.

As can be seen, the electrical COP varies significantly from one system to another. Some systems have good values between 4 and 7. But a few systems show an electrical COP that is so low that the system actually consumes more electricity than a conventional system that uses a vapour compression chiller for cooling. However, a low electrical COP does not necessarily mean that the system concept or the chiller type is not suitable but can have several reasons. One of these reasons is that many systems often operate in part load conditions but the electricity consumption of the surrounding system components is not adapted to the reduced load. This shows the importance of a careful system design and appropriate choice of components. Often the electricity consumption of certain components of the system is simply too high and can be reduced significantly if the system is designed carefully. Since solar cooling is still a relatively new technology, it has to be noted that there is still room for improvement even of the best results shown in Figure 10- 14. To be comparable with a conventional reversible heating and cooling appliance such as an electric heat pump, the electric efficiency of a solar heating and cooling system must be calculated and enlarged to a full year operation. In these conditions, including use in heating mode, the value can easily increase from 5–7 to 10–15. In pre-engineered systems careful design of the system has to be done by the system manufacturer or system supplier. The systems on the market should already be optimized so that the planning effort by the planner or installer of a system is reduced to a minimum and optimal energy performance of the system is ensured. In the following sections experiences from existing systems are presented that show critical issues when designing a system to ensure significant energy savings.

10.4.1 Electricity consumption of auxiliary components Many components in a solar heating and cooling system consume electricity. The electricity consumption of the pumps in the collector circuits and of all pumps in all three loops of the chiller (hot side, heat rejection, chilled water) is not negligible. In addition, in many cases there is a fan in the cooling tower and also the controller and automatic valves consume electricity, even when they are in stand-by operation. Figure 10- 15 shows the electricity consumption of four small-scale solar heating and cooling

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10.4 Recommendations for system suppliers

systems expressed in Wh of electricity per kWh cold produced. While the total electricity consumption of all four systems is in the same order of magnitude, the shares of the different system parts vary significantly. In the first and the fourth system, the predominant part of the electricity is consumed by the heat rejection system (pump and cooling tower fan). In the second system, the heat rejection system is very efficient while the generator loop pump included in the chiller itself consumes most of the electricity. The third system does not include a cooling tower but rejects the waste heat by means of boreholes. This method also leads to relatively low electricity consumption for heat rejection. The second and the third system use an additional pump on the chilled water side. In example 2, this pump is used to charge a cold water storage tank. In example 3, the pump is necessary to deliver cold to a cooling coil in the ventilation system. These system configurations increase the overall electricity consumption.

Fig. 10- 15 Measured electricity consumption of 4 solar cooling systems during the summer months, in Whel per kWh cold produced, HR = Heat rejection

10.4.2 Heat rejection components As the examples show, the heat rejection unit can be a major electricity consumer. This concerns both the fan(s) in the unit itself and the pump in the heat rejection circuit. Figure 10- 16 shows the nominal electricity consumption of a number of heat rejection units that were used in existing small-scale heating and cooling systems. Some of the units are wet cooling towers, others dry coolers or hybrid units where spraying is used to reduce the return temperature of a dry cooler on peak days (hot and dry days). The nominal electricity consumption per kW nominal thermal heat rejection capacity is shown as a function of the nominal thermal heat rejection capacity.

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Fig. 10- 16 Nominal electricity consumption of small-scale heat rejection units (wet and dry cooling towers and hybrid units with spraying) versus nominal heat rejection thermal capacity

It can be seen that the data scatter a lot. Among heat rejection units between 20 and 30 kW (5.6 and 8.4 ton) nominal thermal capacity, the nominal electricity consumption can be as low as 10 W/kW or up to three times as much. In principle, the use of evaporative cooling (wet towers and hybrid ones) can significantly help to reduce the power consumption but nevertheless the figure indicates that a lot of electricity can be saved by installing an energy efficient unit, since the values even scatter for each category of systems. The effect of choosing energy efficient auxiliary components on the total electrical COP of the system is shown in the following example for the summer months. The data come from a system using a 10 kW (2.8 ton) absorption chiller.

Fig. 10- 17 Total electrical COP and fractional primary energy savings of a solar heating and cooling system where three pumps have been exchanged by more efficient ones (2008 to 2009) and calculated values for additionally exchanging the AC dry cooler fan against an EC version consuming significantly less electricity

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10.4 Recommendations for system suppliers

In this example the collector loop pumps and the pump in the generator loop of the chiller have been exchanged in winter 2008/2009. The nominal electricity consumption of the two collector loop pumps was reduced from 298 W to 170 W (1,017.5 BTU/h to 580.45 BTU/h), that of the generator loop pump from 144 W to 44 W (491.68 BTU/h to 150.23 BTU/h). The first column in Figure 10- 17 shows monitored values of the electrical COP for the year 2008 before the exchange of the pumps and in the second column those values after the exchange (summer of 2009). On the right hand side, the impact on the fractional primary energy savings compared to a conventional compression chiller is shown. In addition, it is planned to exchange the AC fan with an asynchronous motor against an EC model (170 W (580.45 BTU/h) nominal electricity consumption compared to 340 W (1,160.9 BTU/h)). When publishing this book, monitoring data were not available yet. Therefore, the effect of this improvement was calculated using the measured data from 2009 and assuming the same running time of the fan. The results are shown in the third columns in Figure 10- 17. The example shows that choosing more efficient auxiliary components has a significant impact on the overall performance of a solar cooling system.

10.4.3 Part load operation Many solar cooling systems operate frequently under part load conditions. This is often the case if the system covers not only a base load with a backup chiller covering the rest of the cooling load but is designed to match the peak cooling loads of the building as well. In this case the flow rates in the external circuits of the chiller and the velocity of the cooling tower fan should be reduced to save electricity.

10.4.4 Pressure drop in the system An important parameter to pay attention to is the pressure drop in the system circuits. High pressure drops lead to high electricity consumption even if highly efficient pumps are used. Therefore, all heat exchangers and other hydraulic equipment such as heat meters or valves should be designed carefully in order to reduce pressure drops. In addition the pressure drops in the internal heat exchangers of the chiller need to be minimized. The following table shows minimum and maximum pressure drops of small-scale thermally driven chillers available on the market in 2009/2010.

Pressure drop (kPa/PSI) Minimum value

Maximum value

Generator

19.6/2.8

88.3/12.8

Evaporator

23.5/3.4

54.9/8.0

Condenser

31.4/4.6

109.8/15.9

Tab. 10- 4: Pressure drop in internal heat exchangers of small-scale thermally driven chillers (values from 2009/2010)

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The table shows that there is a wide range of different pressure drops on the market today leaving a lot of room for optimization for chiller manufacturers.

10.4.5 Nominal flow rates – high temperature differences Chillers are typically operated with high flow rates in all external circuits in order to achieve optimal temperature levels in the generator and condenser that ensure good heat exchange coefficients in the heat exchangers and the highest possible thermal COP values. However, these high flow rates lead to high electricity consumption of the pumps and therefore lower electrical COP values. It is necessary to find an optimum compromise for high thermal COP and high electrical COP. In recent years, sorption chiller manufacturers have deeply worked on this optimization. However, it may still be possible to improve the electrical COP of the overall system by reducing the nominal flow rates recommended by the manufacturer in the external circuits of the chillers.

10.4.6 Use of a cold store The use of cold water storage tank is very convenient if the load does not match the solar cold production profile during the day. However, a significant amount of the stored energy can be lost because of heat losses through the insulation. Moreover, the management of the temperature distribution level with cold stores often leads to efficiency reduction (chilled water produced in the early morning to cool down the storage tank before distributing). Today, many small-scale solar cooling systems include a cold storage tank. In some cases this is a true storage tank, in other cases it is only a small hydraulic switch to balance constant flow rates through the chiller and varying flow rates in the cold distribution system due to changing loads. Disadvantages of a cold storage tank are higher electricity consumption because another pump is necessary and thermal losses from the storage tank. Monitoring results show that systems have been operated successfully without a cold storage tank by operating the chiller in part load depending on the current load in the building. In addition, in many cases the thermal mass of the building can be used up to a certain limit to store the cold.

10.4.7 Influence of heat rejection temperature Another important aspect to pay attention to is the available heat rejection temperature of a system. This greatly influences the thermal COP of a system. Figure 10- 22 shows an example of monitoring data of the same small-scale adsorption chiller installed in two very different locations/ systems. The high thermal COPs are achieved in a Central European location with a borehole as the heat rejection device which ensures low heat rejection temperatures around 20 °C (68 °F) even in the hot summer months. The same chiller installed with a dry cooler in a Southern European

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10.4 Recommendations for system suppliers

location where spraying is used only on very hot days has to cope with heat rejection temperatures around 30 °C (86 °F) which leads to significantly reduced thermal COPs. Both systems work at similar cold water temperatures (12–18 °C / 54–64 °F).

Fig. 10- 18 Influence of the heat rejection temperature on the thermal COP, monitoring data of the same smallscale adsorption chiller in two different locations, °F = °C × (9/5) +32

This example shows the importance of low heat rejection temperatures for the performance of solar cooling systems. Therefore, the heat rejection system used in a specific plant has to be adapted to the local climatic conditions as well as the legal (e.g. rules maintenance requirements) and geologic (optional use of ground for heat rejection) constraints at the location.

289

Chapter 11

Experiences from installed custom made systems Responsible Authors:

Wolfram Sparber, EURAC research, Italy Assunta Napolitano, Altran Italia S.p.A., Italy Tim Selke, AIT – Austrian Institute of Technology GmbH, Austria Marcus Jones, AIT – Austrian Institute of Technology GmbH, Austria Paul Kohlenbach, Solem Consulting, Germany Contributing Authors: Alexander Thür, AEE INTEC, Austria Olivier Marc, University of Reunion Island, France Franck Lucas, University of Reunion Island, France Osama Ayadi, Politecnico di Milano, Italy Ebbe Münster, PlanEnergi, Denmark Edo Wiemken, Fraunhofer Institute for Solar Energy Systems, Germany Jakub Wojciech Wewior, Fraunhofer Institute for Solar Energy Systems, Germany Klaus Huber, Offenburg University of Applied Sciences, Germany Antoine Dalibard, Stuttgart University of Applied Sciences, Germany Clemens Pollenberg, Fraunhofer Institute for Environmental, Safety and Energy Technology UMSICHT, Germany Hilbert Focke, ASIC – Austria Solar Innovation Center, Austria Luis Angel Bujedo, CARTIF, Spain Daniel Mugnier, TECSOL, France Uwe Franzke, Institut für Luft- und Kältetechnik ILK, Germany

Within the following chapter custom made solar thermal heating and cooling systems are discussed. The target of the chapter is to present practical experiences of worldwide installed systems and thereby enhance the theoretical knowledge. The chapter starts with an overview of documented international installations, showing statistics on the implemented technologies and on the fields of application. This introduction is followed by a detailed presentation of three case studies installed under different climatic conditions. The case studies were chosen as examples for solar assisted heating and cooling systems, solar autonomous cooling systems and solar cooling applications in industry. The case studies have been monitored for several seasons and the main measured performance figures are shown. In the following chapter several possibilities of hydraulic schemes, sizing of main components and control strategies are presented. Next to general information, specific recommendations based

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on experiences are given. Where possible installed case studies are presented and the applied choices are discussed. The chapter closes with a section on commissioning.

11.1 Introduction Within the framework of IEA SHC Task 38, technical figures of worldwide installed systems have been collected. The target was to get a better understanding of the status quo and the distribution of the technologies. Further statistics based on 135 custom-made SHC installations, collected in the years 2008–2010 are shown. The statistics include the end use typology, the installed thermally driven chiller technology, solar thermal technology and some figures on how these components are combined. Most of the documented systems are dedicated to the air-conditioning of office buildings as shown in Figure 11- 1. Further applications are air-conditioning for industries, business premises, hospitals and schools. A small portion (6%) of the considered installations are applied at research centres (denoted “Laboratory”) and are used as well for testing purposes.

Fig. 11- 1 Application areas of 135 documented large scale solar heating and cooling installations. (“Other services” include: hospitals, canteens, several kinds of service centres)

As shown in Figure 11- 2, absorption chillers (Abs) have been implemented in most of these cases. Adsorption chillers (Ads) and desiccant evaporative cooling (DEC) systems, based on liquid and solid sorption materials, have been applied only in around 25% percent of the documented systems.

Fig. 11- 2 Percentage of use of single technologies for thermally driven cooling system within 128 large scale systems.

Considering the implemented solar thermal collector technology, flat plate collectors (FP) and evacuated tube collectors (ETC) are mostly applied (see Figure 11- 3). Compound Parabolic Collectors (CPC), Air Collectors (Air) and Parabolic Trough Collectors (PTC) or other types of single-axis

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11.2 Built examples

tracking concentrating collectors have been applied only in a limited number of cases. Considering the geographical distribution no relevant correlations could be identified. Only in the case of PTC the countries of installation are mainly those with high direct solar radiation as for example Australia.

Fig. 11- 3 Percentage of use of different solar thermal collector technologies within   124 large scale systems.

Correlating the different technologies of solar collectors used and the installed thermally driven cooling technology in 112 installations, it resulted that only specific combinations (12) are used. In Figure 11- 4 shows that sorption chillers are usually used with FP and ETC collectors. Air collectors are only coupled with DEC systems. In fact in DEC systems hot air is required for the regeneration of the desiccant wheel. If flat plate or evacuated tube collectors are used in combination with DEC systems a heat exchanger in the return air flow is needed, in order to heat up this air stream for regeneration of the wheel. Considering absorption and adsorption chillers, it can be noticed that evacuated tube collectors is the most chosen technology for absorption chillers while adsorption chillers are mostly combined with flat plate collectors. Performance improvements of flat plate collectors (double glazing or teflon foil to reduce convective losses, highly selective absorber) has led, in recent installations, to an increased application of this technology in combination with absorption chillers.

Fig. 11- 4 Percentage of the used solar collector technology for the different   thermally driven cooling technologies

11.2 Built examples The following three solar heating and cooling systems are presented in detail. In each case the installation is described and the main figures are reported. Furthermore the system’s control is explained and a detailed breakdown of the electrical energy consumption of the system is shown. Finally the performance figures for one season are reported.

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The three systems are located in different climates and serve different applications. The choice of the three examples was based on the availability of detailed monitoring results showing interesting experiences and lessons learned.

11.2.1 Example 1: office building in Gleisdorf – Austria 11.2.1.1 Description of the plant The renovation and enlargement of the existing head office of the town-hall in Gleisdorf was accomplished in 2008. The enlargement included the realization of a new building to serve as service centre (Figure 11- 5). Renovation included the installation of a solar assisted heating and cooling plant within the framework of the European research project “High Combi”1. The main figures describing the plant are shown in Table 11- 1.

Fig. 11- 5 The town hall and the service centre in Gleisdorf with its solar trees and solar collectors on the roof

Town Hall

Service Centre

Floor area

1.321 m²

1.212 m²

Volume

4.599 m³

3.562 m³

Persons

approx. 25

approx. 25

Heating load

94 kW

31 kW

Annual heating demand

109.000 kWh/a

33.000 kWh/a

Cooling load

38 kW

24 kW

Annual cooling demand

29.000 kWha

42.000 kWh/a

Tab. 11- 1 The main features of the Town Hall and Service Centre in Gleisdorf /11.1/ (1m² = 10.764 ft², 1 m³ = 35.315 ft³, 1 kW = 3,412.14 BTU/h, 1 kWcold = 0.2843 ton)

1 European research and development project “High Combi”: HIGH solar fraction heating and cooling systems with COMBInation of innovative components and methods; www.highcombi.eu

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11.2 Built examples

The cooling load of both buildings is matched by an absorption chiller connected to a distribution system with fan coils in the town hall and to chilled ceilings in the service centre. A DEC system supplies dehumidified fresh air to the service centre via a central air handling unit (AHU). Natural ventilation occurs in the Town Hall via window openings. The heat to both thermally driven systems is provided by solar thermal collectors and assisted by a small district heating connection powered by three natural gas boilers which are located in another building nearby. Two different technologies of solar thermal collectors are used: one is located on the roof and the other, added in 2009, is located on special structures called “solar trees”. Heat from solar thermal collectors and district heating is stored in a common water tank. Solar collectors contribute to the heating and DHW supply as well. Space heating is accomplished via radiators and floor heating in the Town Hall and via ceilings in the service centre. Table 11- 2 and Figure 11- 6 show functional diagrams of the system and key figures of the plant.

Thermally driven cooling ­systems

Solar thermal collectors fields

Storage system Back-up system

Technology

Absorption chiller

Nominal capacity

35 kWcold

Heat rejection system

Open wet cooling tower

Technology

Sorption wheel based DEC system

Nominal capacity – Nominal air flow

35 kWcold – 6,250 m³/h

Regeneration power

65 kW

Technology

High performance flat plate collector with integrated teflon foil

Gross area

134 m² gross area

Tilt angle, orientation

22° tilted and 30° West oriented

Typical operation temperature

85–105 °C

Technology

Standard flat plate collector with extra thick insulation

Gross area

168 m² gross area

Tilt angle, orientation

30° tilted and 30° East oriented

Typical operation temperature

85–105 °C

Cold storage

1 m³

Heat storage

4.6 m³

Heat back-up

District heating – 190 kW

Cold back up

none

Table 11- 2 Technical data of the solar heating and cooling system installed at Gleisdorf town hall and service centre (1 m² = 10.764 ft², 1m³ = 35.315 ft³, 1 kW = 3,412.14 BTU/h, kWcold = 0.2843 ton, 1 °C = 33.800 °F)

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11.2 Built examples

Fig. 11- 6 Functional diagram of the solar heating and cooling system of the town hall and service centre in Gleisdorf /11.2/

11.2.1.2 Cooling mode operation The pump of the primary solar loop is put into operation when the temperature in the collector field is 10 °C (50 °F) higher than the temperature in the base part of the heat storage. A special mass flow control for the two different oriented collector fields ensures that the mass flow of the collector with less radiation is reduced in order to reach the same outlet temperature as the collector with high radiation. The loading of the heat storage occurs via stratifying lances. The return to the solar loop lies at the bottom of the storage tank, which in this case is a comparatively small one. The design goal is to enable 100% solar cooling for most of the summer. Nevertheless, the district heating system is used as a back-up and keeps the top 40% of the heat storage heated up to 88 °C (190.40 °F) in summer, if necessary, e.g. during very hot periods.

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Figure 11- 7 shows a typical operation in the cooling mode during a hot day.

Fig. 11- 7 Example of a day of operation. Main inlet and outlet temperatures are shown on the left and power values of the generator and evaporator are shown on the right /11.2/ (1 °C = 33.800 °F, kWcold = 0.2843 ton)

11.2.1.3 Performance results The reported performance results refer to the absorption chiller based system whereas the DEC system is described in chapter 12. The cooling energy supplied by the absorption chiller from May to August 2010 was nearly 13 MWh (44 357 841.23 BTU). The heat provided to the chiller was approximately 25 MWh (85,303,540.82 BTU), resulting in a seasonal thermal coefficient of performance, COPthermal, of 0.52. The solar fraction calculated over the same time was 86%, being 100% in June and July. The radiation measured on the collector plane was 116 MWh (395,808,429.41 BTU), the efficiency of the solar thermal collector production and distribution was evaluated to be 21%. The efficiency of the solar loop was approx. 27%, whereas the efficiency of the distribution to the tank was approx. 80%. Nearly 3 MWh (10,236,424.90 BTU) of electricity were consumed to assist the cooling supply in the same period. Figure 11- 8 shows the distribution of electricity consumption within different components. As can be seen, the dominant electricity consumer is the heat rejection unit and the pump of the heat rejection circuit.

Fig. 11- 8 Breakdown of electricity consumption of the solar cooling application (only based on the energy balance of the absorption chiller, not of   the desiccant cooling system).

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Considering the cooling supply and the electricity consumption, the seasonal COPelectric was approximately 5.7. The primary energy ratio (PER) of the installation over the same period was 1.30.

11.2.1.4 Conclusions This system shows a special solar heating and cooling system, using two technologies in parallel on both the heat and cold supply side. The concept of two different oriented collector fields in combination with a relatively small heat storage (15 l/m² (13 gal/ft²)) minimizes heat losses, enables an early start of operation and extends operating time of direct solar cooling, resulting in a high utilization of solar energy. This solar thermal system creates some complexity in management, but leads to good performance results on the hot side and good performance of the thermally driven absorption cooling. On the other hand it reduces the overall performance results as soon as fossil auxiliary heating (district heating based on natural gas boilers) is activated. Potential for improvement exists and is expected to be exploited by the use of high efficient electric motors for pumps and fans and an improved operation of the building control system.

11.2.2 Example 2: education centre in La Reunion island – France 11.2.2.1 Description of the plant The unique aspect of this installation is that it is an autonomous solar cooling system located in a tropical climate. This means that solar energy is the only heat source driving the absorption chiller and no heat or cold back-up system is available. The aim is to enhance the thermal comfort of four classrooms in the University of Reunion Island (Figures 11- 9 and 11- 10). In this case the indoor thermal comfort is achieved by a control system operated so as to maintain the indoor temperature 6 °C (42.8 °F) below the outdoor temperature.

Fig. 11- 9 Solar air conditioned classroom in the University of Reunion Island

Fig. 11- 10 Solar thermal system located on the roof of the main building of the University of Reunion Island

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The size of the chiller was selected on the basis of the hours of discomfort which resulted from dynamic building simulations. Table 11- 3 reports the main technical data of different components whereas Figure 11- 11shows the functional diagram of the installation.

Thermally driven cooling systems

Solar thermal collectors field

Storage system

Technology

Absorption chiller

Nominal capacity

30 kWcold

Heat rejection system

70 kW open wet cooling tower

Technology

Flat plate collector

Gross area

90 m² gross area

Tilt angle, orientation

0° North

Typical operation temperature

80 °C

Cold storage

1 m³

Heat storage

1.5 m³

Tab. 11- 3 Technical data of the plant installed at the University of Reunion Island located in Ile de La Reunion, France (1 m² = 10.764  ft², 1 m³ = 35.315 ft³, 1 kW = 3,412.14 BTU/h, kWcold = 0.2843 ton, 1 °C = 33.800 °F)

The solar collector field feeds heat into a single-effect absorption chiller (working pair lithium bromide and water). Heat and cold are stored in two different water tanks. Their volume can provide 45 minutes autonomy in hot and cold water production, whereby the cold storage provides decreasing efficiency when used alone, i.e., without operation of the absorption chiller. The cold distribution is realized via fan coils in four classrooms. The hydraulic system to the four classrooms consists of circuits mounted in parallel fed by the main distribution circuit. Each circuit feeding a classroom is equipped with a three-way valve to adapt the cold production to the building loads. Ceiling fans are installed in each classroom to enhance comfort conditions, e.g. when the solar assisted plant is not in operation.

300

11.2 Built examples

Fig. 11- 11 Functional diagram of the solar cooling plant installed at the University of Reunion Island (Ile de La Reunion, France) /11.3/

11.2.2.2 Cooling mode operation The solar loop pump switches on as soon as the value of irradiation, measured by a pyranometer installed close to the collector field, exceeds a value of 250 W/m² (79.25 BTU/h*ft²). The pump stops when the irradiation falls below 200 W/m² (63.4 BTU/h*ft²). The solar loop three-way valve, which is located between the solar collector field and the hot buffer tank, allows the primary

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fluid temperature to increase fast after the solar loop pump has been switched on and permits to realize a stabilised outlet temperature of the solar collector field. The hot water tank is fed when the outlet temperature of the collector field exceeds 70 °C (158 °F) and if the temperature difference between the outlet temperature of the solar collector field and the bottom of the hot water tank is higher than 5 K (–450.67 °F). The valve is closed and the tank feeding is stopped when the difference of temperature falls below 2 K (–456.07 °F); this ensures that the hot water from the tank does not reheat the collector field. The cold distribution pump feeding the fan coils starts when the cold water tank temperature falls below 17 °C (62.6 °F) and when the rooms need cooling. An occupation schedule controls this distribution process (weekly hours, from 10:00 a.m. to 5:00 p.m.). During the non-occupied periods (week-ends and holidays) cold solar production is accumulated in the cold water tank until the temperature reaches 7 °C. Then a “forcing function” has been implemented to dissipate the cold production in order to avoid overheating of the hot tank, which could rapidly lead to overheating of the solar collectors. Figure 11- 12 shows the temperatures and power of a typical day of operation.

Fig. 11- 12 Example of a day of operation. Main inlet and outlet temperatures are shown on the left side and power values of the generator and evaporator are shown on the right side /11.3/ (1 °C = 33.800 °F, kWcold = 0.2843 ton)

11.2.2.3 Performance results The monitoring results reported here refer to the period December 2008 – June 2009, which corresponds to summer time in the Southern hemisphere. The cooling supplied by the absorption chiller was nearly 10.8 MWh (36,851,129.64 BTU). 23.6 MWh (80,526,542.54 BTU) of heat were fed to the chiller, resulting in a seasonal thermal

302

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coefficient of performance, COPthermal, of 0.46. As no back-up is available, the solar fraction is 100%, but comfort conditions could not be reached in all days. However, a significant increase of indoor comfort was achieved during most of the operation time of the system. The radiation measured on the collector plane was 80 MWh (272,971,330.65 BTU). Thus, the efficiency of the solar thermal collector field including collection and distribution losses is approx. 30%. The efficiency of the solar loop was approx. 32% whereas the efficiency of the storage was approx. 92%. Nearly 4 MWh (13,648,566.53 BTU) of electricity were consumed for the entire system. Hence the seasonal electric performance, COPelectric, was 2.7. It has to be considered that 22% of losses were registered in the cold tank due to the limited cooling load. If this could have been avoided, a COPelectric of 3.3 would have resulted. The primary energy ratio (PER) of the installation over the same period was 1.1. Figure 11- 13 shows the breakdown of the single electricity consumers. As can be seen the cooling tower and the pump to the cooling tower are responsible for more than 70% of the overall electricity consumption. This figure is further discussed in the heat rejection sub-chapter, showing that the cooling tower was oversized leading to increased electricity consumption.

Fig. 11- 13 Breakdown of electricity consumption at La Reunion University installation

11.2.2.4 Conclusions The present example in the island of La Reunion presents peculiarities linked to tropical solar systems focused on cooling all year long. The design aimed at reaching low temperature levels to also assure dehumidification of air in the fan-coils installed in the classrooms. The achieved performance is lower than expected and in particular significant electrical consumption of auxiliary components was observed, mostly related to the operation of the cooling tower. This was partially improved by means of an optimal control (refer to subchapter 11.3.1.4). Furthermore, the system was planned to autonomously guarantee comfort condition most of the time, i.e. without the integration of any back-up system. As a consequence, the system was sized based on the peak cooling demand and this resulted in an oversized system for many hours of part load operation.

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11.2.3 Example 3: Industrial application in Grombalia – Tunisia 11.2.3.1 Description of the plant The plant serves as an industrial refrigerator for a beverage factory in Grombalia, Tunisia (Figures 11- 14 and 11- 15). One of the main energy consuming processes in the factory is the cooling of the twenty fermenting tanks in order to dissipate the heat resulting from the exothermic fermentation process. The solar cooling plant is connected to three of these fermenters (having a volume of 300 hectolitres each) in parallel with the existing cooling system which serves as a cooling backup. The solar cooling system was designed to supply cold temperatures down to –8 °C (17.6 °F) to the fermenting tanks jacket heat exchangers. The system was installed in the frame of the European research project “Medisco”2.

Fig. 11- 14 Fresnel solar thermal collectors installed in the system in Tunisia

Fig. 11- 15 Fermenting tanks behind the installed solar thermal collector field

The system has to operate under unique boundary conditions. The cold temperature to be supplied is below zero while the ambient temperature often reaches 40 °C (104 °F) during the hot season. Moreover, the system concept is intended for regions which face water scarcity, it is therefore unadvisable to use water based heat rejection technology and the selection of air cooled chillers is to be prefered. The main figures of the installation are reported in Table 11- 4.

2 European research and development project “Medisco”: MEDiterranean food and agro Industry applications of Solar Cooling technologies, www.medisco.org

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11.2 Built examples

Thermally driven cooling ­systems

Solar thermal collectors fields

Storage system

Technology

Ammonia-water absorption chiller

Nominal capacity

12.8 kWcold

Heat rejection system

Integrated air cooling

Technology

Concentrating linear fresnel collector

Gross area

120 m² gross area / 88 m² aperture area

Tilt angle, orientation

Single axis tracking, North-South

Typical operation temperature

110–180 °C

Cold storage

3 m³

Heat storage

none

Tab. 11- 4 Technical data of the plant installed in Grombalia, Tunisia, within the Medisco research project (1m² = 10.764 ft², 1 m³ = 35.315 ft³, kWcold = 0.2843 ton, 1 °C = 33.800 °F)

The system (Figure 11- 16) consists of a concentrating linear Fresnel collector; the heat produced by the collector is supplied to the generator of a single-effect water-ammonia absorption chiller which cools down a water-glycol cold storage. This chiller was selected to supply cold water at a temperature below zero degrees. The water-glycol cold storage is connected with the cooling load represented by the three 300 hl fermenting tanks shown in Figure 11- 15.

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Fig. 11- 16 Functional diagram of the industrial solar cooling installation in Grombalia/Tunisia

306

11.2 Built examples

11.2.3.2 Cooling mode operation Figure 11- 17 presents the operation of the system on a typical sunny day of summer 2010. As soon as the radiation level measured by the pyranometer reaches the set point (340 W/m² (107.78 BTU/h*ft²)), the solar pump starts and the temperature of the fluid in the solar loop increases. When the set temperature of 120 °C (248 °F) is reached the operation of the absorption chiller begins. The hot water temperature continues to increase up to 180 °C (356 °F) which represents the maximum set temperature of the collector. At this point the controller of the collector partially defocuses some of the collector mirrors to keep the collector working around 180 °C (356 °F) and not exceed this temperature level. The water in the cold loop was cooled down from about 8 °C (46.4 °F) at the beginning of the operation of the system to about –8 °C (17.6 °F) after the chiller had been working for 7 hours on that day.

Fig. 11- 17 Powers and input/output temperatures of the collector and chiller at one day of operation. (1 °C = 33.800 °F, kWcold = 0.2843 ton)

11.2.3.3 Performance results The monitoring results reported here refer to the period April to August 2010. The cooling supplied by the absorption chiller was approx. 7.5 MWh (25,591,062.25 BTU). 12.5 MWh (42,651,770.41BTU) of heat were fed to the chiller, resulting in a seasonal thermal coefficient of performance, COPthermal, of 0.60. This heat amount was completely delivered by the solar collectors, as no heat back-up system is available.

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The direct beam radiation measured on the horizontal plane of the collectors aperture area was 69.7 MWh (237,826,271.83 BTU). The heat produced by the collector was approx. 18.1 MWh (61,759,763.56 BTU) and thus the solar collector efficiency was approx. 26%. More than 30% of the heat gained by the solar collectors was lost through the loop to the absorption chiller. In this case, it has to be taken into account, that the thermal system is working at far higher temperatures than in the case of usual flat plate or vacuum tube collector fields and therefore the thermal losses in the piping to and from the collector field are higher. Moreover, the optical end losses from this short Fresnel collector used for this pilot plant are of importance and an analysis showed that about 5% better optical efficiency values could be achieved with longer Fresnel collectors which would be used in larger systems. The installation in Tunisia is a demonstration plant which is smaller than a typical installation using tracking solar collector technology. 2.0 MWh (6,824,283.27 BTU) of electricity were consumed within this season, so the seasonal electric performance, COPelectric, was 3.7, which is a good value considering the low cooling temperatures of –8 °C (17.6 °F) and the dry heat rejection at high ambient air temperatures. The primary energy ratio of the installation over the same period was 1.5. Further optimization studies showed, that improving the system hydraulics and using variable speed pumps could lead to an improvement of 35% on the primary energy ratio of the system. In Figure 11- 12 the breakdown of the electricity consumers is shown. As can be seen the absorption chiller including the dry air cooling unit is responsible for only 22% of the overall electricity consumption.

Fig. 11- 18 Breakdown of electricity consumption at   the Grombalia installation

11.2.3.4 Conclusions The present example in Grombalia (Tunisia) shows a particular application for solar cooling by providing cooling at low temperature (below 0 °C (32 °F)), to provide cold for an industrial process (wine production). Specific technologies are used on both the heat and cold side: Fresnel collectors for obtaining heat from solar energy at high a temperature level (up to 180 °C (356 °F)), an ammonia/ water absorption chiller for obtaining cooling at a low temperature (down to –8 °C (17.6 °F)) and dry heat rejection without use of fresh water, thus leading to a reduced maintenance effort.

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11.3 Experiences

The Tunisian plant has performed well since it was put into operation (March 2008). Both the solar collector and the chiller were able to work at their maximum design temperatures. Some optimization measures have been taken and improvements especially regarding the electrical energy consumtpion have been achieved.

11.3 Experiences The aim of the following section is to summarize experiences made in operation of custom-made solar cooling installations. The conclusions and presented application areas are not exhaustive but most common solutions are presented and discussed.

11.3.1 Components integration and layouts 11.3.1.1 Existing layouts for the solar circuits The heat collected in the solar thermal collector field is delivered to the final users through the solar circuit. In this section several possible solar circuit layouts are investigated and explained based on the review of twenty installations. Solar loop In many installations, the heat storage system is the point of connection between the solar loop and the different sub systems, i.e., the thermally driven chiller, the heating distribution system, the DHW circuit and so on. Mainly two different approaches can be used for the connection between the solar loop and the hot storage (see Figure 11- 19): 1. 2.

The solar collector field is connected to an external heat exchanger and a second loop is connected to the heat storage. The solar collector field is connected directly to the heat storage.

(a)

(b)

Figure 11- 19 Diagram of the two main types of solar loop layouts: with (a) and without (b) external heat exchanger

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In climates where freezing might occur the solution with external heat exchanger (left in Figure 11- 19) is the common choice as it allows the utilization of two different fluids, typically an antifreeze water solution (glycol) and pure water respectively. The disadvantage of the external heat exchanger is the need of two pumps in the two circuits, which leads to an increased electricity consumption and the increased temperature losses due to the heat exchanger (typically about 6 K (–448.87 °F)). There are mainly two further possible solutions in climates with freezing risks where only one circuit is used: –– In combination with evacuated tube collectors a pure water based system can be installed if coupled with a control strategy assuring that no freezing occurs in the collectors or the connecting tubes. This control strategy implies that warm water is circulated through the whole primary circuit from the hot water tank in cases where the outdoor temperature falls below the freezing point. An example for such an application is the large scale installation at the FESTO Company in Germany where 1218 m² (13,105.7 ft²) vacuum tube collectors are connected to 17 m³ (600.35 ft³) hot water storages and waste heat from the production process is further applied in the system /11.4/. The disadvantage of such an installation is the additional electricity consumption of the pumps for the circulation of warm water for the freezing prevention and the losses of thermal energy through the circulation. The advantage lies in the simplicity using pure water, its heat capacity, its viscosity and the stagnation safety. –– A further possibility is the application of drain-back systems. These systems are characterized by the fact that the primary circuit is not under pressure (but closed) and water drains back in a non-pressurized water storage if the system is not used. Thereby the collector field remains dry and no freezing risk persists in winter time. Also during summer time in the case of the presence of excess solar radiation the collector pump can be switched off and the collectors are drained. A further advantage of such systems is the fact that no expansion vessels have to be foreseen. An example of such an installation is a solar heating and cooling system installed in Perpignan, France, at the CNRS PROMES research centre (cooling capacity 7 kW (1.96 ton) and further described in chapter 10). A critical aspect of this solution is the fact that it can only be applied in installations where a hydraulic scheme is realized, which allows a complete drain-back of the water from the solar collector field. If water remains in single pipe positions or collectors and freezing occurs the system may be damaged. Today the drain back technology is mainly used in small systems but could be adapted to larger solar air conditioning installations; it has however, already been implemented in several solar domestic hot water systems, especially in Holland, Belgium and France. In climates where no freezing occurs an interconnection without heat exchanger can be implemented (Figure 11- 19, b). An example for such a case is represented by the plant installed at the University of Reunion Island (Figure 11- 11).

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11.3 Experiences

Solar loop – thermally driven subsystems connection Considering the connection between the solar loop and the single load sub-systems several hydraulic schemes are possible. The most common solution is a connection via the hot storage system (see Figure 11- 19). An example for such an application is the installation at La Reunion (see Figure 11- 11). The main advantage of this choice is the straight-forward hydraulic scheme and the simple control of the system. A disadvantage might be present in systems with large hot storage buffers and the need to switch on the thermally driven chiller early during the day. In fact if the primary circuit is only connected to the hot storage buffer, it might take several hours before the temperature needed by the thermally driven chiller is reached in the buffer. Therefore the chiller can be switched on only in the middle to late morning, even on nice sunny days. In order to overcome this drawback the size of the storage might be reduced (with possible disadvantages regarding the solar fraction), or means can be implemented for using only part of the storage tank to achieve high temperatures in a short time (see chapter 9) or a possibility of skipping the storage can be implemented. The latter can be done using a by-pass or by the implementation of a hydraulic junction as a key component for the connection between the solar loop, subsystems and heat back-up systems (if available). Such an application is shown in Figure 11- 20.

Fig. 11- 20 Diagram for the connection of the solar system with the heat driven cooling system using an hydraulic junction or heat collector as core component and bypassing the   storage. HJ = hydraulic junction.

The drawback of this solution is the unsteadiness in the use of the solar collector and the heat delivery to the thermally driven chiller which in particular may occur at unsteady solar radiation conditions (e.g. moving cloud). In this context it has to be taken into consideration that thermally driven chillers are machines with a high inertia and it can take up to 30 minutes till they are switched on again. An example for the application of an hydraulic junction is given by the solar assisted cooling system installed in the applied research centre EURAC research in Bolzano, Italy /11.5/ (see Figure 11- 21 and Table 11- 5). In this case the heat storage and the hydraulic junction are connected to a heat distributor to which the single sub-systems are also connected. Therefore the secondary solar loop flow can bypass the storage be connected to the hydraulic junction and thus it can be directly coupled to

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the heating, cooling and DHW loads (as shown in Figure 11- 20). In this case, the heat backup system is realized by natural gas fired boilers and a natural gas driven cogeneration unit. A risk of a high switching frequency of the absorption chiller is avoided in this way /11.6/, since the primary loop is fed by four large single collector fields and the cogeneration unit also provides a constant heat flow.

Fig. 11- 21 The building of the applied research centre EURAC located in Bolzano (Italy). View of the vacuum tube collectors located on the flat roof

Thermally driven cooling systems

Solar thermal collectors fields

Storage system Backup system

Technology

Absorption chiller

Nominal capacity

300 kWcold

Heat rejection system

Open wet cooling tower

Technology

Evacuated tube collectors

Gross area

353 m² gross area/ 253 m² aperture areas

Tilt angle, orientation

22° tilted – South oriented

Typical operation temperature

85–105 °C

Technology

Heat pipe collectors

Gross area

173 m² gross area/ 105 m² aperture area

Tilt angle, orientation

5° tilted – part west oriented and part south oriented

Cold storage

5 m³

Heat storage

2*5 m³

Heat back up

One gas engine cogeneration unit (180 kWel/330 kWth) Two gas boilers (350 kW each)

Cold back up

Two compression chillers (315 kWc each)

Tab. 11- 5 Main figures of the solar heating and cooling installation at the EURAC research centre (1 °C = 33.8 °F, 1m² = 10.76 ft², 1 kW = 0.28 ton)

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11.3 Experiences

11.3.1.2 Integration of storage systems Heat storage integration is a common practice in solar thermal systems to address the mismatch between solar heat gains and the heat demand of the various loads. Solar thermal systems serving heating and cooling purposes do not require any long-term storage as a significant heat demand occurs in summer time. Actually, in solar heating and cooling systems, an additional degree of freedom in storage design exists, as both heat supplied by solar collectors and cold supplied by the thermally driven chiller can be stored to manage the mismatch between the solar availability and the cooling demand. In the majority of documented installations, water tanks are applied as thermal storage technology and mostly serve to store solar heat rather than cold (see Figure 11- 22). Such a choice helps to manage the mismatch between the solar availability and the heat demand all year round, i.e. not only when the thermally driven cooling system is in operation. Because of this reason and in order to assure stable operation of the thermally driven chiller, the authors generally recommend the installation of a heat buffer. The decision on the installation of an additional cold storage has to be made based on the particular boundary conditions.

 

Fig. 11- 22 Results from statistics carried out on 70 documented installations

In the cases where only one tank is present, it is usually charged with solar energy in the upper part whereas cold flow is drawn out from the bottom of the tank to be heated up again by the solar collectors (see also chapter 9). In several cases hot water storages with charging lances are used in order to optimize tank stratification. The discharging of the tanks is carried out by leaking hot water out of the top of the tank and entering cold water coming from the users into the bottom of the tank. When two or more tanks are used, parallel, series and hybrid connections are applied. On the base of practical experiences, the authors recommend, in general, to try to keep the hydraulic scheme as simple as possible and to reduce the number of single heat storages in order to reduce thermal losses, electrical energy consumption by additional pumps and the probability of mistakes or non-optimal functioning of the three way valves. Nevertheless because of space and volume availability in technical rooms, a combination of two or more heat storages often is necessary. An example of a solar heating and cooling system applying several storages is the plant serving the applied research centre CARTIF, located in Boecillo – Spain /11.7/. The main characteristics of the system are shown in Table 11- 6, Figures 11- 23 and 11- 24. Within the system, two solar thermal

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collector fields (using two collector technologies, namely flat plate collectors and evacuated tube collectors) and four hot water storages have been implemented. For each solar thermal field, two tanks with a volume of 2 000 litres each are connected in series to increase the temperature difference between the inlet and outlet of the solar collectors. Thanks to a three way valve the four tanks can also be connected in series so that the evacuated tube collectors can further heat up the fluid heated by the flat plate collectors instead of only mixing them before entering the absorption chillers.



Fig. 11- 23 View of the flat plate collector field on the roof of the applied research centre CARTIF, located in Boecillo – Spain.

Thermally driven cooling system Technology

Solar thermal collectors fields

Storage system Back up system

Absorption chiller

Nominal capacity

35 kWcold

Heat rejection system

Open wet cooling tower

Technology

Evacuated tube collectors

Gross area

40 m² gross area

Tilt angle, orientation

40º, azimuthally tracking

Typical operation temperature

45 °C in winter time and 90 °C in summer time

Technology

Flat plate collectors

Gross area

37.5 m² gross area

Tilt angle, orientation

40º South

Typical operation temperature

45 °C in winter time and 90 °C in summer time

Cold storage

1 m³ water tank

Heat storage

4*2 m³ water tanks

Heat back up

235 kW gas boiler

Cold back up

16 kW compression chiller

Tab. 11- 6 Main figures of the solar heating and cooling installation at CARTIF research center in Boecillo – Spain (1 °C = 33.800 °F, 1 m² = 10.76 ft², 1 kW = 0.28 ton)

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Fig. 11- 24 Scheme of the integration of storage tanks in the installation at CARTIF

In some cases, storage of the cooling overproduction too can result in a higher exploitation of the available solar radiation. The decision whether or not a cold storage should be implemented strongly depends on the time pattern of the cooling demand and on the applied thermally driven chiller technology. Furthermore the installed distribution system plays an important role. If conventional fan coils are used, the chilled water might be distributed at 7 °C because of the standard sizing of the fain-coil’s heat exchangers. This temperature permits active dehumidification. As this is close to the minimum temperatures of cold water produced by e.g. Li-Br absorption chillers, the energy storing capacity is very limited. However, if radiant ceilings are used or fan coils are installed without the need for dehumidification of room air a cold water temperature of about 17 °C will be sufficient and a cold storage can be implemented effectively. In the following, the temporal distribution aspects are discussed on the base of the SHC installation of CARTIF – Spain. In this installation overall 8m³ of hot water storage volume and 1m³ of cold water storage have been installed. The load curve for this system is primarily given by the air-conditioning need of the building and the presence of the building occupants. During summer time the core working hours are scheduled between 7:00 am and 3:00 pm. This means that the main cooling load occurs several hours earlier than heat from the solar thermal collector field is available necessitating storage to compensate the shift. Monitoring data of operation in the year 2007 show a very limited average yearly efficiency of the storage system. This has several reasons, but a main reason is that the available solar heat being delivered from the collector field to the hot storage from 3:00 pm onwards was not used by the thermally driven chiller as the cooling load was very limited in the afternoon hours. On the other hand, only a limited amount of the thermal energy fed into the hot storage, needed at temperatures above 85 °C, was still available for the thermally driven chiller in the early morning hours of the following day. This led to a significant utilization of the installed back-up boiler (more than 50% of the overall delivered heat).

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As shown by this specific example, the dimensioning of the thermal storages is strongly dependent on the time distribution of the cooling demand and has therefore to be individually designed. A further aspect which can suggest the addition of a cold water tank is the technology used for the thermally driven chiller. According to /11.8/, sorption chillers can be classified on the basis of their operation which can be continuous, semi continuous and batch (see also chapter 4). Adsorption chillers usually feature a discontinuous operation mode. Such principle causes a discontinuous cold generation and can create a mismatch between the cooling supply and need. The installation of a cold water tank can help to manage such a mismatch.

11.3.1.3 Integration of back-up systems On the basis of the fraction of solar use with respect to the overall thermal energy need, solar heating and cooling systems can be subdivided in two categories: 1. In solar autonomous systems the cooling load is covered exclusively by solar energy. In these systems only the solar collectors supply heat to the thermally driven chiller, and eventually to the heating and sanitary hot water systems; no back-up system is available to provide heat to the thermally driven chiller and no back-up chiller is installed. 2. In solar assisted systems only a part of the cooling load is covered by solar energy. In these systems the solar collectors can represent an additional heat source assisting a conventional heat source for the delivery of the needed thermal energy to operate the thermally driven chiller; usually the same heat source is also used to provide heat to the DHW and the heating system. For cooling applications a back-up on the cold side is also possible, using conventional vapour compression chillers. Therefore, when discussing the integration of back-up systems, the following main aspects have to be taken into consideration: –– The position of back-up systems, i.e. back-up on hot side, on cold side or on both sides –– The typology of back-up technology and the type of driving energy (e.g. fossil fuel, bio-fuel, electricity) –– The hydraulic integration of the back-up in the system As is shown in Figure 11- 25, usually solar assisted cooling systems are installed with a backup system on the hot side and possibly with a backup system as well on the cold side. Only in 20% of the cases of the available documented systems, a backup system was installed only on the cold side.

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Fig. 11- 25 Results from statistics carried out on 64 documented installations

Heating dominated climates usually require the integration of heat back-up systems to face solar heat lacks in winter time. As heat back-up systems can be used to feed the thermally driven chiller in summer as well, the integration of a cold back-up system can apparently be avoided (reducing the investment cost of a further component). Such a choice can be done, but it is only recommended under circumstances where the thermal back-up is needed during the cooling season for a very limited number of hours (see section 3 of chapter 7). Among the documented installations the most common types of heat back-up sources are based on: –– boilers mostly driven by natural gas but also by oil and biomass –– heat pumps, which feature the advantage that they can also be used as cold back-up systems in summer time –– cogeneration units which provide heat in parallel to the electricity production –– district heating systems Cold back-up vapour compression chillers are the most common choice. From the layout point of view, heat and cold back-up systems can be connected to the solar collectors and the sorption chiller in series or in parallel, respectively. Integrating the heat back-up systems in series generally gives the opportunity of exploiting even low temperatures from the solar collectors and thus increasing their efficiency. Nevertheless, following the experiences made in various installations, it is important to clarify the flexibility of the heat back-up system with regard to the incoming temperatures and eventually adapt the hydraulic circuit by installing a three way valve for a specific temperature control. Furthermore, the layout with back-up in series often leads to complex schemes which are costly and difficult to run properly. For instance there exists an high risk of wasting back up heat in the collector field if the control system is inadequate. Back-up systems can either be connected to the same storage as the solar collectors or the heat back-up system can be connected in parallel and supply the load directly, as in the case of the installations at FESTO and EURAC where the hot storage system only serves the solar source. In the latter case the stored solar energy is used only when the temperature in the top of the tank is suitable for the desired application (heating or cooling). While the first case features a simpler

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layout, the second case enables a more flexible heat management. This, of course, requires that the back-up heat source react fast enough to adjust the heat production to the actual needs on the load side. About the integration of multiple back-up systems especially in complex systems, a general accepted conclusion is that using compression chillers as the cold back-up source instead of using gas boilers as heat back-up source can result in an energetically more efficient and financially convenient operation. The use of a heat back-up source can be advantageous whenever it derives from heat recovery, e.g. from waste heat as for example in the case of the system installed at the FESTO company. Also the coupling with a Combined Heat and Power systems (CHP) might seem attractive, but attention has to be paid as such systems require specific load curves on the heating, cooling and electrical energy demand in order to be used efficiently. In /11.6/ experiences are described and monitoring data are shown for the system installed at EURAC research centre combining solar, CHP, a sorption chiller, gas boilers and vapour compression chillers.

11.3.1.4 Integration of heat rejection systems The heat rejection system represents a very important component within a solar heating and cooling system. This is due to several factors: –– In general in systems applying thermally driven chillers much more heat has to be rejected in comparison to cases where compression chillers are implemented due to the lower coefficient of performance values (COP values) –– Heat rejection units in many cases represent the single component with the highest electrical energy consumption in a solar heating and cooling system (see examples in the first part of this chapter) –– Furthermore the heat rejection unit can play a relevant role with regard to the maintenance costs of solar heating and cooling systems Therefore special attention should be paid to the choice, the dimensioning and the control of the heat rejection unit. Wherever available, the use of heat sinks such as swimming pools, ground water, lakes, rivers or rain water basins should be considered. As in many cases these heat sinks are not available the most common technologies to be applied are dry, hybrid or wet cooling towers. Even using these technologies large differences regarding the electric energy consumption in relation to the produced cooling effect could be found in monitored systems. In order to reduce the electrical energy consumption, it is highly recommended to choose the proper type and size of cooling tower especially considering the fact that the heat rejection unit will work under part load conditions most of the time. Furthermore, it is recommended to implement pumps which allow a variable mass flow and apply a control strategy which permits to control the speed of the fans within the cooling tower.

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In the following, monitoring results of selected installations are reported. In Figure 11- 26, the energy consumption of the single components of the solar cooling system installed at EURAC research in Italy is shown. As can be seen from the Figure the consumption of the fan of the wet cooling tower and the pump to the wet cooling tower account for 78% of the overall energy consumption; similar values were found in other systems like the ones shown earlier in this chapter.

Fig. 11- 26 Breakdown of electricity consumption of the solar cooling system using an absorption chiller installed at EURAC research   in Bolzano/Italy.

In this system the back-up vapour compression chiller is connected to the same heat rejection unit. In Figure 11- 26 the distribution is shown when only the absorption chiller is covering the cooling demand. If the case in which the cooling demand is covered only by the vapour compression chiller is considered, it can be seen that the main consumer is the chiller itself followed by the wet cooling tower fan and pump to the wet cooling tower which account for 38% of the overall energy consumption. Overall in this installation, approx. 50% more electric energy is needed by the conventional chillers to produce one kWh of useful cold compared to the operation of the solar thermally driven cooling. At the University of Reunion Island, the monitoring data revealed that most of the electricity consumption was due to the fan in the cooling tower and the pump in the heat rejection loop. During the commissioning phase it was decided to put the fans into operation as soon as the chiller was switched on and the heat rejection pump started. However, under particular weather conditions heat rejection was possible without fan operation. Therefore, in 2009 was decided to change the control algorithm such that the fan was only put into operation when the temperature from the condenser exceeded 38 °C and to switch it off as soon as this temperature fell below 33 °C. Thereby the energy consumption of the cooling tower could be reduced to 31% of the overall electricity consumption, below the level of the pump to the cooling tower which was responsible for 41% of the overall electricity consumption. In Figure 11- 12 the energy distribution is shown after the implementation of the optimized control strategy.

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As a third example the SHC system installed in Skive – Denmark is used. This example is of interest as a special control strategy was applied in order to further reduce the electric energy consumption on the SHC system. Within this system the heat rejection is done by one dry cooler and one wet cooling tower (see Figure 11- 27), whereby the dry cooler is also used for the heat rejection from the compression chiller. Pumps of the cold water and cooling water loops are frequency controlled and the pump speed is gradually increased as the cold demand grows. To keep the chilled water temperature constant, the cooling water temperature is regulated instead of the hot water temperature feeding the desorber, as suggested in /11.9/. The regulation is carried out in such a way that in correspondence with an increase of the driving temperature also the cooling water temperature is increased (and vice versa) without any change in the cold water temperature. The cooling water temperature returning to the chiller can be regulated by varying the volume flow in the cooling tower and the frequency of the fan. As long as the cooling load increases, the frequency of the two pumps in the wet cooling tower loop and the frequency of the fan increase as well up to their maximum values. Then, if the cooling demand still increases, the dry cooler is used to reject the heat. The electricity consumption is reduced thanks to partial operation of pumps and fans. It was reported that these measures have improved the electrical efficiency of the overall SHC system by about 10%.

Fig. 11- 27 Representation of the heat rejection system of the SHC plant in Skive – Denmark

11.3.2 Component sizing Within the present subchapter, some experiences are reported with regard to the sizing of the main components of SHC systems. It is important to underline that the component sizing is an individual choice depending on a series of parameters for each single SHC system (e.g. heat and

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cold load profiles to be matched, architectural and space constraints, economics...), therefore it is difficult to give general key figures and dynamic simulation is highly recommended for each individual application. Nevertheless in order to support the designer in the planning phase of an SHC system three aspects to be considered are outlined in the following: –– Figures from worldwide reported installations with regard to solar collector surface and hot storage volumes in relation to the installed cooling capacity and hot storage volumes in relation to the solar thermal collector surface –– Thoughts and reasons which might lead to enhance or reduce the size of single components –– Monitoring results of the thermal COP of thermally driven chillers and efficiencies of installed collector areas –– Size of the solar thermal collector field For the sizing of the solar thermal collector field two main system types have to be distinguished, namely solar autonomous cooling systems and solar assisted (heating and) cooling systems. Under autonomous solar cooling systems we understand systems with no back-up for cold production if the cold available from solar energy is not sufficient. Those systems can be adopted whenever an increase in comfort is needed but not a predefined target indoor temperature (and humidity) has to be assured all year round. Usually dynamic simulations are carried out in order to forecast the building needs and the solar availability at the same time and to optimize the interaction between the different components. The collector field and chiller will be sized in a way that the target temperature (and humidity) will not be met for a certain number of hours per year, i.e. the resulting discomfort is acceptable to the building owner. The smaller this number of hours the larger the collector field and chiller will be. As a result the system will be oversized for many hours in which the system will be working under part load conditions. This may lead to a reduced overall efficiency of the system. Under solar assisted (heating and) cooling systems we understand systems which have a back-up that enables production of cold even when the cold available from the solar energy is not sufficient. If the back-up is installed on the hot side, i.e. works as back-up for the solar collector, the thermally driven chiller has to be sized to match cooling loads at design conditions. If the back-up is installed on the cold side, the vapour compression chiller has to be sized such that even with minimal solar cold production the cooling load at design conditions is matched. In solar assisted systems the relationship between the sizes of both, the solar thermal sub-system, the heat driven chiller sub-system and the building load is less evident as back-up systems in combination with storages can balance the cooling load. Therefore a greater degree of freedom in the design process is given. As a result, the selection of a solar fraction (SF) can be a consequence of strict constraints, such as the available areas for the installation of solar collectors or the available budget.

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If the decision is taken to cover only a small fraction of the heating and cooling demand with the solar driven system, for both back-up solutions (hot and cold side) energy efficient components should be used as they will provide a large part of the requested energy. On the other hand as the existing load, in comparison to the SHC system, is large, this system can run under optimal conditions many hours per year leading to a high energy and economic efficiency of the SHC system. Figure 11- 28 shows the probability distribution for the specific collector area per nominal cooling capacity of the thermally driven chiller for the two main collector technologies – flat plate collectors (FP) and evacuated tube collectors (ETC). The data have been collected on the base of 88 reported SHC systems installed worldwide. As the climatic conditions, the building type and the usage of the building are very different, these values vary considerably. Nevertheless they might give an orientation about the collector size especially in the beginning of the planning phase of an SHC system. According to Figure 11- 28, the most probable ratio between installed vacuum tube collectors and thermally driven chiller capacity is given as 2.6 m² ETC/kWTDC and in the case of flat plate collectors this ratio is given as 3.4 m² FP/kWTDC.

Fig. 11- 28 Probability density for the ratios collector surface (m²) / cooling capacity (kW) applied in documented installations for different solar thermal collector technologies.

 In Figure 11- 29 the mean thermal COP of 6 selected installations is shown; average values were achieved over the time of one or of several cooling seasons. As can be seen this figure varies between 0.35 and 0.58 and especially in one system the average thermal COP falls significantly below the COP of the installed absorption chiller at nominal operation conditions. In order to achieve a high average COP a proper sizing of the single components with regard to the overall demand and the temporary distribution of the demand plays a fundamental role.

Fig. 11- 29 Mean thermal COP in selected installations monitored over one or more cooling seasons



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A key design figure for the solar thermal collector is the average efficiency of the solar thermal collector field. As shown in the monitoring results of 7 installed systems (Figure 11- 30) this figure varies from 12% to 32%; these values correspond to the annual average value. In comparison with solar thermal DHW systems, the efficiency of the solar thermal collector fields in solar cooling applications tends to be lower as the driving temperature requested by the chiller is considerably higher.

Fig. 11- 30 Average yearly collector efficiency in selected installations monitored over one or more cooling seasons



Size of the hot storage volume The selection of the volume of the hot storage depends mainly on the following aspects: 1. What is the typical time for which building loads (heating, cooling) shall be covered by the heat storage but no or limited solar radiation is available at the same time? Or in other words: how big must a heat storage be in order to minimize not used solar gains, i.e., solar gains which cannot be used due to missing heat loads at the same time? 2. Is a cold storage available in the system or not? The hotels located in tourist areas, e.g. at the sea side, might be used as an example: In such applications the main heat load for DHW and thermally driven cooling usually occurs in the evening hours. In these cases, the storage tanks have to be large enough in order to store the thermal energy produced by the collectors during the day which cannot be used at the same time in order to be available in the late afternoon and evening hours. In office buildings, the opposite might be the case (see example of the research centre CARTIF discussed earlier in this chapter) when the peak load occurs in the morning hours due to standard office times. In Figure 11- 31, the probability distribution of the hot storage volume per installed solar thermal collector area of 55 documented installations is shown. As can be seen the mean value is located around 50 l hot storage volume / m² solar thermal collector field.

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Fig. 11- 31 Probability density of the ratios between hot water tank volume and collectors’ area in documented systems

 In Figure 11- 32 furthermore the probability density curve of the hot storage volume per nominal capacity of the thermally driven chiller of 43 documented installations is shown. As can be seen the mean value is located around 75 l hot storage volume / kW cooling capacity of the absorption chiller. In order to have a more useful representation, single installations with ratios higher than 200 lHWT/kWAbs are not shown in the figure.

Fig. 11- 32 Probability density of the ratios between hot water tank volume and installed cooling capacity of the thermally driven chillers in documented systems

 Due to a temporal mismatch between solar radiation and load needs the selection of the volume of the hot storage depends on whether a storage tank is also installed on the cold side. If a cold storage is installed, it might be the choice of the control to produce cold water at the moment when the necessary heat is available and therefore use the cold storage mainly to store the energy. However, this choice has not been a common approach up to now. The number of documented installations using a cold storage is too small to draw clear conclusions on average values.

11.3.3 Control strategies The control of solar cooling systems plays an important role in order to minimize energy use and cost while maintaining internal comfort requirements. The control of a solar cooling installation is largely dependent on its configuration and application. The design of a solar cooling system, like the design of any Heating Ventilation and Air Conditioning (HVAC) system, is a trade-off between initial cost (primarily the initial capital expenditure), operating cost (maintenance and primary energy cost), and indoor comfort. In general the objectives of control include; –– Safety –– Autonomous operation of system

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–– –– ––

Minimum cost of energy Maximum occupant comfort Fault detection and correction

HVAC systems in buildings are typically controlled using a two-level control structure. The more general level of supervisory control specifies set points and time dependent modes of operation for the lower level local-loop control, which attempts to meet the set point using an actuator. The following section focuses on higher level supervisory control strategies. The main goals of solar cooling control strategies are: –– Avoid malfunctioning of both the overall system and the various technical components and to ensure trouble-free operation –– Meet the cooling loads (i.e., guarantee certain indoor thermal comfort conditions in terms of relative humidity and temperature of indoor air) –– Minimize the primary energy use. This means that the implemented energy consumers (pumps, ventilators, chiller etc.) of the overall system have to operate with the lowest possible energy consumption and that the use of solar energy for operation of the thermally driven cooling system and other building loads has to be maximized

11.3.3.1 Overall control and configuration of SHC systems Closed cycle thermally driven cooling systems can typically be divided into three main fluid circuits. The three fluid circuits defined by the chiller are the generator or driving heat circuit, the evaporator or chilled water circuit, and the absorber/condenser or cooling water circuit.



Fig. 11- 33 Schematic drawing of the three circuits of sorption chillers /11.10/

From the perspective of the key components of solar thermally driven chiller-based systems, manifold control strategies can be identified according to these three circuits. One major controlled variable is the chilled water outlet temperature, which needs to be controlled depending on the cooling load and the type of chilled water distribution system. Due to the working principle

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of the thermally driven chiller, all of the input parameters – namely inlet temperatures and mass flows of all three hydraulic circuits – define the chilled water outlet temperature which therefore results in 6 potential control parameters. Conceptually, the control strategy can take into account the control of one or more of these parameters matching the required set-point of the chilled water temperature. The necessary technical field equipment is mainly based on actuators like valves, pumps, and cooling tower fans in order to control the system. In the following, various common or typical control strategies are presented by means of a general description of the control strategy and by more detailed information about the specific control issues. Some general boundary conditions for each hydraulic circuit result from the particular system boundary: –– Chilled water circuit: Depending on the applied air-conditioning distribution system, like fan-coils or chilled ceilings, different chilled water temperatures are required. Following the specific needs the chilled water temperature supplied to the distribution system is controlled and the control of the other circuits has to follow. –– Driving heat circuit: The general control task is to provide hot water at a temperature to the chiller according to the operating conditions recommended by the manufacturer. –– Heat rejection circuit: The cooling water temperature supplied to the chiller is controlled in order to operate the heat rejection system efficiently and to achieve high electrical coefficients of performance of the overall SHC system.

11.3.3.2 Control strategies of the chilled water circuits for absorption chillers Two general control strategies can be distinguished with regard to the chilled water temperature. The first one is the “free floating strategy” which allows different chilled water temperatures but aims at benefitting from solar heat as much as possible. The second one keeps the chilled-water temperature constant. These two control strategies can be characterized as follows: –– The “free floating strategy” is characterized by a chilled water demand large enough so that the chiller can operate whenever driving heat is available by means of maximizing the cold production at the cost of a variable chilled water temperature. This strategy is applied in solar autonomous heat driven chiller-based systems with no integrated back-up system and/or in assisted SHC systems if there is a sufficient cooling demand and/or cold storage volume available. –– The “constant chilled water temperature strategy” alternatively represents an active attempt by the control equipment to reach and to keep a chilled water temperature set-point, at the cost of potentially reducing the chilling capacity. In the case the distribution system is designed for constant or very low chilled water temperature – like fan-coils designed for dehumidification – the SHC system has to produce constant chilled water temperatures. In most cases, solar heat cannot be the only driving heat source and back-up systems have to be integrated in the overall SHC system in order to guarantee constant chilled water temperatures.

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Additionally, it is possible to control the chilled water temperature by varying the mass flow of chilled water using a variable speed pump or mixing the chilled water by using a mixing valve on the chilled water circuit. The control options for the chilled water temperature for example strongly depend on the system configuration and type of chilled water distribution. For systems with a low temperature chilled water distribution, with cold supply temperatures close to the lower limit of the absorption chiller (e.g., 5–6 °C (41–42.8 °F)) the chilled water storage does not provide a real storage capacity (this is only valid for LiBr-water chillers but not for ammoniawater chillers which are able to produce cold below the freezing point of water. In this case the chilled water outlet temperature of the thermally driven chiller (TDC) needs to be controlled either by the generator inlet temperature, the evaporator mass flow rate or by the control of the cooling water temperature. However, the control of the generator inlet temperature in a range of 70 to 95 °C (158 to 203 °F) reduces the temperature level in the hot storage tank at part load conditions and therefore improves the efficiency of the solar system. Furthermore, the number of system shutdowns and startups are significantly reduced which improves the overall system efficiency. The control of the evaporator mass flow rate is able to assure a stable chilled water supply temperature at variable generator inlet temperatures, but does not reduce the number of system shutdowns and startups at part load conditions. The control of the cooling water supply temperature reduces or improves the thermal COP of the Thermally Driven Chiller to assure a stable chilled water supply temperature at variable generator inlet temperatures. For the overall system efficiency, the increase of cooling water temperature is justified if the electricity consumption of the cooling tower is significantly reduced through a fan speed control and no backup system is used for the heat supply. A control of the cooling water temperature through a three way valve is not recommended in this case. For systems with higher temperature chilled water distribution (e.g. 16 °C (60.8 °F) supply temperature) the chilled water storage offers a useful storage capacity, if the produced chilled water temperature is allowed to drop significantly below the chilled water supply temperature set point at part load conditions. In this case no generator inlet temperature control or evaporator mass flow control are required. If flexible generator inlet temperatures according to the temperature level in the hot storage tank are allowed, the backup system should only be in operation if the chilled water temperature in the chilled water storage tank increases above the required chilled water supply temperature. For high system efficiency, the backup heating system should provide hot water at the upper limit of the absorption chiller (e.g. 95 °C (203°F)).

11.3.3.3 Control strategies of the driving heat circuits The primary task of the heat production system is to provide a sufficient driving heat temperature and capacity where the sorption chiller can operate in an energy efficient way and provide the required cold. The control strategy should prioritize the use of solar heat instead of the thermal back-up solution in the case of a heat generation system using a hot back-up.

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To operate the solar thermal system, the control strategy of the primary solar circuit is based on criteria about when to start and how to operate the solar system. Various starting criteria are applicable, such as the temperature difference between the hot water storage tank and the solar collector, or a minimum solar irradiation value on the solar collector. In the case that the primary circuit uses a variable speed pump, the pump can be controlled to obtain a given temperature at the outlet of the solar collector system.

11.3.3.4 Control strategies of the cooling circuit There are two main strategies in operating a thermally driven chiller in order to meet the required conditions in the chilled water cycle: –– The chiller operation is controlled by adjustment of the driving heat temperature. All other chiller inlet temperatures are not controlled and are dependent on the overall system conditions such as cooling load and weather. –– The chiller is operated by actively controlling the cooling water inlet temperature supplied by the heat rejection system. This second control strategy is far less used in comparison to the first one but can lead to important reductions of the electrical energy consumption of the heat rejection system [12]. As a practical example the installation in Skive – Denmark was presented earlier in this chapter.

11.3.3.5 Recommendations on control issues In the following, general statements and recommendations related to control strategies are listed, these have been stated during the collection of experience from several SHC plants. Limited control standards – Up to now, most systems are unique and only a few system providers offer comprehensive control strategies. The development of standard control strategies for a wide range of different system configurations is currently still a task limited to research and single market actors. Therefore the effort to develop and to implement a suitable control strategy of a system should not be underestimated. Complexity of control – It is obvious that the complexity of the control strategy depends strongly on the complexity of the hydraulic scheme. The control strategy can be developed if the system configuration is fixed and the comfort and system requirements are defined. A complex system implies a complex control strategy leading to a higher risk of misunderstanding or mal-function during operation. Choice of the control equipment – Solar-assisted air-conditioning systems include various numbers of components, which need to be controlled efficiently in order to cover the cooling requirements and to meet expectations in overall system efficiency. On the other hand, the implementation of control features (sensors, flow meters) increases the system costs having a certain influence on the economics of the system. Therefore the control devices should be selected in order to meet

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the highest possible system efficiency (primary energy saved, electric efficiency) with the lowest possible control effort (complexity, cost). Minimum required monitoring equipment – In order to monitor the energy performance of a system, it is highly recommended to implement extra hardware and software. Normally the sensor equipment and measured data points of the implemented control system do not fulfill the requirements of an appropriate assessment of system performance. . Experience has shown that through evaluation of monitoring data and following optimization measures the systems performance can be strongly enhanced. Therefore installation of a minimal monitoring system is recommended. High solar fractions – If a SHC system with a thermally driven chiller is employed and a fossil-fuelled heat source is used as the back-up, a high solar fraction is necessary in order to achieve significant primary energy savings. An appropriate design of the solar system, i.e., suitable dimensioning of the solar collector and system-integrated energy storage, is necessary for this purpose. The control strategy should facilitate the use of as much solar heat as possible. Especially for solar-assisted airconditioning with additional heat sources such as cogeneration units, district heating and/or boilers, the control priority should be given to solar heat followed by the other heat sources. Adaptability and flexibility – A key target of the control strategy of a system is to minimize the primary energy use per produced kWh cold or conditioned air. That means for the control to drive the implemented energy consumers (pumps, ventilators, chiller etc.) of the overall system at their minimal energy consuming operation point. Consequently the control strategy has to be flexible and adaptable in order to optimize the overall system. The control system and software should enable easy adaptation of control algorithms, parameters and set values. This is of importance as in many cases the use of the building may change, leading to a relevant change of the temporal demand distribution. In many of the above presented case studies the applied control software and hardware is more or less open source. In the following, some precise hints are given on the experience of plants where the hydraulic scheme and the control has been modified after the first seasons of operation. With regard to the cooling tower, a fan speed control is recommended instead of a three-way valve for the control of the cooling water temperature. With regard to the pumps, it is recommended to use high energy efficient variable speed pumps and ensure a good hydraulic design with low pressure drops (e.g. tube cross section, tube length, rambling tubing). With regard to the chilled water distribution system, a mass flow control of the distribution pump is recommended which keeps a stable temperature difference between supply and return. If a chilled water distribution systems with high supply temperatures (e.g. 16 °C (60.8°F) / 18 °C (64.4°F)) and cold water storage is applied:

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–– ––

Allow variable generator inlet temperatures according to the temperature in solar hot water storage tank. Choose supply temperatures from the backup-system as high as possible (generally limited to 95 °C (203°F) for a single effect chiller but higher than a traditional nominal temperature produced by a standard condensing gas boiler fixed at 60 °C (140°F)) for a high thermal COP of the thermally driven chiller.

If a chilled water distribution system with low supply temperatures (e.g. 7 °C (44.6 °F) / 12 °C (53.6 °F)) is applied, it is recommended to control the generator inlet temperature in order to reach a constant cold water outlet temperature of the thermally driven chiller.

11.3.4 Commissioning The aim of the commissioning process can be in general described as follows: “The commissioning process makes sure that new or reconstructed buildings will be of the same specifications as have been agreed upon in the contract documents during the planning stage and in the owner’s project requirements” /11.11/. For solar heating and cooling processes this means that after finishing the design and installation phase the systems functionality is checked at the moment of the handover of the system to the custumer. But it also means that the energy performance figures are monitored and checked and compared to the performance figures declared in the design documents. This cannot be done during a single commissioning phase but needs monitoring of key energy performance figures for a certain period (at least a period covering all typical operation conditions). This is possible only if a monitoring system is installed that for the analysis of the energy fluxes within the system. Such a monitoring system and the control and evaluation of data requires a certain investment. On the other hand the extended commissioning of a complex energy system leads in many cases to energy efficiency improvements of 20% or more without reducing the comfort for the building users /11.12/. These improvements of the efficiency of the system result in reduced operation costs and therefore often lead to an overcompensation of the extra initial investment cost within the first years of operation. Based on the monitoring results of solar heating and cooling systems worldwide, the authors strongly recommend the execution of an extended and comprehensive commissioning process. As the available experience among the actors within the construction phase and the degree of standardization of solar heating and cooling systems is considerably lower than that of conventional HVAC systems, the probability of mistakes, malfunctions or non-optimal control are considerably higher. Within Task 38 of the Solar Heating and Cooling program a technical report on commissioning was produced /11.13/. This is downloadable for free from the website. In this report relevant

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i­nformation and references to the commissioning process have been collected. Furthermore a check list to be applied specifically for solar heating and cooling systems has also been produced and is also available for download.

/11.1/ /11.2/ /11.3/

/11.4/

/11.5/

/11.6/

/11.7/

/11.8/

/11.9/

/11.10/ /11.11/ /11.12/ /11.13/

Thür, A., Vukits, Solar Heating and Cooling for the solarcity Gleisdorf. Journal of sustainable energy Vol 1, NO 2, June 2010 Vukits, M., Thür, A., Solar Heating and Cooling for the City of Gleisdorf – Optimisation of the Control Strategy, EUROSUN 2010, Graz, Austria Marc. O, et al, Experimental investigation of a solar cooling absorption system, operating without any backup system under tropical climate. Journal Energy & Buildings, Volume 42, June 2010 E. Wiemken, J. W. Wewiór: Solar air-conditioning in the German Solarthermie 2000plus programme: installed plants and first monitoring results. 3rd International Conference Solar Air-Conditioning; Palermo, Italien, 30. Oktober – 2. September 2009 Troi, A., Filippi, H. and Sparber, W. Practical Experience with Solar-Assisted Cooling in an Office and Educational Building in South Tyrol / Northern Italy. Wels: 1st International Conference on Solar Air Conditioning, 2005 Napolitano A., Troi A., Sparber W., “Monitoring activities for the optimization of a tri-generation plant including a solar heating and cooling system,” 2nd International Conference Solar Air-Conditioning – proceedings, October 2007: Tarragona, Spain L. A. Bujedo, J. Rodriguez-Santiago, P.J. Martícnez-Beltran, L.R. Rodríguez and J.Vicente Comparing Different Control Strategies and Configurations for Solar Cooling. Eurosun 2008, Lisbon (Portugal), 2008 P. N. Melograno, R. Fedrizzi, W. Sparber, G. Franchini, Test Procedures for Sorption Chillers Based on the Working Mode. Eurosun – International Conference on Solar Heating, Cooling and Buildings, Graz, October 2010 Annett Kühn, José Luis Corrales Ciganda, Felix Ziegler. (2008): Comparison of control strategies of solar absorption chillers, Proceedings of the 1st International Conference on Solar Heating, Cooling and Buildings (Eurosun), 7–10 October 2008, Lisbon, Portugal Paul Kohlenbach. Solar cooling with absorption chillers: Control strategies and transient chiller performance. PhD thesis, Technische Universität Berlin, 2006. ASHRAE (2005): ASHRAE Guideline 0–2005. The Commissioning Process. Atlanta General Services Administration (GSA) (2005): The building commissioning guide. Washington Uwe Franzke (2010): IEA SHC Task 38: Commissioning – a technical report of subtask B

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Chapter 12

DEC systems: built examples and experiences Responsible Author:

Marcello Aprile, Politecnico di Milano, Italy

Solar driven desiccant evaporative cooling (DEC) systems have increasingly raised interest during the last few years, even if desiccant technology covers a smaller share of the SAC market compared to absorption and adsorption chillers. Such systems are most suited for supplying fresh air to indoor spaces, such as offices, lecture and conference rooms, and leisure and sports buildings. Therefore, their main purpose is to control indoor air quality and humidity although, depending on the particular application, they might also supply sensible cooling. According to the Task 38 survey on SAC existing installations, there exists a clear predominance of the solid desiccant technology, with system sizes ranging from 1,250 to 18,000 m³/h (10,594 cfm). System sizes of 60,000 m³/h (35,315 cfm) and above are achieved by the DEC liquid technology. Regarding the interaction with solar systems, liquid-based flat plate collectors and air-based collectors are the most common choice, although CPC and ETC also have applications. The complexity of DEC systems varies with the type of heat source and the degree of integration with the overall heating and cooling distribution plant. For example, the DEC cycle can be driven by solar heat only or a mix of solar and other sources (e.g. district heat, natural gas, waste heat from a co-generation plant or biomass). Moreover, the solar based heat generation system can simultaneously meet heat demands other than the DEC system, such as the domestic hot water preparation system. Finally, the cooling capacity of the DEC cycle can be enhanced by means of an external supply of cooling power, which can be provided by conventional (i.e. mechanical vapour compression) or thermally driven (i.e. absorption/adsorption) water chillers. For some of the existing installations, the resulting picture is that of a complex custom made system. In the following, selected monitored and documented solar DEC plants will be described in detail with the main focus on the air handling unit (AHU) subsystem and its performance. The aim is to present built examples of solar-assisted DEC plants and to comment on the hydraulic scheme and the air handling unit configuration. Since flexibility exists in setting up the control strategy of such plants, as a consequence of the different operation modes under which a DEC system can run, space is dedicated to the description of the implemented control logic. Summary performance

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figures are provided at the end of the plant description: as the presented cases all differ from each other in many aspects (load, climate, heat source, AHU configuration), care should be taken in comparing the performance of the different plants. The chapter ends with a summary of the lessons learned throughout the experimental activities within Task 38, with the aim of providing useful technical hints to planners. Advice concerns the preliminary design decisions, such as the choice of the system configuration and the key components (sorption rotor, rotary heat exchanger, humidifiers and fans), the sizing of the main system components and the setting up of the control strategy.

12.1 Built examples 12.1.1 ENERGY base The ENERGYbase office building in Vienna, Austria, was designed as a passive house building with renewable technologies to cover the energy demand for heating, cooling, ventilation and lighting. The energy systems of ENERGbase are designed to use both a water based and an air based energy distribution system. The basic temperature control of the office areas is realized by the water based heating and cooling system. The air treatment system is designed in order to control indoor humidity and to supply fresh air to the offices. In particular, air treatment during summer is provided by a 100% solar thermally driven DEC system, composed of two identical units of 8,240 m³/h (4,850 cfm) each (see Table 12- 1). In cooling mode, the air treatment system was conceived to supply fresh air to the building at a moderately cool temperature (22 °C (71.6 °F)) and relative humidity below a maximum value (80%). Location: Vienna, Austria Latitude

48°12’ N

Longitude

16°22’ E

Elevation

220 m

Application

Office

Solar collector field

285 m²

Heat storage capacity

15 m³

Conditioned area

5,025 m² (529.7 ft³)

DEC system volume flow rate

2 × 8,240 m³/h (4,850 cfm)

Tab. 12- 1 ENERGYbase plant design data

Fig. 12- 1 The ENERGYbase building

The simplified hydraulic scheme is shown in Figure 12- 2 (only one of the two units is shown). During the summer, the solar system supplies hot water to the regeneration coils of the two DEC

334

12.1 Built examples

systems through independent circuits, of which the supply temperature is controlled by means of three-way mixing valves, this is because the sorption rotor has to be protected from air temperatures above 70 °C (158 °F). Each unit is assembled with the following equipment: –– Li-Cl sorption rotor, casing width 1770 mm, depth 450 mm. –– Aluminum rotary sensible heat exchanger, casing width 1770 mm, depth 400 mm. –– Spray humidifiers with 4 nozzles, water recirculation and automatic cleaning process of the humidifier water basin. –– Bypass flaps around the sorption rotor and the rotary heat exchanger, on both the supply and return air streams. –– Auxiliary electrical heater on the regeneration air stream (for emergency only). –– The bypass across the supply humidifier is meant for winter operation only, when a fraction of the supply air stream is first diverted to indoor garden spaces before being supplied to the office rooms.

solar collector

heat exchanger T

heat storage 45C°

T

T

filter

humidifier

garden space

heater

fan

heater

fan

bypass

bypass

filter

Ambient

humidifier

Sorption wheel

filter

bypass back-up heater

Heat exchanger

room

bypass

Fig. 12- 2 Simplified hydraulic scheme (the electrical back-up heater is used for emergency cases only)

The controlled variable is the supply air temperature, which is set at 21 °C (69.8 °F) in heating mode and 22 °C (71.6 °F) in cooling mode. Different operation modes are activated in sequential order, according to the load. The priority is assigned from the least to the most energy consuming mode. During cooling mode, the control sequence is set as follows: 1. ventilation: variable speed fans are used and controlled according to the static pressure difference on the supply and return flow, all other equipment is turned off and the bypasses around the rotary heat exchanger and sorption wheel are open

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12  DEC systems: built examples and experiences

2.

3.

4.

5.

6.

heat recovery: the bypass around rotary heat exchanger is closed, the rotary heat exchanger is turned on when the return temperature is lower than the outdoor temperature, the speed is PI controlled up to 10 RPM return air adiabatic cooling: the return humidifier is turned on when the return temperature is higher than the outdoor temperature, the frequency of the pump inverter is PI controlled between 20 and 50 Hz solar heat regeneration: the bypass around sorption rotor is closed, the wheel is turned on and the speed is set at 20 RPH, the two-way valve on the hot water return from the regeneration coil is opened and PI controlled, the regeneration air temperature is limited up to 70 °C (158 °F) supply air adiabatic cooling: the supply humidifier is turned on, the frequency of the pump inverter is PI controlled between 20 and 50 Hz and the relative humidity at the humidifier outlet is limited up to 80% auxiliary heater regeneration (emergency case only): in an emergency case the electrical heater is manually turned on when the solar plant is not working and no heat from the hot buffer storage is available; the electrical heater is PI controlled and the regeneration air temperature is limited up to 70 °C (158 °F)

During winter, both the sorption rotor and the rotary heat exchanger are used for energy recovery. In order to prevent frost formation, the return air at the sorption rotor inlet is kept above 8 °C (46.4 °F) and the sorption rotor slows down and eventually stops as the return air at the sorption rotor outlet approaches 0 °C (32 °F). When the rotary heat exchanger is also used, return air at its outlet is maintained above 9 °C (48.2 °F) in order to cope with the minimum temperature requirement at the sorption rotor inlet. The purpose of the air handling units is to supply primary air, thus the cooling power is given by the enthalpy difference between supply and outdoor air, multiplied by the supply air mass flow rate. In summer 2010, the following monthly performance figures were reported: –– cooling and heating energy supplied to the fresh air stream –– regeneration heat –– solar collector output –– available solar energy input –– electricity consumption of the DEC system (without the contribution of solar pumps) –– electricity consumption of the solar pumps The electricity consumption of the DEC system includes the fans, the distribution pumps from heat storage to AHU coils, the motors of the two rotors and the pumps of the humidifiers. From the available data, the following performance indicators have been calculated: –– Thermal COP –– Solar energy utilization The results are shown in Table 12.2.

336

12.1 Built examples

Month

Jun

Jul

Aug

A) Cooling energy, kWh

3,971

6,980

4,879

B) Heating energy, kWh

438

331

405,9

C) Regeneration Heat, kWh

7,990

12,994

7,924

D) Solar collector output, kWh

8,654

14,348

8,516

E) Available solar energy input, kWh

44,707

48,859

42,011

F) Electricity DEC, kWh

2,456

2,736

2,491

G) Electricity solar, kWh

39

60

38

Thermal COP, (A/C)

0.50

0.54

0.62

Solar energy utilization factor, % (D/E)

19%

29%

20%

Tab. 12- 2 ENERGYBase plant energy performance in summer 2010

The thermal COP values (0.50–0.62) are in agreement with the state-of-the-art, although the solar plant is probably oversized. Finally, the electricity consumption of the solar loop is negligible if compared to that of the DEC system.

12.1.2 Munich Airport In spring 2008 Munich Airport decided to realize a pilot plant for solar air conditioning for one of the canteens. The plant was erected at the end of 2008 and the commissioning took place in February 2009. In the following period, a continuous monitoring and optimization of the plant and the control strategy was conducted. The cargo canteen of Munich Airport has an area of 400 m². Around noon, the canteen space is occupied by about 100 to 120 people. Location: Munich, Germany Latitude

47°8’ N

Longitude

11°35’ E

Elevation

519 m

Application

Canteen

Solar collector field

75 m² (807.29 ft²)

Heat storage capacity

3 m³ (105.944 ft³)

Conditioned area

400 m² (4305.6 ft²)

DEC system volume flow rate

10,000 m³/h (5,885.78 cfm)

Tab. 12- 3 Munich Airport design data

Fig. 12- 3 Exterior view with the liquid desiccant cooling system on the top and the solar collector field in front of the canteen

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12  DEC systems: built examples and experiences

The desiccant cooling system (¹) is liquid-based and mainly driven by solar heat, with district heat as backup. The liquid desiccant cooling system (¹) is equipped with speed controlled fans and an improved heat recovery unit in order to reduce auxiliary energy demand and running costs. Furthermore, the occupied area is equipped with CO2 sensors and thus the volume flow can be controlled in depending on the real occupancy of the canteen. The innovative dehumidification unit uses a lithium chloride solution as the absorbent. The driving heat for the process must be in a temperature range of 55 to 70 °C (131 °F to 158 °F). The system is sized in order to improve the conditions for the integration of solar heat. First, the air heater for the winter mode is designed for a maximum inlet temperature of 45 °C (113 °F). Second, the amount of lithium chloride solution is enough for dehumidifying the ambient air for two hours by 4 g/kg without a heat supply. The AHU is mounted on the roof of the building. The driving heat for the liquid desiccant cooling system is delivered by a 75 m² (807.29 ft²) flat plate collector field. The specific collector type (¹) is filled with inert gas and hermetically sealed. Besides providing a higher thermal insulation, air tightness prevents the flue gases at the airport from coming into contact with the selectively coated absorbers in the collector. The 36 collectors are divided into 6 equally sized collector fields, connected in parallel. Measurements during the commissioning phase confirmed that the flow distribution between the 6 fields is very uniform. The collectors are mounted on a flat roof with gravel filled plates and facing south with an inclination of 30° (86 °F).The solar system is divided into primary and secondary circuits, separated by a plate heat exchanger. To reduce the temperature stress on all the components and thus increasing durability, a dry cooler removes excess heat from the collector field, i.e. heat which cannot be used due to a reduced heat load. Figure 12.4 shows the hydraulic scheme of the solar air conditioning system. The buffer storage is solely heated by solar energy. The back-up heat is integrated “in line” with a plate heat exchanger. This decreases the losses of the buffer storage and improves the conditions for a higher collector yield.

338

12.1 Built examples

solar collector

heat exchanger T

dry cooler

backup

heat storage 45C°

T

T

room

filter

heater

fan

desorber

fan

fan

absorber

filter

Ambient

Ambient spray heat exchager

Fig. 12- 4 Hydraulic scheme of the solar AC system at Munich Airport. The dry cooler is only used in cases when solar heat cannot be used due to a reduced heat load (excess heat)

Figure 12- 5 shows the thermal energy flow diagram for the period June 2009 to April 2010. Since the AHU supplies an all-air system for space heating and cooling, the delivered energy at the end user is calculated as the enthalpy difference between supply and return air multiplied by the supply air mass flow rate. It can be seen that only 673 kWh (2,296,371.32 BTU) of excess heat from the collector field had to be removed to the ambient. This indicates that the collector field is almost perfectly balanced to the system.



Fig. 12- 5 Energy flow diagram (June 2009 to April 2010)

The thermal COP for the period from June 2009 to April 2010 was 1.2; note that only cooling operation was taken into account for the calculation of this value. This higher value is possible due to the efficient use of evaporative cooling within the process. The system enables a very efficient use of heat recovery during heating operation by pre-heating fresh air using warm building return air in the

339

12  DEC systems: built examples and experiences

plate heat exchanger. The electrical COP for the same period was 4.9 for cooling, 53.5 for heating and 20.5 for cooling and heating. It has to be noted that in the calculation of the electrical COP, the energy consumption of the fans is not included. The specific collector yield was estimated to be 407 kWh/(m² year), with a peak of 73 kWh/m² during August. (1 m² = 10.76 ft², 1  Wh = 3,412.14 BTU)

12.1.3 DREAM Unipa A small scale solar air-conditioning system addressing air conditioning in hot humid climates has been installed and tested at the Department of Energy and Environmental Research – Università degli Studi di Palermo (Unipa), Italy. The SAC system is composed of the solar system, a DEC air handling unit, a vapour compression chiller and a radiant ceiling system (see Table 12- 4). A gas boiler is also used as back-up. Location: Palermo, Italy Latitude

38°11’ N

Longitude

13°36’ E

Elevation

14 m

Application

Office

Solar collector field

22.5 m² (242.19 ft²)

Heat storage capacity

0.6 m³ (21.19 ft³)

Conditioned volume

450 m³ (15,891.60 ft³)

Radiant ceiling surface

78 m² (839.58 ft²)

DEC system volume flow rate

1,500 m³/h (882.87 cfm)

Chiller cooling power

24 kW (6.72 ton)

Chiller rated EER

3.47 –

Tab. 12- 4 DREAM plant design data

Fig. 12- 6 The DREAM DEC system

The DEC system, which is mainly driven by solar heat, provides fresh air to an office room of 450 m³ (15,891.60 ft³). The chiller supplies cooling power to both the DEC system and the radiant ceiling system. Additionally, the DEC system makes use of a fraction of the chiller condensation heat for pre-heating of the regeneration air (see Figure 12- 7). The DEC air handling unit is equipped with a silica gel sorption rotor (diameter 700 mm, depth 200 mm), an aluminum rotary heat exchanger (diameter 700 mm, depth 200 mm), an evaporating pad, a condensing cooling coil, a non-condensing cooling coil, a regeneration pre-heating coil and a regeneration heating coil. Since the humidity ratio of outside air is normally high and the required humidity ratio of supply air is very low, no humidifier on the supply side was installed. The two cooling coils and the regeneration heating coil are controlled by means of a three-way

340

12.1 Built examples

mixing valve on the return line, in a variable flow – constant supply temperature arrangement. During the typical cooling operation (design conditions), outdoor air at 36 °C (96.8 °F) and 20 ­g/kg humidity ratio is cooled to 18–20 °C (64.4 °F to 68 °F) and dehumidified to 8 g/kg humidity ratio. This is achieved by means of a three stage process: condensing cooling, DEC and non-condensing cooling. The first stage ensures that air humidity does not exceed 15 g/kg before entering the sorption rotor. The DEC stage allows dehumidification of the supply air at the desired humidity ratio. The final stage cools down supply air at the desired temperature. The DEC system operation is defined according to outdoor and indoor conditions. The controller sets the seasonal operation mode (cooling, heating or ventilation) according to the outside temperature: heating below 20 °C (68 °F), cooling above 24 °C (75.2 °F) and ventilation in between. In ventilation mode, only the fans are switched on. In order to reduce power consumption, the supply air intake downstream of the sorption rotor opens and the return air stream is diverted to the bypass duct around regeneration coil and sorption rotor. solar collector

solar loop

backup

heat storage

T T

T

bypass

filter

fan

cooler 1

filter

Ambient

cooler 2

heater 2

Sorption wheel

fan

heater 1

humidifier

Heat exchanger

room

filter

ambient load

auxiliary chiller

dry cooler

distribution

Fig. 12- 7 Hydraulic scheme (only the cooling case is drawn); for heating a connection between solar loop and cooler 2, which then is used to preheat supply air, is possible

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12  DEC systems: built examples and experiences

The cooling operation modes are set in sequence as follows. 1. Indirect evaporative cooling – The rotary heat exchanger is turned on at 10 RPM and the return humidifier is switched on at full power, the supply air intake is downstream of the desiccant wheel and the bypass of the desiccant wheel is open. 2. Desiccant cooling – The sorption rotor is turned on at 15 RPH, solar heat is supplied by opening the regeneration coil three-way valve, the supply air intake is upstream of the condensing coil and the bypass of the desiccant wheel is closed. Due to operation of the cold radiant ceilings, the vapour compression chiller is often in operation and pre-heating of regeneration air downstream of the rotary heat exchanger occurs. 3. Auxiliary cooling – The auxiliary condensing and non-condensing cooling coils are activated to meet the cooling loads in the case that the desiccant cycle is not sufficient to reach the desired supply air conditions (20 °C, 8 g/kg). The two coils are controlled independently. The condensing coil outlet humidity ratio is maintained below a value of 15 g/kg, whereas the non-condensing coil air outlet temperature is set at 20 °C (68 °F). The control of the AHU in heating operation has two operation modes. 1. Heat recovery – In this mode of operation the recovery of the heat related to the return air is performed by means of the rotary heat exchanger. The supply air intake is lateral, downstream of the desiccant wheel and the bypass of the desiccant wheel is open. 2. Active heating – Besides the heat recovery, the non-condensing coil is now operated as a heating coil providing solar heat to the supply air stream (not shown in the hydraulic scheme in Figure 12- 7.). A set point of 30 °C (68 °F) for supply temperature is used. Since the purpose of the plant is to supply primary air, the cooling power is calculated as the enthalpy difference between supply and outdoor air, multiplied by the supply air mass flow rate. During summer, the following energy performance figures were monitored (see Table 12- 5): –– total cooling energy supplied to the fresh air stream –– cooling energy supplied to the fresh air stream by the chiller (through cooler 1 and cooler 2) –– regeneration heat from solar (through heater 2) –– regeneration heat from chiller condenser (through heater 1) –– solar collector output –– available solar energy input –– electricity consumption of the DEC system (without the contribution of solar pumps) –– electricity consumption of the solar pumps –– electricity consumption of the chiller The cooling energy supplied to the radiant cold ceiling by the chiller was not monitored because the radiant cold ceilings were not used during the first monitoring campaign. The electricity consumption of the DEC system includes the fans, the distribution pumps from the heat storage to the AHU coils, the motors of the two rotors and the pumps of the humidifiers. On the basis of the monitored data, the following performance indicators were calculated:

342

12.2 Experiences

–– –– ––

Thermal COP Solar energy utilization EER of the vapour compression chiller

Due to the DEC hybrid scheme, the definition of thermal COP is not straight forward. Thermal COP can be assessed under the perspective of the classical DEC cycle by isolating the contribution of the chiller. In particular, the cooling supplied by the chiller is subtracted from the total AHU cooling effect, whereas the heat rejected by the chiller and recovered in the AHU as regeneration heat is added to the solar regeneration heat. Month

May

Jun

Jul

Aug

Sep

A) Total cooling energy to supply air, kWh

599

2,000

3,374

2,846

1,532

B) Cooling energy to supply air from chiller, kWh

146

807

1,941

1,138

1,007

C) Regeneration heat from solar, kWh

525

1,232

1,316

1,390

829

D) Regeneration heat from chiller condenser, kWh

214

647

722

451

348

E) Solar collectors output, kWh

808

1,765

1,876

1,962

1,396

F) Available solar energy input, kWh

2,087

4,556

4,741

4,698

3,481

G) Electricity DEC, kWh

239

514

589

565

446

H) Electricity solar, kWh

4

9

9

10

7

I) Electricity chiller, kWh

50

330

691

508

382

Thermal COP, (A–B)/(C+D)

0.61

0.63

0.70

0.93

0.45

Solar energy utilization factor, % (E/F)

39%

39%

40%

42%

40%

Chiller average EER, B/I

2.95

2.44

2.81

2.24

2.64

Tab. 12- 5 DREAM Unipa plant performance in summer

12.2 Experiences DEC system configuration Among the most recent installations of solar-assisted DEC systems, two main system configurations have been implemented: 1. Classical DEC cycle with solar system and back-up heat source 2. Hybrid DEC cycle with solar system and back-up chiller (electricity or heat driven, with or without condensation heat recovery) The choice of the best configuration is strongly dependent on climate and load. Due to its limited dehumidification capacity at low regeneration temperatures, the classical DEC system is best suited for moderate or low design outdoor humidity (e.g. below 14 g/kg).

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12  DEC systems: built examples and experiences

Supply air humidifier In moderate to low humidity climates, it is worthwhile exploiting direct evaporative cooling, especially at the beginning and towards the end of the cooling season. However, effective direct evaporative cooling requires fine controllable humidifiers. The humidifier should have fast response and should be controllable with sufficiently fine capacity increments between 0 and 100%. Moreover, freshwater treatment is normally necessary in order to prevent damage to the humidifiers and the diffusion of salt particles inside the supply air. All this might result in high installation costs and, if excessive or not-strictly-necessary humidifier cleaning occurs, this can result in a waste of electricity. In some circumstances (e.g. extreme hardness of the available freshwater), energy and economic feasibility should be assessed. It is possible that a hybrid solution with sensible air cooling could be a better choice. Pre-cooling for enhancing dehumidification capacity The hybrid solution with condensing air cooling, occurring before rather than after the sorption phase, is necessary in climates with high outdoor humidity (e.g. above 15 g/kg). In such conditions, direct evaporative cooling is useless during most of the operation time. Moreover, the heat rejected from the chiller can be recovered for pre-heating of the regeneration air. Return air humidifier Indirect evaporative cooling is often operated at full capacity, thus making on/off controlled humidifiers a suitable choice: an evaporative pad could be preferred to a more expensive air washer or water atomizer. Humidifier effectiveness, which realistically can achieve 90%, is negatively affected by air stream misdistribution and a proper AHU design should consider also this aspect (e.g. positioning and type of filters). Selection of the rotors For both the classical and the hybrid DEC system, selecting the proper desiccant sorption rotor and rotary heat exchanger is critical for the overall system performance and durability. The rotary heat exchanger effectiveness is determinant in achieving the desired capacity and thermal COP. As a general guideline, only heat exchangers having effectiveness higher than 75% should be used. Concerning the desiccant wheel, material stability should be considered since sorption components must withstand high thermal loads. Materials, indeed, may operate at high regeneration temperatures in summer conditions, when dehumidification is required, as well as at high relative humidity due to fog formation in winter. Both operating conditions should be considered in the design process. For example, frost formation can be avoided through a pre-heater. Dehumidification capacity is another important aspect. The planner should be aware that not all sorption wheels are designed for low regeneration temperatures. At fixed design conditions, the larger the dehumidification the better, and very good dehumidification capacity at low regeneration temperatures (60 to 70 °C (140 °F to 158 °F)) can be achieved only with high performing sorption rotors. Good performances can be achieved with both silica gel and LiCl-cellulose wheels. When considering LiCl-cellulose wheels, care should be paid in order to avoid possible damages from oversaturation (i.e. exposure to humid air for prolonged periods) or excessively high regeneration temperature.

344

12.2 Experiences

Purge sections across rotors The use of purge sections in both the sorption rotor and the rotary heat exchanger is not recommended. In the case their use is made necessary, special attention should be paid to the correct air pressure distribution across the rotors (this might affect the positioning and the number of fans). Purge sections can help to limit contamination of the supply air deriving from pollutants carried over in the return air. However, the best choice is to limit, when possible, contamination in the return air. For example, a good recommendation is to avoid the mixing of return streams from toilets and similar rooms: odours would be adsorbed by the sorption wheel and possibly released to supply air during wheel rotation. Rotor frame and sealing across channels Good sealing devices around the rotors are also very important in order to limit cross-leakages between supply air and return air. Cross-leakages are particularly detrimental on performance downstream from the heat exchanger when the return humidifier is running at full capacity: in such situations, mixing of dry supply air with saturated return air occurs, thus decreasing the dehumidification potential of the whole system. AHU internal bypass and openings Pressure drops within a DEC system can be much higher than those of a conventional AHU operating with the same flow rate, mainly due to the presence of the two rotors. In addition to the correct sizing of the rotors, some other solutions might help mitigate the parasitic consumption problem associated with the higher pressure drops. For example, additional air intakes and bypasses can reduce the total pressure drop when the desiccant wheel is not needed (e.g. during heating and adiabatic cooling mode) and efficient variable speed plug fans can lower the electricity consumption as compared to more conventional centrifugal fans. Sensors Sensors for control and supervision should be carefully installed. The most difficult measurements are downstream of the humidifiers and the sorption wheel, due to the strongly non homogeneous air temperature and humidity distribution in these zones. However, measurements at the two inlets and two outlets are sufficient for proper controlling and monitoring of the system. DEC sizing Undersized DEC systems will suffer from high room temperature and humidity, which in turn will cause the DEC cooling power to decrease and the room temperature and humidity to rise. Therefore, sizing should be addressed carefully. The sizing procedure of a DEC system is not different from that of a conventional AHU. The difficulty lies in the correct prediction of design data and in the gathering of the sorption rotor performance at the design point. If outdoor air conditions are relatively easy to set, the same cannot be said for return air conditions. In particular, depending on the air ventilation system and the room shape, the return air temperature could be higher than the temperature set in the occupants’ zone. As a consequence, the heat recovery effect could be lower than expected in summer. Moreover, outdoor conditions vary with climate. When outdoor

345

12  DEC systems: built examples and experiences

humidity is higher than 15 g/kg, the dehumidification capacity of the sorption rotor is fully used and there is no margin for sensible cooling through evaporative cooling in the supply air stream. The combination of high return temperature and high outdoor humidity thus causes a too high supply air temperature. An example of the realistic capacity and expected performance of various system components of a classical DEC system is reported in Table 12.6, in which cooling power is defined as the total heat (latent and sensible) extracted from the supply air stream while flowing across the AHU. Simulation models can be of help not only in the design of the DEC process and the sizing of the heat sources including solar collector system but also for identification of process faults during system operation. In calculating supply air outlet conditions, the fan heat should be considered, as non-negligible air temperature increases, in the order of 0.5 to 1.5 °C (32.9 °F to 34.7 °F), are to be expected for each fan.

Supply air inlet conditions

32 °C, 12.0 g/kg

Return air inlet conditions

26 °C, 10.5 g/kg

Supply air outlet conditions

22 °C, 9.0 g/kg

Regeneration temperature

75 °C

Sorption rotor humidity drop

5.5 to 6 g/kg

Sorption rotor frontal air velocity

2 m/s

Rotary heat exchanger effectiveness

75 to 80%

Specific humidifier water flow (return or supply)

0.1 (l/min) / (1000 m³/h)

Specific cooling power (latent and sensible)

4.5 to 5.5 kW / (1000 m³/h)

Specific regeneration heat

8 to 10 kW / (1000 m³/h)

Typical range of thermal COP

0.50 to 0.65

Tab. 12- 6 Example of design data and expected performance of a classical DEC cycle (1°C=33.8°F, 1kW=0.28ton, 1m³/h=0.59 cfm)

Control tips Actuators of the classical DEC system typically include: fan inverters, sorption rotor and rotary heat exchanger inverters, regeneration temperature valve and pump switch, return humidifier switch and supply humidifier inverter, supply air heater valve and pump switch. During cooling mode, the sorption rotor rotational speed is set at a very low value, about 10–20 RPH, and the rotary heat exchanger at 10 RPM (such values are recommended for 2 m/s frontal air velocity and 200 mm depth wheels). Sorption rotor and rotary heat exchanger should have opposite rotational directions in order to enhance the heat recovery effect. Regeneration temperature can be varied in order to control dehumidification capacity.

346

12.3 Control strategy definition

During heating mode, two possibilities arise, depending on climate. In very cold climates, the sorption rotor is used for enthalpy recovery by increasing its revolution speed to a few revolutions per minute, at a minimum. The heat recovery wheel is not in operation. Practical experience has shown that the effectiveness of the desiccant wheel is better without the heat recovery and at very low outdoor temperatures, condensation and freezing in the desiccant wheel might occur if the relative humidity of the return air is too high (as it would be after heat recovery). In mild, cold climates, with sufficiently high outdoor humidity, the sorption rotor is not used and only heat recovery is effected. In both cases, the supply air heater brings the supply temperature to the desired set point.

12.3 Control strategy definition For DEC systems, the definition of a clear control strategy is not straight-forward and starting with a proven strategy is of fundamental importance in order to accelerate system optimization during commissioning. In principle, two types of control strategies can be set: temperature based and humidity-and-temperature based. In temperature based control, actuators are controlled in sequence as a function of the temperature mismatch, i.e. the difference between air target and actual temperatures (see Figure 12- 9). During summer, an additional limit to the supply humidifier outlet humidity could be set in order to avoid raising the humidity in the conditioned space (e.g. maximum supply air humidity ratio to 10 g/kg). In humidity-and-temperature based control, the air humidity mismatch drives humidification in winter and dehumidification in summer, whereas the air temperature mismatch drives in sequence heat recovery and supply heater in winter and adiabatic cooling (both indirect and direct) in summer (see Figure 12- 10). The controlled temperature or temperature-and-humidity could be that of the supply, the return or the room air, depending on the application. In all-air systems, return or room air are suitable choices, whereas supply air is the usual choice in primary-air systems. Finally, target values should be adjusted according to the season, in order to exploit free ventilation as much as possible.

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12  DEC systems: built examples and experiences

Fig. 12 -9 Example of temperature based control. In cooling mode, the return humidifier (1) is on/off controlled by temperature (Tc), with deadband in order to avoid cycling. Simultaneously, the rotary heat exchanger (3) is proportionally controlled from 0 to 100% rotational speed. With the increasing of Tc, the supply humidifier (2) is activated with proportional control from 0 to 100% capacity along with sorption rotor (4) and regeneration heating coil (5). In heating mode, heat and humidity recovery is initially performed by the sorption rotor (4), which is proportionally brought to full speed for enthalpy recovery. With the decrease of Tc, the heating power exchanged at the supply air heating coil (6) is proportionally increased.

Fig. 12- 10 Example of humidity-and-temperature based control. In cooling mode, the return humidifier (1) is on/off controlled by temperature (Tc), with deadband in order to avoid cycling. Simultaneously, the rotary heat exchanger (3) is proportionally controlled from 0 to 100% rotational speed. With the increasing of Tc, the supply humidifier (2) is first activated with proportional control from 0 to 100% capacity. In turn, the corresponding raise of the controlled humidity (wc) triggers the functioning of the sorption rotor (4), which is quickly brought to the desired rotational speed for dehumidification (e.g. 15 RPH). At this stage, also the heating power at the regeneration heating coil starts to increase, until the maximum allowed regeneration temperature (e.g. 75 °C (167 °F)) is achieved. In heating mode, heat and humidity recovery is initially performed by the sorption rotor (4), which is proportionally brought to full speed for enthalpy recovery. With the decrease of wc, the supply humidifier is proportionally controlled. In turn, the corresponding decrease in Tc is offset by the increase in the heating power exchanged at the supply air heating coil (6).

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Chapter 13

Summary and outlook Responsible Authors:

Hans-Martin Henning, Fraunhofer Institute for Solar Energy Systems, Germany Mario Motta, Politecnico di Milano, Italy Daniel Mugnier, Tecsol, France

13.1 Overall technology status Today, there are different technological options for the use solar thermal energy in the air-conditioning of buildings or refrigeration services. Overall, the key components of solar cooling systems are market available and have reached a good level of maturity. Several types of solar thermal collectors, produced by numerous manufacturers in many countries, can be used to provide heat for the operation of thermally driven cooling equipment. Non-tracking solar thermal collectors are flat plate collectors, evacuated tube collectors or solar collectors for direct air heating, which can be operated in particular with desiccant systems. In the last few years, many efforts have been made to improve these collectors towards higher efficiency, even when operated at temperatures above those required for standard building applications (DHW and heating); these higher temperature levels are needed for process heat generation or for some thermally driven cooling equipment. Examples are double glazed flat plate collectors with antireflective coatings to minimize optical transmission losses or collectors using concentrators with low concentration factors to reduce thermal losses. The group of manufacturers of single-axis tracking solar collectors has significantly increased in the last decade. Today, different options are available for medium temperature collectors (operation temperature range from 120 °C to 250 °C (248 °F to 482 °F) and above). Parabolic trough as well as linear concentrating Fresnel solar collectors are commercial products. These collectors are among the few viable answers to the need of renewable heat at temperatures above 100 °C (212 °F); and they are particularly suitable for regions with a high ratio of direct solar over defuse radiation. Medium temperature collectors are well suited to be combined with high efficient thermally driven chillers such as double-effect or triple-effect absorption cooling machines. They can also be employed to produce a high temperature lift in combination with a single-effect chiller for instance in applications where cold at temperatures below 0 °C (32 °F) is needed and a dry heat rejection system has to be used in combination with high ambient air temperatures, e.g. in dry-hot climatic conditions. Thermally driven sorption chillers, that use either a liquid or a solid sorbent to produce chilled water or cold at temperatures below the freezing point of water, e.g. for ice production are m ­ arket

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available. Today, sorption chillers are available in a wide spectrum of capacities, from 8 kW 2.24 ton) up to multi megawatt systems. While in the small capacity range mainly single-effect systems are available, for large capacities double-effect machines and even triple-effect machines can be employed. Desiccant cooling air handling units can be employed for conditioning the ventilation air using low-temperature heat from solar thermal collectors. In such open cycles, mainly systems using solid sorbents fixed into the matrix of desiccant wheels are currently in use. But systems using liquid sorption materials have also entered the market in the recent years. These systems are well adapted for coupling with a solar heat source and provide an interesting way to store energy in the concentrated desiccant solution. From the technological point of view, it can be concluded that the two main subsystems of common solar cooling plants (i.e., solar thermal system and thermally driven cooling) are efficient, mature and reliable. Nevertheless there are still certain important shortcomings for solar air-conditioning and refrigeration at the system integration level. Many systems fail to achieve the planned energy savings because of failures in proper design and energy management that result in a high overall electricity consumption of auxiliary components. A large potential for improvement exists in the heat rejection subsystem, which has often not received sufficient attention. The choice of the correct component, the proper design of the heat rejection loop and its control aspects can have a deep influence on the solar cooling plant performance. The reduction of complex hydraulic schemes has been recognized as a main priority too. It has emerged that often the analyzed systems were far too complex and as a consequence reduced performance, non-optimal control and high maintenance efforts have been faced. However, examples exist that show a high level of overall system performance and lead to a high satisfaction of the users. Thus it is clear that it is possible to achieve good results, in terms of energy performance and user satisfaction, in cases of properly designed, installed and commissioned plants. A collection of various positive experiences has been given in this book. Some of the basic design principles, which are summarized below can be used to avoid the risk of insufficient performance.

13.2 Energy performance Solar cooling systems have proven to save energy when compared to conventional standard solutions. The amount of energy saved strongly depends on system design and operation. Key figures that determine the achieved energy saving are the solar fraction and the overall electricity demand for auxiliary components, such as fans (e.g., the fan in the cooling tower) and the pumps in the hydraulic circuits. Some basic design guidelines have to be followed in order to assure that energy savings are achieved for a particular system.

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13.3 Basic design guidelines and operation principles

Appropriately designed systems will lead to primary energy savings in the range of 25% up to 75% in comparison to standard reference solutions. The potential achievable savings depend on the load composition, the available space for solar collectors and on the overall energy efficiency of the conventional solution. In particular the primary energy ratio of electricity production has a direct impact on the primary energy efficiency of the production of cold with electrically driven vapour compression systems. As a result primary energy savings of solar cooling will be higher in countries where electricity is produced and distributed with low primary energy efficiency. Further high primary energy savings are achieved in particular if the solar collector system is able to serve other loads beside cooling, e.g. heating and domestic hot water production. This is true in building applications or for the generation of process heat or steam in the case of industrial applications. Solar heat can also be part of concepts which integrate various different heat sources such as industrial waste heat or heat from co-generation. Such designs can make sense in particular when solar heat can cover daily cooling peak loads by avoiding the operation of primary energy based backup sources. Some of the best performing systems today achieve an electric COP (see chapter 7) value of 6 to 8, (average value for a complete cooling season), meaning 6 to 8 units of useful cooling per unit of electricity used in the overall system. In principle, values above 10 can be reached and future efforts have to assure that these values will be achieved in realized installations. With this performance index, solar cooling becomes the most efficient renewable cooling technological option in terms of primary energy, today available.

13.3 Basic design guidelines and operation principles A conclusion of basic design considerations, as pointed out in chapter 8 of this handbook, and which are based on experiences made in realized solar cooling installations, led to the following basic design guidelines. Compliance of these guidelines will assure a high energy saving potential: –– Most solar cooling systems will need a back-up source to provide cooling when the solar energy available is not be enough to cover the loads. In these cases either a second heat source, such as a back-up burner, to drive the thermally driven cooling equipment or a conventional vapor compression chiller should be employed. –– If a solar thermally driven cooling or air-conditioning system uses a fossil-fuelled heat source as back-up, a high solar fraction of the driving heat of the thermally driven cooling device is necessary in order to assure significant primary energy savings. This requires an appropriate design of the solar system, i.e., it is necessary to correctly dimension the solar collector field and system-integrated energy storage. Typically, dynamic system simulation models have to be employed in order to allow for a reliable forecast of the solar fraction. –– The above statement is more important and critical for systems operating with a low thermal COP value, such as single-effect systems than for systems operating with high

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thermal COPs such as double- or triple-effect systems since in the first case the heat originating from fossil fuel combustion is converted to cooling at a lower efficiency. The installation of a conventional vapour compression chiller as a second source to provide cooling is preferable from an energy saving point of view. Such a solution would be particularly appropriate for large installations where a large overall amount of cooling power is needed. In these cases even a small fraction of solar cooling will lead to primary energy savings and the solar system will mainly serve to reduce electric energy consumption as well as peak electricity loads. Solar-thermally autonomous systems that do not use any back-up for cooling may also be installed. In these cases, energy savings are always achieved. But there is no guarantee of meeting the cooling loads and maintaining the indoor climate within comfort conditions. Nevertheless, such an approach may be applicable, in particular, when the load side provides a large thermal inertia. Examples are a solar cooling system connected to a building with thermal activation of the building mass such as thermally activated concrete slabs or solar cooling of storage rooms such as in the food and agro industry sector (e.g. wine). In any case, the use of the solar collector field must be maximized through the exploitation of the solar heat source to match other loads such as space heating or domestic hot water production. Particularly in climates with high cooling loads during summer and heating loads during winter, the solar system can also contribute significantly to the meeting of heating loads, especially in regions with high solar gains throughout the year, as often occurs in the Mediterranean.

In addition some basic design and operation principles which turned out to be very useful should be considered in order to realize potential energy savings in practical operation: –– An important design item is to keep systems as simple as possible in order to reduce risks of errors during installation, operation and maintenance. –– Generally a careful design and planning is needed in order to define an optimal sizing of the key components and an appropriate design that adapts well to the actual structure load. Component oversizing should be avoided for economic reasons. Consequently, the system should not be designed to cover peak loads, that only occur a few hours of the year. In addition, intrinsic safety measures (drainback strategy for example) against solar collector overheating can be adopted in some systems where power failure can often occur or when the cooling load significantly fluctuates from one day to the other –– All auxiliary components have to be highly energy-efficient. –– An operation strategy has to be developed that leads to the most energy-efficient operation under full and part load conditions and gives the best reliability for longterm operation. –– A careful commissioning phase of the system is necessary to ensure system operation as planned. –– An ongoing monitoring (“continuous commissioning”) is helpful in order to enable long term operation with the highest possible performance.

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13.4 Economics

13.4 Economics The initial cost (investment cost including planning, assembly, construction and commissioning) for solar cooling systems is significantly higher than the corresponding cost for best practice standard solutions – this is a well-known fact for almost all solar energy systems and many other systems using renewable energies. The finitial cost of realized solar cooling installations is typically between 2 to 5 times higher compared to a conventional state-of-the-art system, depending on local conditions, building requirements, system size, and of course on the selected technical solution. In recent studies, the initial cost for overall systems (in Europe) ranged from below 2000 € per kWcold to 5000 € per kWcold (kWcold = 0.28 ton) and was even higher in some particular cases. This large range is due to the different sizes of the systems, different technologies, different application sectors, and other boundary conditions. The cost saving during operation very much depends on the boundary conditions. Boundary conditions in favor of a short payback time are: –– High solar radiation lead to high gains of the solar system. –– A long cooling season leads to a large number of hours where the system is used. –– Other heat loads such as sanitary hot water and/or building heating increase the usefulness of the solar system. –– High prices of conventional energy make a solar alternative more competitive. Looking at the overall life cycle cost of a SAC system (excluding any incentives or funds) in comparison to a conventional standard solution the situation looks much better than in the case of the initial cost. Depending on the particular conditions, SAC systems will, in many cases, amortize within their lifetime. Under promising conditions even payback times of ten years and less can be obtained. Increasing prices for conventional energy will of course contribute to an increased economic viability of solar cooling. Moreover the operation cost for a solar driven system will be much less dependent on price fluctuations of conventional energy and can be predicted much more reliably at the design stage than for conventional systems completely based on fossil energy and/or electricity from the grid. The application of solar driven cooling technology has some other advantages that are often difficult to translate into an economic figure, but may be important aspects during the decision making process: –– Solar cooling applications lead to (primary) energy savings and thus help to reduce the dependence of finite energy fuels, which have to be imported in many countries. –– Correspondingly, the application of solar air-conditioning or refrigeration systems will lead to reduced CO2 emissions and thereby contribute to a reduction in climate change and related effects. –– Systems using thermally driven cooling cycles show additional environmental benefits since they typically employ refrigerants with no ozone depletion potential and no or a very small global warming potential.

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Solar cooling systems can also be used for nearly all heating applications in a building or industry. The large solar collector field also provides heat for purposes other than cooling and thus helps to avoid fuel or electricity consumption for heating applications. These systems can contribute to grid stability in regions where a considerable share of the daily electricity consumption coming from the grid is used for air-conditioning by conventional techniques.

13.5 Outlook In summary, solar cooling using solar thermal collectors and thermally driven cooling equipment is an interesting option under many boundary conditions and should always be considered as an alternative in the planning phase of a building project where active cooling is needed. Overall, the most favorable conditions for a successful market implementation of SAC systems are: –– Applications with a high need for heating and cooling (and sanitary hot water). –– Places with a high solar energy potential i.e. high solar radiation. –– Conditions characterized by a good correlation between loads and solar gains since this reduces the storage need. –– Places with a high cost for conventional energy. –– Places with good infrastructure and qualified services for technical support in heating and cooling systems. A major obstacle is that in many places that fulfill the above requirements, no or very little experience with solar energy use exists and often HVAC and refrigeration installations are of a comparatively low standard. Companies that offer overall solutions and have the capability of providing maintenance services (e.g., using remote monitoring and control approaches) will be able to exploit this market opportunity leading possibly to a cooling energy contracting service and to guarantee of minimum solar cooling results. A major motivation for the installation of a solar air-conditioning or refrigeration system is its environmental soundness due to the reduced consumption of conventional energy and the employment of environmentally friendly refrigerants. This can be an important insentive for an investor even if the system is not considered economically competitive yet, that is the return on the investment is below the normal expectations for commercial investments. Moreover, in a background where, in the legislation, the use of renewable energies begins to be coupled to the production of cold services (air-conditioning and refrigeration) see EU directive 28/2010; solar cooling remains one of the few viable concepts for renewable cooling.

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13.5 Outlook

Concretely speaking, the following market opportunities are viable: –– Tertiary buildings, such as office buildings and hotels, in regions with sufficient solar gains: here in particular technologies employing non-tracking solar collector technologies will be employed and the solar system will be used for heating, cooling and sanitary hot water (if needed). –– Production buildings in sunny regions that need cooling for industrial processes (e.g., in the food industry): depending on the required temperature level of the cooling process either non-tracking solar collectors or single-axis tracking solar collector technologies will be employed. Large factory roofs can serve for the placement of the collectors but also ground installed systems are usual, in particular in areas where ground is not a limiting resource. –– In sunny regions in particular, a potential for the application of SAC in residential houses exists. Heating and cooling solutions making use of solar thermal energy for single family houses as well as for multi-family houses are particularly interesting in new buildings and in combination with highly energy efficient buildings which allow for radiative solutions (e.g., using the floor and / or the ceiling for heating/cooling). Overall, renewable energies will play an increasingly important role in future energy systems due to the strong need to limit CO2 emissions originating from conventional energy sources. SAC technology is one of the important solutions applicable on the demand side. This technology provides a market opportunity for many involved stakeholders including building owners, planners, manufacturers, and installation companies. Planners, engineers and end-users will continue to use solar thermal energy for air-conditioning, even if a greater effort is still necessary today during planning and sometimes also during operation. However, the experience gained in future installations will continue to contribute to an ongoing learning process and help to make this technology increasingly successful. Cost reductions due to solutions with a higher level of standardization and shortening of the design process will also be achieved. System operation will become more reliable and the effort for maintenance will be reduced by continued equipment development. Overall, solar cooling is a highly exciting and challenging goal, we use energy from the sun to meet our cooling and air-conditioning demand, a demand which has been created, in great part, by the sun.

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Chapter 14

Appendix 14.1 The IEA Solar Heating & Cooling Programme The International Energy Agency (IEA) was founded in 1974 as an autonomous body within the Organisation for Economic Co-operation and Development (OECD). It carries out a co-operative programme on issues concerning energy, including joint research and development of new and improved energy technology. The Solar Heating and Cooling (SHC) Programme was one of the first IEA research agreements to be established. Since 1976, its members have been collaborating to develop technology that uses the energy of the sun to heat, cool, light and power buildings. The following countries, as well as the European Commission, are members of this agreement: Australia, Austria, Belgium, Canada, China, Denmark, ECREEE, France, Germany, Italy, Mexico, the Netherlands, Norway, Portugal, Singapore, South Africa, Spain, Sweden, Switzerland and the United States. The mission of the SHC Programme is to facilitate an environmentally sustainable future through the greater use of solar design and technology. Current Tasks of the IEA Solar Heating and Cooling Programme are: Task 39 – Polymeric Materials for Solar Thermal Applications Task 40 – Net Zero Energy Solar Buildings Task 41 – Solar Energy and Architecture Task 42 – Compact Thermal Energy Storage Task 43 – Solar Rating & Certification Procedures Task 44 – Solar and Heat Pump Systems Task 45 – Large Scale Solar Heating and Cooling Systems Task 46 – Solar Resource Assessment and Forecasting Task 47 – Solar Renovation of Non-Residential Buildings Task 48 – Quality Assurance and Support Measures for Solar Cooling Systems Task 49 – Solar Heat Integration in Industrial Processes Task 50 – Advanced Lighting Solutions for Retrofitting Buildings Task 51 – Solar Energy and Urban Planning For more information: www.iea-shc.org

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14.2 TASK 38 Solar Air-Conditioning and Refrigeration Duration: September 2006 to 31 December, 2011

14.2.1 Objectives The main objective of the Task was the implementation of measures for an accelerated market introduction of solar air conditioning and refrigeration with focus on improved components and system concepts. Main activities to achieve the objectives were –– activities in development and testing of cooling equipment for the residential and small commercial sector; –– development of pre-engineered system concepts for small and medium size systems and development of optimized and standardized schemes for custom made systems; –– reports on the experiences with new pilot and demonstration plants and on the evaluation and performance assessment procedure; –– provision of accompanying documents supporting the planning, installation and commissioning of solar cooling plants; –– analysis of novel concepts and technologies with special emphasis on thermodynamic principles and a bibliographic review; –– performance comparison of available simulation tools and applicability for planning and system analysis; and –– market transfer and market stimulation activities, which include information letters, work-shops and training material as well as this 3rd edition of the Handbook for Solar Cooling for Planners. –– For more information about Task 38: http://www.iea-shc.org/task38/

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14.3 TASK 38 management structure

14.3 TASK 38 management structure 14.3.1 Operating Agent Hans-Martin Henning Fraunhofer-Institute for Solar Energy Systems ISE Heidenhofstr. 2 79110 Freiburg, GERMANY [email protected]

14.3.2 Subtask Leaders Subtask A – Pre-engineered systems for residential and small commercial applications Dagmar Jähnig AEE INTEC, AEE – Institute for Sustainable Technologies Feldgasse 19 8200 Gleisdorf, AUSTRIA [email protected] Subtask B – Custom-made systems for large non-residential buildings and industrial applications Wolfram Sparber EURAC research Viale Druso/Drususallee 1 39100 Bolzano/Bozen, ITALY [email protected] Subtask C – Modelling and fundamental analysis Etienne Wurtz Institut National d’Energie Solaire 50, avenue du lac Léman BP 332 73375 Le Bourget du Lac Cedex, FRANCE [email protected] Subtask D – Market transfer activities Mario Motta Department of Energy, Politecnico di Milano Piazza Leonardo da Vinci 32 20133 Milano, ITALY [email protected]

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14.4 Institutions participating in Task 38 Australia Division of Information Technology, Engineering and the Environment and Institute for Sustainable Systems and Technology University of South Australia The Mawson Centre 5095 Mawson Lakes, South Australia AUSTRALIA CSIRO Division of Energy Technology PO Box 330 2300 Newcastle, NSW AUSTRALIA

Austria ASIC – Austria Solar Innovation Center Durisolstraße 7/Top 50 4600 Wels AUSTRIA AEE INTEC, AEE – Institute for Sustainable Technologies Feldgasse 19 8200 Gleisdorf AUSTRIA Institute of Thermal Engineering Graz University of Technology Inffeldgasse 25 8010 Graz AUSTRIA AIT Austrian Institute of Technology Österreichisches Forschungs- und Prüfzentrum Arsenal Ges.m.b.H. Giefinggasse 2 1210 Vienna AUSTRIA

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14.4 Institutions participating in Task 38

AB Energieeffizientes Bauen Institut für Konstruktion und Materialwissenschaften Universität Innsbruck Techniker Str. 13 6020 Innsbruck AUSTRIA S.O.L.I.D. Gesellschaft für Solarinstallation und Design m.b.H Puchstraße 85 8020 Graz AUSTRIA

Canada Queens University – Department of Mechanical and Material Engineering KingstonON CANADA Queens University – Department of Mechanical and Material Engineering KingstonON CANADA

Denmark Ellehauge & Kildemoes Vestergade 48H, 2.tv. 8000 Århus C DENMARK AC-Sun Aps Vitus Bering Innovation Park Chr. M. Østergårdsvej 4 8700 Horsens DENMARK PlanEnergi Jyllandsgade 1 9520 Skørpingi DENMARK

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Refrigeration and Heat Pump Technology Danish Technological Institute Teknologiparken Kongsvang Allé 29 8000 Aarhus C DENMARK

France INES – Université de Savoie BP 332 – Savoie Technolac 50 Avenue du Lac Léman 73377 Le Bourget du Lac FRANCE TECSOL SA. 105 av Alfred Kastler – BP 90434 66 004 PERPIGNAN Cedex FRANCE Laboratoire d’Étude des Phénomènes de Transfert et de l’Instantanéité: Agro-industrie et Bâtiment – LEPTIAB Avenue Michel Crépeau 17042 La Rochelle Cedex 1 FRANCE CNRS-LIMSI, BP 133, Rue J. von Neumann, 91403 Orsay Cedex FRANCE

Germany zafh.net Zentrum für angewandte Forschung Nachhaltige Energietechnik Hochschule für Technik Stuttgart Kienestraße 41 70174 Stuttgart GERMANY

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14.4 Institutions participating in Task 38

Institut für Energietechnik Technische Universität Berlin Marchstraße 18 10587 Berlin GERMANY ILK Dresden GmbH Bertolt-Brecht-Allee 20 01309 Dresden GERMANY Department of Solar and Systems Engineering, Kassel University, Kurt-Wolters-Str. 3 34125 Kassel GERMANY Fraunhofer Institut für Solare Energiesysteme ISE Heidenhofstr.2 79110 Freiburg GERMANY SolarNext AG Theodor-Sanne-Str. 6 83233 Bernau am Chiemsee Germany Fraunhofer-Institut für Umwelt-, Sicherheits- und Energietechnik – Umsicht Osterfelder Str. 3 46047 Oberhausen GERMANY ZAE Bayern Walther-Meißner-Str. 6 85748 Garching GERMANY Institut für Thermodynamik und Wärmetechnik (ITW) Universität Stuttgart  Pfaffenwaldring 6 70550 Stuttgart GERMANY

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Beuth Hochschule Luxemburger Straße 10 13353 Berlin GERMANY Ingolstadt University of Applied Sciences Esplanade 10 85049 Ingolstadt GERMANY

Greece GRoup Energy Conservation (GREC) Institute for Environmental Research & Sustainable Development, National Observatory of Athens (NOA) I. Metaxa & Vas. Pavlou, Lofos Koufou, 15236, Palaia Penteli, Athens GREECE

Italy Department of Energy, Politecnico di Milano Piazza Leonardo da Vinci 32, 20133 Milano ITALY Department of Energy, Università degli Studi di Palermo Viale delle Scienze bldg 9 90128 Palermo ITALY EURAC Research European Academy Bolzano Viale Druso/Drususallee 1 39100 Bolzano/Bozen ITALY

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14.4 Institutions participating in Task 38

CREAR – Centro interdipartimentale di Ricerca per le Energie Alternative e Rinnovabili Universita degli studi di Firenze c/o Dip.to di Energetica „S.Stecco“ Via di Santa Marta, 3 50139 Firenze ITALY OLYMP ITALIA SRL Via Orbassano, 16 10090 Bruino ITALY Dipartimento di Ingegneria Industriale e Meccanica (DIIM) Universita di Catania Viale A. Doria 6 95125 Catania ITALY

Malta Institute for Sustainable Energy University of Malta, Triq il-Barrakki Marsaxlokk MXK 1531 Malta

Mexico CIE-UNAM (Centro de Investigacion en Energia), Universidad Nacional Autonoma de Mexico Privada Xochicalco S/N Temixco, Morelos 62580, México

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Portugal DER/INETI, Edificio G Estrada do Paço do Lumiar 1649-038 Lisboa PORTUGAL

Spain Energy Division, Renewable Energies Area Fundación CARTIF Parque Tecnológico de Boecillo, parcela 205 47151 Boecillo, Valladolid SPAIN IKERLAN – Centro de investigation tecnològicas Juan de la Cierva 1 01510 Miñano SPAIN Sistemes Avançats d’Energia Solar Tèrmica SCCL AIGUASOL ENGINYERIA C/ Roger de Llúria, 29 3r 2a 08009 Barcelona SPAIN INTA Ctra. San Juan del Puerto-Matalascañas, km. 34 21130 Mazagón, Huelva SPAIN Instituto de Ciencias de la Construcción Eduardo Torroja (CSIC) c/ Serrano Galvache N° 4 28033 Madrid SPAIN mira energia Tortellà, 6 2–2 08005 Barcelona SPAIN

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Sistemes Avançats d’Energia Solar Tèrmica SCCL AIGUASOL ENGINYERIA C/ Roger de Llúria, 29 3r 2a 08009 Barcelona SPAIN Unidad de Eficiencia Energética en la Edificatión CIEMAT Avenida Complutense 22 28040 Madrid SPAIN Centro national de energías renovables – CENER Avda. Ciudad de la Innovacion n 7 31621 Sarriguren, Navarra SPAIN

Switzerland Laboratory of Solar Energetics and Building Physics (LESBAT) HEIG-VD – School of Business and Engineering Route de Cheseaux 1 1400 Yverdon-les-Bains SWITZERLAND Hochschule für Technik Rapperswil HSR Institut für Solartechnik SPF Oberseestrasse 10 8640 Rapperswil SWITZERLAND Hindenburg Consulting – Solar.Cooling.Comfort Hauptstrasse 28 4469 Anwil Switzerland

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USA National Solar Thermal Test Facility Sandia National Laboratories MS 1127 Albuquerque, NM 87185 USA

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