6E9?07/7y

INSTITUTIONEN FOR VARME- OCH KRAFTTEKNIK KRAFTVERKSTEKNIK LUNDS TEKNISKA HOGSKOLA

Cogeneration Based on Gasified Biomass - a Comparisonof Concepts

by

Fredrik Olsson

DISTRIBUTION OF THIS DOCUMENT IS UNLIMITED FOREIGN SALES PROHIBITED 0jL^

Thesis for degree of Licentiate of Engineering

ISRN LUTMDN/TMVK—7033--SE

DIVISION OF THERMAL POWER ENGINEERING DEPARTMENT OF HEAT AND POWER ENGINEERING LUND INSTITUTE OF TECHNOLOGY P.O. BOX 118, S-221 00 LUND 1999 SWEDEN DISCLAIMER

Portions of this document may be illegible in electronic image products. Images are produced from the best available original document. ISRN LUTMDN/TMVK—7033 —SE

CogenerationBased on Gasified Biomass - a Comparison of Concepts

by

Fredrik Olsson

Thesis for degree of Licentiate of Engineering

Lund Institute of Technology Division of Thermal Power Engineering - VOK

January 1999 Dokumentotgivare Dokumenmamn Dokumentbeteckning LU/LTH Licentiatrapport ISRN LUTMDN/IMVK—7033—SE Handlaggare Utgivningsdatom Arendebeteckning Januari 1999 Forfattare Fredrik Olsson

Dokumenttitel och undertitel Cogeneration Based on Gasified Biomass - a Comparison of Concepts

Referat (sammandrag) In this report, integration of drying and gasification of biomass into cogeneration power plants, comprising gas turbines, is investigated. The thermodynamic cycles considered are the combined cycle and the humid air turbine cycle. These are combined with either pressurised or near atmospheric gasification, and steam or exhaust gas dryer, in a number of combinations. An effort is made to facilitate a comparison of the different concepts by using, and presenting, similar assumptions and input data for all studied systems. The resulting systems are modelled using the software package ASPEN PLUS™, and for each system both the electrical efficiency and the fuel utilisation are calculated. The investigation of integrated gasification combined cycles (IGCC), reveals that systems with pressurised gasification have a potential for electrical efficiencies approaching 45% (LHV). That is 4 - 5 percentage points higher than the corresponding systems with near atmospheric gasification. The type of dryer in the system mainly influences the fuel utilisation, with an advantage of approximately 8 percentage points (LHV) for the steam dryer. The resulting values of fuel utilisation for the IGCC systems are in the range of 78 - 94% (LHV). The results for the integrated gasification humid air turbine systems (IGHAT) indicate that electrical efficiencies close to the IGCC are achievable, provided combustion of the fuel gas in highly humidified air is feasible. Reaching a high fuel utilisation is more difficult for this concept, unless the temperature levels in the district heating network are low. For comparison a conventional cogeneration plant, based on a CFB boiler and a steam turbine (Rankine cycle), is also modelled in ASPEN PLUS™. The IGCC and IGHAT show electrical efficiencies in the range of 37 - 45% (LHV), compared with a calculated value of 31% (LHV) for the Rankine cycle cogeneration plant. Apart from the electrical efficiency, also a high value of fuel utilisation is important in cogeneration. When flue gas condensation is utilised, the conventional plant can achieve values of fuel utilisation exceeding 100% (LHV) and in this respect the new technologies are inferior. In addition to the thermodynamic calculations, some critical components and development needs are reviewed. One area still requiring development is the gas cleaning, critical issues being, for example, hot gas filtration, tar cracking and ammonia removal.

Referat skrivet av Forfattaren Forslag till ytterligare nyckelord

Klassifikationssystem och -klass(er)

Indextermer (ange kalla)

Omfing Ovriga bibliografiska uppgifter 129 sid SprSk Engelska Sekretessuppgifter ISSN ISBN 0282-1990 Dokumentet kan erh Silas frSn Mottagarens uppgifter Institutionen for Varme- och Kraftteknik Box 118 221 00 Lund Pris Summary

The risk of a global climate change, due to increasing concentrations of green house gases in the atmosphere, presently increases the interest in biomass as an energy source. Hence, a growing market for conversion technologies using biomass to produce heat, electricity and/or liquid fuels is expected.

In Sweden, biomass is used extensively for heat production, e.g. in district heating networks, and also cogeneration of heat and electricity is attracting interest. The prevailing technology, solid fuel boiler and steam turbine, is characterised by high fuel utilisation and low power to heat ratio. In order to increase this ratio, other power plant concepts incorporating biomass gasification and gas turbines, could be considered.

In this report, integration of drying and gasification of biomass into power plants comprising gas turbines is investigated. The thermodynamic cycles considered are the combined cycle and the humid air turbine cycle. These are combined with either pressurised or near atmospheric gasification, and steam or exhaust gas dryer, in a number of combinations. The resulting systems are modelled using the software package ASPEN PLUS™, and for each system both the electrical efficiency and the fuel utilisation are calculated.

In this context, an effort is made to facilitate a comparison of the different concepts by using, and presenting, similar assumptions and input data for all studied systems. Moreover, the complexity of the different systems are kept at similar levels, to assure that no concept is unduly favoured. Judging if two power plant concepts are equally complex is, however, difficult. Here the choice of plant layouts is mainly based on common sense and some knowledge of present-day power plant technology.

The investigation of integrated gasification combined cycles (IGCC), reveals that systems with pressurised gasification have a potential for electrical efficiencies approaching 45% (LHV). That is 4 - 5 percentage points higher than the corresponding systems with near atmospheric gasification. The type of dryer in the system mainly influences the fuel utilisation, with an advantage of approximately 8 percentage points (LHV) for the steam dryer. Similarly, the systems with near atmospheric gasification show higher heat losses and, consequently, lower fuel utilisation, than the pressurised systems do. Hence, the system with pressurised gasification and steam dryer has a fuel utilisation of approximately 94% (LHV). For the system with near atmospheric gasification and exhaust gas dryer, the corresponding figure is only 78%.

The influence of some operational parameters and process alterations are also studied, and a possible further improvement of 1 - 3 percentage points in electrical efficiency is found.

In addition to the thermodynamic calculations, some critical components and development needs are reviewed. One area still requiring development is the gas cleaning, critical issues being, for example, hot gas filtration, tar cracking and ammonia removal.

Estimates of the investment cost for these technologies are found in the literature. Using these, together with the calculated values of electrical efficiency and fuel utilisation, the cost of electricity from IGCC cogeneration plants is estimated to 0.37 - 0.42 SEK/kWh e. These

i figures vary considerably with the fuel price, the annual operating time and the scale of the plant, but generally the IGCC cannot be considered competitive in Sweden today.

The results for the integrated gasification humid air turbine systems (IGHAT) indicate that electrical efficiencies close to the IGCC are achievable, provided combustion of the fuel gas in highly humidified air is feasible. Reaching a high fuel utilisation is more difficult for this concept, unless the temperature levels in the district heating network are low.

For comparison a conventional cogeneration plant, based on a CFB boiler and a steam turbine (Rankine cycle), is also modelled in ASPEN PLUS™. The IGCC and IGHAT show electrical efficiencies in the range of 37 - 45% (LHV), compared with a calculated value of 31% (LHV) for the Rankine cycle cogeneration plant. Apart from the electrical efficiency, also a high value of fuel utilisation is important in cogeneration. When flue gas condensation is utilised, the conventional plant can achieve values of fuel utilisation over 100% (LHV) and in this respect the new technologies are inferior.

n Summary i

Nomenclature 5

1 Introduction 6

1.1 Background...... 6 1.2 Objectives ...... 7 1.3 Method...... 8 1.4 Delimitations...... 9 1.5 Outline of the thesis...... 9

2 Biomass 10

2.1 Overview...... 10 2.2 Properties...... 10 2.3 Model fuel...... 12

3 Integrated gasification combined cycle 13

3.1 Overview...... 13 3.2 Models...... 15 3.2.1 Dryer...... 15 3.2.2 Gasifier...... 22 3.2.3 Gas cleaning...... 28 3.2.4 Gas turbine ...... 32 3.2.5 Boost compressor...... 38 3.2.6 Fuel compressor...... 39 3.2.7 Heat recovery steam generator...... 39 3.2.8 Steam turbine ...... 41 3.3 Parametric study...... 43 3.3.1 Pressurised gasification and exhaust gas dryer...... 45 3.3.2 Pressurised gasification and steam dryer...... 48 3.3.3 Near atmospheric gasification and exhaust gas dryer...... 52 3.3.4 Near atmospheric gasification and steam dryer...... 55 3.3.5 Ambient temperature design value...... 58 3.3.6 Fuel moisture content...... 60 3.3.7 Heat loss between dryer and gasifier...... 64 3.3.8 Temperature and pressure of fuel gas...... 65 3.3.9 Pressure in steam dryer...... 67 3.3.10 Temperature differences in heat recovery steam generator..... 69 3.3.11 Superheating in raw gas cooler...... 70 3.3.12 Uncooled gas turbine ...... 72 3.3.13 Advanced gas turbine ...... 74 3.4 Results and general discussion...... 76 3.4.1 Comparison of IGCC systems...... 76 3.4.2 Economy of IGCC...... 80 3.4.3 Plant availability...... 84 3.4.4 Development needs ...... 86

3 4 Integrated gasification humid air turbine 88

4.1 Overview...... 88 4.2 Models...... 90 4.2.1 Humidifier ...... 90 4.3 Parametric study...... 92 4.3.1 Pressurised gasification and exhaust gas dryer...... 94 4.3.2 Pressurised gasification and steam dryer...... 96 4.3.3 Near atmospheric gasification and exhaust gas dryer...... 98 4.3.4 Near atmospheric gasification and steam dryer...... 100 4.4 Results and general discussion...... 102 4.4.1 Comparison of IGHAT systems...... 102 4.4.2 Development needs ...... 103

5 Rankine cycle 105

5.1 Overview...... 105 5.2 Models...... 105 5.3 Parametric study...... 106 5.4 Results and discussion...... 108 5.4.1 Results of parametric study...... 108 5.4.2 Development needs ...... 109

6 Comparison of presented technologies 111

6.1 Thermodynamic comparison...... 112 6.2 Technological comparison...... 114 6.3 Concluding remarks...... 115

7 Acknowledgements 117

References 118

Appendices

I Calculation sequence and process flow diagram

II Cost of heat from boiler plant m T,Q-diagram for IGHAT systems with different district heating supply temperatures.

4 Nomenclature

A area cp constant pressure heat capacity cv constant volume heat capacity d diameter G Gibbs' free energy H enthalpy h specific enthalpy M molecular weight m mass flow n polytropic coefficient p pressure Q heat flow S entropy T temperature v specific volume w weight fraction W moisture content on a wet basis X vapour fraction a heat transfer coefficient y ratio of heat capacities Ap pressure drop ec cooling effectiveness T|e electric efficiency of motor/generator T)m mechanical efficiency T|p polytropic efficiency T|s isentropic efficiency X heat of vaporisation

Subscripts a air bl blade des design value e electricity f fuel st stage th thermal w water

5 1 Introduction

1.1 Background

Today most of the global power production is based on combustion of fossil fuels, i.e. coal, oil and natural gas. This is, however, not a sustainable solution in the long term, since the supplies of these fuels are finite. For coal, the most abundant of the three, the known sources economically available with present-day technology are estimated to last for about 228 years at the present rate of consumption. The corresponding figure for oil is 43 years [NUTEK, 1997].

There are also more acute problems regarding these fuels, for example, the risk of a global climate change due to increasing concentrations of greenhouse gases in the atmosphere. The major concern here is carbon dioxide (C02). This has caused an interest in biomass as an energy source, since it is considered neutral with respect to the C02-concentration in the atmosphere. The reason for this is that in a system where biomass is grown in a sustainable way, the C02 released from combustion of biomass is absorbed again by the next generation of growing plants.

The limited reserves of fossil fuels and the concern for the greenhouse effect are two reasons why the world-wide use of biomass is projected to increase substantially in the near to mid­ future [Faaij, 1997; Johansson et al., 1996]. Consequently, there will also be a growing market for conversion technologies using biomass to produce heat, electricity and/or liquid fuels.

The prevailing method for generating electricity from biomass is combustion of solid wood to produce steam for a steam turbine in a Rankine cycle. Due to material considerations, limiting the steam data, the resulting electrical efficiency is normally rather low. To make better use of the chemical energy in the fuel, cogeneration of heat and power is an attractive possibility.

In most cogeneration schemes there is a certain heat demand, e.g. a district heating network, determining the size of the power plant. As a practical consequence of this there is often an upper limit to the size of the plant. Also the fact that biomass is a dispersed resource, requiring collection and transportation, imposes an upper limit on the size. Depending on location this upper limit can be in the range of 10 - 100 MWe [Bridgwater, 1995]. The lower limit for the plant output is normally set by increasing specific investment costs.

In Sweden, the major part (>90%) of the electricity is produced by hydro power and nuclear power. There is, however, a political uncertainty concerning the future for nuclear power, indicating a possible need for new generating capacity in the mid-future. Moreover, the possible increase in hydro power is presently limited, due to laws restricting further exploitation of larger rivers. Hence, new capacity will probably be based on thermal power plants utilising various fuels.

Consequently, the interest in power production based on biomass is increasing also in Sweden. In this connection, a relatively high domestic biomass potential, a tradition of using biomass as a fuel, e.g. in pulp and paper, and sawmill industries, and substantial potentials for cogeneration both in district heating plants and in industry, are advantageous.

6 If the main purpose is to produce the greatest possible amount of electricity in cogeneration (where the heaf demand is fixed), the ratio of power to heat should be as high as possible. At the same time, a high total fuel utilisation1 should be maintained. In steam power plants burning solid wood, the fuel utilisation is normally around 90% and the power to heat ratio is typically 0.5 [Stahl, 1997]. If flue gas condensation is applied the fuel utilisation can be over 100%, based on the lower heating value of the fuel.

To increase the power to heat ratio, thermodynamic cycles other than the Rankine cycle could be considered. One possibility could be the combined cycle, which connects the Rankine cycle to a gas turbine (Brayton) cycle. Also other cycles involving gas turbines could be considered, with the aim to take advantage of the rapid development of these machines regarding efficiency and environmental performance.

A drawback is the difficulty in using solid fuels like biomass in a gas turbine, which traditionally requires high-quality gaseous or liquid fuels. Possible ways to handle this problem are closed loop gas turbine cycles with external combustion of biomass, direct combustion of pulverised wood in the gas turbine combustor, or gasification of biomass and subsequent combustion of the produced gas in the gas turbine combustor.

The concept of integrated gasification has been subjected to numerous studies, and two main paths seem to have developed, performing the gasification either at elevated pressure or at a pressure close to atmospheric. Furthermore, the use of either steam or exhaust gas for drying of the wet biomass before gasification has been proposed. There is, however, still a need for a systematic comparison among plants which differ in these respects.

1.2 Objectives

The main purpose of this work is to evaluate alternative concepts for combined production of heat and power from gasified biomass. The thermodynamic cycles considered are the combined cycle and the humid air turbine cycle. These are combined with either pressurised or near atmospheric gasification, and steam or exhaust gas dryer, in a number of combinations. The evaluation is primarily thermodynamic, considering electrical efficiency and fuel utilisation. Consequently, the economics of the studied concepts will not be included (with one exception). Nevertheless, it is believed that the results, in terms of thermodynamic performance of well defined plant concepts, can be used as input data to later economic analyses.

Systems similar to the ones considered here have been studied previously by for example Barbucci et al. (1994), Consonni and Larson (1996b), Faaij (1997) and Steinwall (1995). Often it is difficult to compare results from different investigations, because input parameters and assumptions are different or not clearly presented. Furthermore, the complexity of the various systems, e.g. the number of pressure levels, flow splits and heat exchangers used in the heat recovery system, often differs substantially between different investigations. One objective here is to facilitate the comparison of different concepts by using, and presenting, similar assumptions and input data for all studied systems, and by trying to calculate systems of comparable complexity.

1 Fuel utilisation = (heat delivered to process or district heating + electricity)/(fuel mass flow * heating value)

7 As a result, the-difference in thermodynamic performance among the studied concepts will be quantified. In addition to this, a plant based on conventional technology for biomass based cogeneration, comprising a solid fuel boiler and a steam turbine, is calculated and presented as a reference case.

Furthermore, variations in electrical efficiency and fuel utilisation, resulting from some parameter variations and process alteration, are calculated. The objective of this is both to reveal how sensitive the calculation results are to the assumed input data, and to determine if various concepts are equally sensitive to changes in input data. The focus of this analysis is on the integrated gasification combined cycle and, to a lesser extent, on the integrated gasification humid air turbine cycle.

In order to obtain the thermodynamic performances predicted in this report, for the various technologies, some development work still remains. Also, to ascertain satisfactory environmental performance and availability, some pieces of equipment require further development. It is the authors intention to review some of the development needs of the new technologies utilising biomass gasification. Again, the largest effort is put into describing the requirements of the integrated gasification combined cycle technology.

1.3 Method

This report is the result of a combination of literature studies and computer-aided thermo­ dynamic calculations. Literature studies have given knowledge of the state of the art of conversion technologies for production of heat and power from gasified biomass. This is presented both in the chapters on drying, gasification and gas cleaning, and in the introductory overviews of entire power plant systems. Also, the sections on availability and development needs, are mainly based on literature studies. Moreover, the literature study has given important information on input parameters for the thermodynamic calculations.

The calculations have been performed by means of the software package ASPEN PLUS™. A number of plant layouts have been defined and modelled using existing component models, and the steady state performances of these systems have been calculated. A number of parameters have been varied in order to investigate their influence on plant performance.

In order to make the comparison among different concepts as reliable as possible, an effort has been made to keep the input data equal for all systems whenever possible. The input parameters used are also clearly presented so that every reader can make his/her own judgement about their influence on the result.

The question of comparable system complexities is more difficult, since there is no clear-cut definition of system complexity. Possible measures of complexity could be, for example, the total cost of equipment, the difficulty to control the process, or the number, and nature, of pieces of equipment used in the system. The approach taken here is based on the number and nature of the equipment used. In practice, this means that the same layout is used, whenever possible, for those subsystems which are common to all concepts studied. Other subsystems, such as dryers and gasifiers, are not the same in the various concepts. As a consequence, there are some differences in the parts of the heat recovery system which are used to integrate these

8 subsystems into the power plant. This may also influence the layout of other subsystems, and hence, the complexity of the entire plant. Moreover, three different thermodynamic cycles, all using different pieces of equipment, are studied. To compare the complexity of the resulting systems is even more complicated. Hence, the presented concepts are not completely equal in this respect, but it should be noted that an effort has been made to limit the differences.

1.4 Delimitations

Biomass considered in this work is normally referred to as woody biomass. This includes by­ products from wood product manufacturing plants, logging residues from timber harvesting operations and non-commercial products from standing forests. Other types of biomass, e.g. agricultural crops and wastes, are not included.

All calculations are performed at design point, i.e. it has been assumed that dryers, gasifiers, turbomachinery, heat exchangers and other equipment are available at any size required to match the components in a system. Also, the parameter variations are performed as design calculations, meaning that a new system is designed for every case.

Some of the components used in the calculations do not exist in that form today, e.g. a gas turbine managing the much larger mass flow through the turbine resulting from an increased flow of fuel gas. It is, however, believed that such components could be built using more or less present-day technology. This is also reflected in the choice of parameters used for these components.

Changes in component efficiencies etc. resulting from differing plant sizes are not considered in the calculations. The values used are believed to be reasonable for plants with a fuel input in the range 80- 100 MWf.

1.5 Outline of the thesis

Chapter 1 provides the background and the objectives of this thesis. In chapter 2 some properties of biomass, and the model fuel used in the calculations, are presented. Chapters 3,4 and 5 deal with the integrated gasification combined cycle, the integrated gasification humid air turbine, and the Rankine cycle, respectively. Each of these chapters provides a description of the technology and of the models used. The results of the calculations, as well as some discussion of the technologies, are then given at the end of each chapter. Also, the results of the parameter variations, performed for the integrated gasification combined cycle and humid air turbine cycle, are presented in chapters 3 and 4, respectively. In chapter 6, finally, the three technologies are compared and some differences are high-lighted.

9 2 Biomass

2.1 Potential

According to Nilsson et al. (1996), the total annual quantity of wood fuels available in the southern part of Sweden in the years 1998 - 2008 corresponds to 40 TWh/year (higher heating value, HHV). Reductions for small-scale domestic heating (-10 TWh) and ecological restrictions result in a potential around 26 TWh. These figures do not include industrial by­ products like bark and sawdust.

An earlier investigation [SOU 1992:90] indicated that the total potential for the entire country would be 48 - 59 TWh/year (lower heating value, LHV). When industrial by-products, recycled building timber and black liquor are included, the potential is estimated to be 94 - 109 TWh/year. It is important to note that this potential is based on annual growth rates and the presumed use in the pulp and paper, and sawmill industries. Note also that in this case, the consumption for domestic heating has not been subtracted.

Both investigations referenced above are based on the assumption that removal of wood fuel is compensated for by using wood ash as a fertiliser. If this is not possible, ecological considerations are likely to restrict the potential.

The figures presented above are somewhat uncertain and depend largely on the development on the pulp and paper market, as well as on the price levels in the fuel and timber markets. Nevertheless, it is obvious that a significant potential for increased fuel production exists.

2.2 Properties

The properties of biomass from woody materials vary among different species of trees. One way to characterise the material is by its chemical composition. Both ultimate and proximate analyses are used, and in tables 2.1 and 2.2, normal ranges for the compositions are given [Reed, 1981].

Oven-dry wood Oven-dry bark Volatile matter 77-87 70-74 Fixed carbon 13-21 24-28 Ash 0.1-2 1-3

Table 2.1. Range of proximate analysis data for biomass materials. Weight percent, dry basis. [Reed, 1981]

Tables 2.1 and 2.2 give the properties on a dry basis. But fresh biomass also contains water, the amount of which depends not only on the species but also on the time of the year. Normal moisture contents are in the range 40 - 60% (wet basis). The moisture content affects, among other things, combustion properties and energy required for transportation.

10 Element Ultimate analysis Another important property for a fuel is the C "49.7 - 53.5 heating value. The heating value of a H 5.8 - 6.5 carbon-containing fuel is experimentally N 0.0 - 0.3 determined by means of a bomb o S d o calorimeter. Since the enthalpy difference between reactants and products is 0 40.3-43.1 measured at 25°C, the water of combustion Ash 0.1 -2.2 is present in the liquid state, and thus the test gives the HHV. Higher heating values Table 2.2. Range of ultimate analysis data for woody biomass are often in the range for woody materials. Weight percent, dry 20 - 22 MJ/kg (dry basis). basis. [Reed, 1981]

A number of correlations exist for estimating the HHV from the ultimate analysis of a fuel. One is the IGT equation, an empirical equation based on experimental data from more than 700 coal samples. When used for fresh biomass, the average error is approximately 1.5 - 2% [Reed, 1981]. The HHV on a dry basis, given in kJ/kg, is calculated as

HHVdry =2.326- (146.58 -C+ 568.78 -if + 29.45- 6.58 • Ash -51.53 • (O + N)) (1.1)

where C, H, Ash, 0 and N are weight percentages of the elements on a dry basis.

Another, somewhat better correlation is very similar to (1.1), but based on heats of formation.1

f f 17244 -H Yl 2.326- 141 • C + 615 • H -10.2 • N + 39.95 • 5 - (l - 0.01 • Ash) ■ ------+ 149 l c J) (1.2)

The errors for this equation are said to be within the experimental errors [Reed, 1981].

This is converted to wet basis using

HHV = {\-W)-HHVdry (1.3)

There is also the lower heating value, which assumes the water of combustion to be in the gaseous state. The lower heating value on a wet basis is then given by

( mh2o LHV - HHV - W + (l-W)-wH •A; (1.4) M 'HnO V H,2 y

1 In the reference, the equation is given as r (17244-fiA HHVdry = 2.326- 141-C + 615-tf-10.2- N + 39.95-5-{\-Ash)- c r149 , i.e.

without multiplying the weight percentage of ash by 0.01 and with a different placement of the penultimate parenthesis. When checking the derivation and calculation results presented by Reed (1981), equation (1.2) is the correct form.

11 Here the heat of vaporisation of water, A#.# > both for the original moisture in the fuel and for the water resulting from oxidation of the hydrogen content, has been subtracted.

2.3 Model fuel

In tables 2.3 and 2.4, the ultimate and proximate analyses of the model fuel used in the calculations in this work are presented. When compared with typical values given above, the fuel can be characterised as a mixture of wood chips and bark, i.e. forest residues.

Element Ultimate analysis Proximate C 49.00 analysis H 6.40 Volatile matter 80 N 0.45 Fixed carbon 18 S 0.05 Ash 2 0 42.10 Ash 2.00 Table 2.4. Proximate analysis data for model fuel. Weight percent, dry basis. Table 2.3. Ultimate analysis data for model fuel. Weight percent, dry basis.

The default moisture content for the wet fuel delivered to the plant is 50% (wet basis). Using equations (1.2), (1.3) and (1.4) then gives the following heating values, as received.

HHV =10.216 MJ/kg LEV = 8.296 MJ/kg

These values are used to calculate plant efficiencies throughout this report.

12 3 Integrated Gasification Combined Cycle

3.1 Overview

As a result of several decades of development, the electrical efficiency of the gas and steam turbine combined cycle is now approaching 60% (LHV) for large plants burning natural gas [Briesch et al., 1995; DeMoss, 1996; Jeffs, 1996]. At the same time, coal is utilised in condensing steam turbine power plants at a maximum efficiency in the range of 45 - 47% (LHV) [Kjaer, 1993; Lindhardt and Lind-Hansen, 1994].

Now efforts are made to enable the use of low cost coal in high efficiency combined cycles, a combination that could decrease the emissions (primarily CO2) and the cost of electricity. Coal is, however, not a suitable fuel for modem gas turbines requiring gaseous or liquid fuels of high purity. To solve this problem, the concept of gasification of the coal and subsequent cleaning of the fuel gas has been proposed, yielding what is known as an Integrated Gasification Combined Cycle (IGCC).

The thermodynamic potential of the IGCC has been recognised in several studies, e.g. [Bohm, 1994; Weinzierl, 1994; Pmschek et al., 1998]. To show the feasibility of the technology in practice, a number of pilot and demonstration plants involving gasifiers, cleaning equipment and complete power plants have been erected. In table 3.1 some features of five major ongoing IGCC demonstration projects are presented [Bergmann et al., 1997; Holt, 1998]. 1 2

Gasification technology Electric Power Start of Total efficiency output operation hours of (LHV) (MWe) on syngas operation, April -98 PSI/DESTEC, DESTEC two-stage 40 262 1995 6800 Wabash River, entrained flow. USA Oxygen-blown. Cold gas cleaning. Tampa Electric, Texaco entrained flow. 42% 250 1996 6300 Polk-County, Oxygen-blown. USA Cold/hot1 gas cleaning. Sierra Pacific, KRW fluid bed. 42 95 1998 Pinon Pine, Air-blown. USA Hot gas cleaning. SEP/Demcolec Shell entrained flow. 43 253 1993 11000 Buggenum, NL Oxygen-blown. Cold gas cleaning. ELCOGAS Prenflo entrained flow. 45 320 1998 Puertollano, E Oxygen-blown. Cold gas cleaning.

Table 3.1. IGCC demonstration plants. [Bergmann et al., 1997; Holt, 1998]

1 Slip stream for test of hot gas cleaning planned 2 Design value, not yet achieved

13 The test runs, which for some of the plants are quite extensive, revealed some technical problems, and after solving these, the availability at the Tampa Electric and Buggenum plants has been around 75% in the last quarters. Still, the main challenges are to prove operating reliability comparable to or better than present-day solid fuel fired units and to reduce the capital cost [Holt, 1998].

Turning to biomass, there is of course the same potential for improvement, since present-day technology for power production, comprising solid fuel boilers and steam turbines, shows rather poor efficiencies. This is partly because the plants are normally rather small, well below 100 MWe, and partly because the alkali and chloride content of many biofuels limits the possible steam temperature in the boiler. Faaij (1997) states net electrical efficiencies in the range of 20 - 40% (LHV) for such plants, the higher figure obtainable in systems larger than 100 MWe or when biomass is co-fired with coal. Bain et al. (1996), as well as Consonni and Larson (1996a) give similar figures, e.g. electrical efficiencies below 25% (LHV) for 25 MWe plants.

An IGCC plant for biomass differs from the corresponding coal-based plant in some important aspects. Firstly, the gasification can be performed at a somewhat lower temperature due to the higher reactivity of biomass, and the low sulphur content normally means simpler gas cleaning equipment. On the other hand, the need for drying of the raw fuel before gasification is often significant when using biomass.

Another aspect is the size of the power plant. Since biomass is an energy source characterised by geographical dispersion, it is normally believed the power plant should be rather small, typically 10-100 MWe [Bridgwater, 1995], in order to limit the need for transportation.

As for coal, research and development have resulted in a number of test facilities and demonstration plants. Yan (1998) has made an excellent review of the state of the art of biomass gasification power generation technologies, and here only two demonstration projects will be treated briefly.

One demonstration plant was built in Vamamo, Sweden, by a joint venture company owned by Foster Wheeler International Inc. and Sydkraft AB. There, full integration of a biomass gasification plant and a combined cycle is demonstrated. The first tests with a fully integrated plant burning product gas in the gas turbine combustor were performed in October 1995 [Stahl and Neergaard, 1996]. Since then, approximately 2000 hours of integrated operation have been logged, and several tests have been performed [Sjunnesson, 1999].

The cogeneration plant has an electrical output of 6 MWe and delivers 9 MW* to the district heating network of Vamamo. This corresponds to an electrical efficiency of 32% and a total fuel utilisation of 82% (LHV, fuel with 15% moisture). These figures are rather ordinary, but the main purpose of this plant is to demonstrate the operability of a fully integrated plant, rather than to show the high efficiencies expected with this technology [Stahl, 1997].

At the power plant, the fuel, which is wood chips from forest residues, is dried in a separately fired exhaust gas dryer before it is fed to the pressurised gasifier of a circulating fluid bed type. The gas from the gasifier passes a cyclone before it is cooled to 350 - 400°C and passed through a ceramic hot gas filter. After the filter, the gas is led directly to the gas turbine combustor. The combined cycle has one pressure level in the heat recovery steam generator

14 (HRSG), and additional steam is produced in the gas cooler. The steam turbine is fed with steam of 40 bar/455°C, and the condenser is cooled by district heating water [Stahl and Neergaard, 1996].

Another power plant of the same size, 8 MWe, is now under construction in Yorkshire, England, and the start of operation is planned for 1999 [Rensfelt, 1997]. In this plant, the gasifier, from TPS Termiska Processer, is a circulating fluidised bed operating at near atmospheric pressure. This means that the produced gas must be compressed to the operating pressure of the gas turbine combustor. The gas cleaning will be performed at a lower temperature than that in Vamamo, utilising a filter and a wet scrubber for removal of particles and ammonia.

The combined cycle is similar to the one in Vamamo, with one difference being the possibility of supplementary firing in the heat recovery steam generator. Another difference is that the exhaust gas leaving the heat recovery steam generator is used to dry the fuel in an integrated dryer. The fuel consists of short rotation coppice (willow and poplar) grown on dedicated plantations (2600 hectares) in the vicinity of the plant [Beenackers and Maniatis, 1996].

3.2 Models

In this section some major IGCC components, such as, dryer, gasifier, gas cleaning equipment and turbomachinery, will be presented together with the corresponding models used in the calculations.

3.2.1 Dryer

Fresh wood fuels have moisture contents of 40 - 60% (wet basis) as stated above, depending on the species and time of the year. With such high water content, the fuel is sensitive to attacks from micro-organisms when stored. This is one reason why biomass is dried. In gasification applications, there are also other reasons for drying the fuel, e.g. improved flow characteristics (less clogging) in the fuel handling system and higher heating value of the product gas.

One way of characterising drying processes is by the type of drying medium. Dryers using air or exhaust gas are considered relatively simple and inexpensive. Their main drawbacks are high energy requirement and the fire hazard due to the combination of oxygen and elevated temperature. The fire hazard limits the operating temperature, which in turn increases the time for drying.

Steam dryers are often more complex, and the investment cost is higher. The advantages are a faster process, since higher temperatures are possible (no oxygen), and better possibilities of recovering the latent heat from the evaporated moisture. Steam drying is also considered to have better environmental performance, because pollutants in the form of particles and volatile species are trapped in a liquid phase when the steam is condensed. The need for flue gas treatment is thereby shifted to a need for water treatment of a comparably small flow [Magnusson, 1993].

15 In the exhaust gas dryers presented in table 3.2, the fuel to be dried is brought into direct contact with the exhaust gas. This can be accomplished in a rotary dryer, a pneumatic conveying dryer, or some other configuration. The exhaust gas provides the energy required to evaporate part of the moisture in the fuel, and the same gas is used to transport the moisture from the dryer.

Drying medium exhaust gas exhaust gas steam Type rotary drum pneumatic conveyer pneumatic conveyer Fuel types handled wood chips, bark bark, wood chips, peat bark, wood chips, peat Particle size coarse-medium medium-small medium-small

FG temp, in 300-400°C 150-300°C -

FG temp, out 80-120°C 80-120°C - Pressure ~1 bar ~1 bar 1.5-5 bar Electricity 50-200kJ/kg DS 200-400 kJ/kg DS1 300-400 kJ/kg DS12 consumption

Table 3.2. Characteristics of some dryers. [Wimmerstedt and Halls from, 1984; Magnusson, 1993; Ihren, 1991; Wimmerstedt and Linde, 1998]

In figure 3.1, a pneumatic conveyer dryer is shown schematically. This type of dryer is often used for drying by-products at saw mills, prior to further processing, and for drying wet fuels (e.g. bark, peat) immediately before combustion. For satisfactory operation, the dried particles must be rather small, typically ~1 mm, and hence, a mill or crusher is often included. These dryers can achieve moisture contents as low as 2 - 10%. There are, however, also other constructions capable of handling coarser material. Then somewhat higher final moisture contents (12%) are reported [Magnusson, 1993].

Exhaust gas

Wet fuel

Dried fuel Exhaust gas

Figure 3.1. Pneumatic conveyer type dryer.

1 The higher figure includes extensive milling. 2 The higher figure includes milling to particle size <10 mm.

16 In steam dryers, the energy required to evaporate part of the moisture content of the fuel is supplied in the form of steam. Figure 3.2 shows the concept of a pneumatic steam dryer. The moist fuel is supplied to the dryer where it meets superheated steam at 1.5 - 5 bar. This steam is used to convey the biomass through a number of tube bundles. High-pressure steam (8-15 bar) is condensed on the outside of these bundles, heating the biomass and the low-pressure steam. The normal temperature driving force is reported to be in the range of 30 - 60°C [Wimmerstedt, 1995; Wimmerstedt and Hallstrom, 1984].

During this process, water is evaporated from the biomass. Finally, the dried fuel is separated from the low pressure steam in a cyclone. Some of the steam is returned to the fuel intake, and the excess steam is discharged from the system for further treatment [Magnusson, 1993].

For satisfactory operation, the fuel particles must be rather small, typically <10mm, so that they can be carried through the dryer by the flow of the superheated steam1. Hence, some sort of mill or crusher is often included to reduce the size of the wood chips [Wimmerstedt and Hallstrom, 1984].

high pressure wet fuel

low pressure

♦ condensate

Figure 3.2. Pneumatic type steam dryer.

Having presented an overview of dryer types suitable for biomass drying it is now time to present the models used in the calculations. Since the scope of this report is not to perform design calculations for the dryer, the presented dryer models are simply required to fulfil the energy equation, with temperature levels similar to those presented in literature.

In this report, a default moisture content of 50% is used for the biomass supplied to the processes. When the fuel is to be gasified, the moisture content is decreased to 15% in a drying process prior to gasification. The influence of moisture content on the heat and power

1 Other steam dryers, e.g. one from NIRO A/S, accept fuel particles up to 35 mm [Barbucci et al., 1994]

17 production will also be investigated by varying these values between 30 - 60% (wet basis) for the raw biomass and 10 - 20% for the dried biomass delivered to the gasifier.

Two types of biomass dryers have been modelled in ASPEN PLUS™, one pressurised steam dryer and one atmospheric exhaust gas dryer. Both are assumed to be of the pneumatic conveyer type. The main reason for these choices is to facilitate the comparison between an exhaust gas dryer and a steam dryer. Furthermore, the relatively small particle sizes required in these dryers are well suited for subsequent gasification in a fluid bed gasifier.

In the calculations, a heat loss to the surroundings, amounting to 5% of the heat transferred in the dryer, has been included for both types. This assumption is based mainly on Wimmerstedt and Hallstrom (1984), since results from actual measurements of operating dryers have been difficult to obtain.

The steam dryer model used here can be seen in figure 3.3. The wet biomass is introduced into the dryer and transported through it by means of low-pressure steam generated from the wet biomass itself. For start-up it would be necessary to provide low-pressure steam for this purpose. According to Fyhr (1996) the pressure level does not affect the drying rate significantly, and a choice can be made based on thermodynamic and practical considerations.

In a RSTOIC-reactor, the drying is modelled as a chemical reaction1 where some of the wet biomass is converted to water as the moisture content decreases. The dried biomass and the water is then transported to a heat exchanger where the materials are heated to the saturation temperature and the water from the fuel is evaporated.

The required energy is supplied through condensation of high-pressure steam. Various pressure levels are tested, the only constraint being a minimum temperature difference of 30°C in the dryer to ascertain reasonable drying rates [Wimmerstedt and Hallstrom, 1984]. The high-pressure condensate is pumped back to the steam generator. Another possibility investigated is extracting steam from the steam turbine at a suitable pressure, thereby avoiding a separate evaporator and steam drum for the drying steam.

In the FLASH2 model, the evaporated steam is separated from the dried biomass. Finally, this steam can be condensed to supply energy to district heating or feed water preheating.

The steam leaving the dryer is saturated or slightly superheated [Wimmerstedt and Hallstrom, 1984; Andersson, 1989], and here it has been assumed that the evaporated moisture is 1°C superheated. It is also assumed that the dried biomass exits the dryer at the same temperature12.

1 Biofuel is a ”non-conventional” component in ASPEN PLUS™, and as such is assumed to have a molecular weight of 1 kg/kmol. The chemical reaction for biomass drying then becomes: 1 biomass(wet) => (1/18.0152) H20. The extent of this reaction is set to achieve the prescribed moisture content of the dried fuel. 2 When the fuel exits the dryer the pressure decreases and the saturation temperature also decreases to 100°C. The fuel is cooled to this temperature through flashing of remaining moisture and the released steam could be used to preheat incoming fuel [Wimmerstedt and Linde, 1998]. This is not assumed here. Instead this energy is lost to the environment and the moisture content is constant, giving somewhat higher heat losses for this system than would be theoretically possible.

18 Exhaust gas

Wet fuel

RSTOIC Dried fuel

Dried fuel

FLASH2

Figure 3.3. Model of steam dryer.

Depending on how soon after drying the dried fuel is supplied to the gasifier, some of the sensible heat of the fuel is wasted by heat losses to the surroundings (e.g. in lock-hoppers and bins). This has to be taken into account when modelling the whole process. In the base case, it is assumed that the dried fuel is cooled down in the storage to 15°C before it is fed to the gasifier.

To operate this type of dryer some electricity is used to power mills and fans. Based on Ihren (1991) a total requirement of 400 kJ/kg DS has been assumed. This amount is kept constant also when varying initial and final moisture contents, although the amount of drying ought to affect also the electricity requirements.

It has further been assumed that the consumption in the fans, which are used to circulate the low-pressure steam, amounts to 140 kJ/kg DS. This latter amount is then regarded as useful in the form of heat in the dryer. As a comparison, Hulkkonen (1994) uses values in the range of 70 - 90 kJ/kg DS for low-pressure steam recirculation, and Wimmerstedt and Linde (1998) report a value of 125 kJ/kg DS for feeding and circulation fans in an existing dryer.

The excess low-pressure steam from the moist fuel is condensed and further cooled by means of district heating water in a heat exchanger. Due to the temperature levels in the district heating network, the cooled condensate still contains sensible energy. Although not utilised in the base case, this energy could be used to preheat the fuel before the dryer.

The second dryer model in this work represents an exhaust gas dryer. In the literature, various methods for calculating the energy required for drying in exhaust gas dryers, and the corresponding temperature levels, can be found.

19 When studying a system comprising a combined cycle with integrated drying and gasification, Ihren (1991) assumed that the energy required to dry wood chips from 45% to 15% moisture (wet basis) equals the heat of vaporisation for the evaporated moisture, plus the sensible heat required to heat the fuel and moisture to the wet bulb temperature. Based on the required drying energy and a presumed exhaust gas outlet temperature from the dryer, the temperature of the exhaust gas at dryer inlet is calculated.

In a simplified calculation on the energy requirement of an exhaust gas dryer, Wimmerstedt and Hallstrom (1984) made the same assumption as above, but disregarded the energy required for heating the dry fuel to the wet bulb temperature. In spite of this, they noted that the real fuel temperature may be even higher, especially for low final moisture contents. They also stated that the exhaust gas does not reach saturation and that a reasonable outlet temperature should be at least 10°C above the saturation temperature at outlet. In practice, the outgoing exhaust gas temperature is said to be 90°C or above.

In this work, the minimum outlet temperature of the exhaust gas out of the dryer has been set to 80°C in the base case. In some cases, this temperature may be higher to assure that the exhaust gas temperature is at least 10°C above the exhaust gas dew point temperature 1 at outlet. It is further assumed that the dried fuel leaves the dryer at the dew point temperature2.

The heat duty in the dryer is calculated as the energy required to heat the fuel and moisture to the saturation temperature, plus the energy for vaporisation of some of the moisture and the energy for heating the evaporated moisture to the pre-set outlet temperature. This also enables the calculation of the required exhaust gas temperature at the inlet, and hence, the available heat for steam production in the HRSG upstream from the dryer.

The model used for the exhaust gas dryer can be seen in figure 3.4. As in the case of the steam dryer, the wet biomass is supplied to a RSTOIC-reactor used to model the chemical reaction where some of the wet biomass is converted to water as the moisture content decreases. The water and the dried biomass are then heated in a HEATER-block (A), after which the water is separated from the dried biomass in a FLASH2-model. The water is taken to a second HEATER-block (B) where it is evaporated and mixed with the exhaust gas. It is now possible to calculate the saturation temperature of the exhaust gas.

Having determined the saturation temperature, the outlet temperature of the dried fuel is known and thereby also the heat duty, Q, in the first HEATER-block (A). This heat should also be supplied by the exhaust gas, and this information is passed to the second HEATER- block (B). The resulting exhaust gas temperature is then compared with the minimum allowable temperature, and if necessary, the steam production in the HRSG is adjusted. The exhaust gas can then be cooled further in a heat exchanger connected to the district heating

1 Dew point temperature is used, instead of adiabatic saturation temperature or wet bulb temperature, to simplify the calculations. When the exhaust gas is close to saturation the difference between these temperatures is small, and the corresponding error in exhaust gas and fuel outlet temperatures are minor. ‘ If the surface of the wood is assumed to be wet during the entire drying process, then the outlet fuel temperature equals the saturation temperature. For this to be valid, the diffusion resistance inside the fuel chips must be low. The model, therefore, best represents a case with small wood chips, high moisture content and limited moisture reduction. In this case, the final moisture content is rather low, implying that the material temperature may be somewhat higher. The influence on total plant performance is, however, small.

20 network. In the model, the possibility of exhaust gas recirculation has been included only as an electric loadT

Exhaust gas

Wet fuel

RSTOIC Dried fuel FLASH2

Exhaust gas + Dried fuel HzO(g)

Figure 3.4. Model of exhaust gas dryer.

The dried biomass is then transferred to a storage and, again, it is assumed that the dried fuel is cooled down in the storage to 15°C before it is fed to the gasifier. This corresponds to a loss in sensible energy.

Ihren (1991) provides some information regarding the electricity consumption in various exhaust gas dryers. The figures vary between 180 and 590 kJ/kg DS, mainly depending on the required amount of milling. Coulson and Richardson (1991) give a typical figure of 200 kJ/kg moisture evaporated for a pneumatic dryer, and Wimmerstedt and Hallstrom (1984) give the values 220 - 430 kJ/kg DS or 200 - 400 kJ/kg evaporated water (55% in, 10% out), stating that a major part of the energy supplied to the mill is also useful for drying.

Wimmerstedt and Linde (1998) report a total electricity requirement of 460 kJ/kg DS for a pneumatic conveyer type of dryer with milling and exhaust gas recirculation. Of this, 140 kJ/kg DS was used in the circulation fan. Kurkela and Solantausta (1994) in their calculations assumed a significantly lower consumption, in the range of 40 - 90 kJ/kg DS for feeding, milling and drying of wood, in a similar dryer.

In this report a pneumatic dryer is assumed, and the recycling of exhaust gas is believed to introduce an additional electricity requirement because of the fans. Using figures for a pneumatic dryer from Ihren (1991) and allowing for some additional fan work due to the recycling of exhaust gas, the result is a total electricity consumption of 330 kJ/kg DS. One third, or 110 kJ/kg, is used in the fans and included as heat in the energy balance of the dryer. If the mill was more closely incorporated in the dryer (e.g. ABB Flakt Biomasster) the major part of its electricity consumption could also be considered useful in the drying process, but this is not the case here.

In this type of dryer, up to 75% of the terpenes in the fuel can be emitted with the flue gas [Nyren, 1992]. This constitutes both an environmental problem and a loss of chemical energy, neither of which has been regarded in this model.

In order to separate fine fuel particles from the flue gas, cyclones and filters are often used [Magnusson, 1993]. This probably causes pressure drops which would result in a higher gas

21 turbine back pressure, thereby decreasing the output from the gas turbine. This has not been considered here: Instead, it is assumed that the fans in the dryer compensate for this pressure drop.

3.2.2 Gasifier

Various types of gasifiers are used to gasify carbonaceous materials like coal and biomass, yielding different products. A classification may be based on the method for supplying heat for the process (direct or indirect), the gasifying agent or agents used (air, oxygen, steam) or the operating pressure (pressurised or near atmospheric).

There are also numerous methods of creating a gas/solid particle interface to facilitate the chemical reactions. Presented below in figure 3.5 and table 3.3 are three reactor types, which differ substantially in the way this interface is created. These are moving bed, fluid bed and entrained flow gasifiers [Kristiansen, 1996; Reed, 1981].

Fuel Fuel + Air/O-

Reduction

Combustion

Air/O-

Moving bed (updraft) Fluid bed Entrained flow

Figure 3.5. Gasifier types.

Type Moving bed Fluid bed Entrained flow Operating temperature 1 (°C) 400-700 900-1100 1200 Fuel size (mm) 5-80 5 0.1 Feeding dry dry dry/slurry Oxidant air/oxygen air/oxygen oxygen

Table 3.3. Important characteristics of three types of gasifiers. [Kristiansen, 1996]

1 Temperatures are given for coal; biomass can usually be gasified at lower temperatures due to higher reactivity.

22 Of these, mainly the first two types are used for biomass [Beukens and Schoeters, 1985], whereas there are examples of all three types being used for coal gasification [Kristiansen, 1996].

The higher reactivity of biomass compared with coal enables the use of air as oxidant instead of pure oxygen, thereby avoiding the need for an air separation unit. The main drawback using air, is the lower heating value of the produced gas due to the large amount of nitrogen present.

For the purpose of producing power from gasified biomass, the fluid bed is considered to be the better option, mainly due to its greater fuel flexibility, better mixing and temperature control, and higher volumetric capacity [van Ree, 1994]. Using a circulating fluid bed, where char particles entrained in the gas flow are collected in a cyclone and returned to the gasifier, also enables very high carbon conversion.

Gasification of a fuel includes several physical and chemical processes, i.e. preheating, drying, devolatilisation (pyrolysis), tar cracking and gasification [Beukens and Schoeters, 1985; Double et al., 1989]. These processes can take place at different locations in the gasifier (moving bed) or simultaneously everywhere in the reactor (fluid bed).

For biomass, as opposed to coal, the effects of pyrolysis are important since the content of volatile matter is high. Some of the pyrolysis products merely add to those of gasification, giving a product gas of different composition from that if all gases from the pyrolysis participated in subsequent gas phase reactions [Beukens and Schoeters, 1985]. The tars produced by pyrolysis can either exit as part of the product gas, or be cracked into gaseous components, a process helped by higher temperature and/or catalysts. The gaseous components resulting from the cracking then may or may not participate in further gas phase reactions. General models for these phenomena are not available (to the author’s knowledge), though experimental data and models for pyrolysis of single particles exist.

When it comes to modelling the gasification chemistry, there are several levels of sophistication. The simplest model is a ’’black box” with known compositions in to and out from the gasifier, based on experimental data [Craig et al., 1994]. Methods for actual modelling of the gasification range from equilibrium models, assuming a fixed reaction temperature and two equilibrium equations [Bjerle et al., 1995], to transient models, based on mass and energy balances yielding a set of partial differential equations [Beukens and Schoeters, 1985]. Which model is chosen depends on the purpose of the model, e.g. designing a gasifier requires a very detailed model, whereas a simpler model might be sufficient for estimating gas composition when varying some operational parameters.

In this report, the aim is to study the thermodynamic behaviour of a complete system. Therefore, a relatively simple model that simulates steady state operation of a circulating fluid bed gasifier has been created. The basic idea in this model is that certain reactions attain chemical equilibrium.

The equilibrium composition depends on temperature, pressure and initial composition, and can be calculated through minimisation of Gibbs ’ free energy

G = H-T-S (3.1)

23 The change in Gibbs ’ free energy for a reaction is

AG = Zo, • AGfj (3.2) where 0/ is the reaction coefficient for substance z, and AGp is the Gibbs ’ free energy of formation from elements in their standard state.

According to Beukens and Schoeters (1985), the gasification process can be modelled with the following set of reactions:

C+2"O2 <=> CO (I)

C + O2 <=> CO2 (D)

C + C02 4=> 2CO m

c+h2o<=>co+h2 (IV)

C + 2H2 <=> CH4 (V)

CO + H20 <=> C02 + h2 (VI)

CHa+H2O^CO + 3H2 (vn)

Reactions (I)-(V) are heterogeneous (gas-solid) reactions representing char gasification, while (VI) and (VII) are homogeneous (gas phase) reactions.

At normal gasification temperatures, reactions (I) and (II) are fast, and are often considered to be irreversible [Beukens and Schoeters, 1985; Reed, 1981].

For reactions (III)-(VII), equilibrium compositions can be calculated for various temperatures, pressures and initial compositions. However, it is not certain that the calculated equilibria are attained in real gasifiers, mainly because the time required for reaching equilibrium may be longer than the actual residence time in the gasifier. Factors affecting how fast equilibrium can be attained are the operating temperature, the reactivity of the feedstock, the quality of the gas/solids contact and the type of reactor [Beukens and Schoeters, 1985; Reed, 1981].

In this work, the gasification process has been modelled in ASPEN PLUS™, using the theory presented above. No attempt has been made to model the drying and pyrolysis steps in the gasifier. This is a common strategy when modelling coal gasifiers where the content of volatile matter is low [Beukens and Schoeters, 1985]. This approach should be appropriate also for biomass, provided that there are small fuel particles with a large gas/solid interface and a high rate of mass transfer, together with long enough residence time. It has been used, for example, in Bjerle et al. (1995).

24 Although the original intention was to use the same model for both pressurised and near atmospheric gasification, the final models differ somewhat. This is due to differences in the gas cleaning methods, as described below.

In the case of pressurised gasification, the gas is cooled and passed through a filter at relatively high temperature, and then delivered directly to the gas turbine combustor. Hence, no further chemical reactions occur before the gas is combusted. Nor is any gaseous constituent added to or extracted from the gas before entering the combustor, and no additional compression work is required.

Therefore, the exact fuel gas composition is not essential for the overall cycle results, as long as the calculated composition closes the gasifier mass and energy balances. What is important is the split between sensible and chemical energy leaving the gasifier, since a major part of the sensible energy is recovered in the bottoming steam cycle, whereas the chemical energy is passed to the topping gas turbine cycle. Knowing the input composition and temperature, the gas outlet temperature and the gasifier heat loss, the fraction of incoming energy leaving as sensible energy is also known. With these quantities known, the chemical energy in the gas is also determined, and hence, more or less independent of the actual gas composition [Hughes and Larson, 1997].

For near atmospheric gasification, this approach proved to be insufficient. For the gasifier, the fraction of incoming energy leaving as chemical energy is still independent of the actual gas composition. The gas composition does, however, influence the subsequent equipment. Firstly, the gas is cooled to a low temperature and scrubbed with water to remove particles and gaseous contaminants, and then the gas is compressed to the gas turbine combustor delivery pressure. Here, the gas composition affects both the amount of water extracted from the gas in the scrubber and the required compression work. Hence, some knowledge about the actual composition is required.

The ASPEN PLUS™ model used for pressurised gasification is shown schematically in figure 3.6 below.

Fuel gas

Fuel (s) SPLIT

RYIELD RGDBBS RGIBBS HEATER Qi<

Inert gas Ash + unbumt

Figure 3.6. Pressurised gasifier model.

First, the dried biomass is decomposed in a RYIELD-reactor into solid carbon, oxygen, nitrogen, hydrogen, water, sulphur and ash. This could possibly be interpreted as a pyrolysis step, where the biomass is divided into volatiles and char. In a real fluid bed gasifier, some of

25 the constituents produced in the pyrolysis step are present in the product gas, mainly higher hydrocarbons and tars [Double et al, 1989]. This is not the case in this model. Instead, the approach taken here is to assume that any tars and higher hydrocarbons are cracked within the gasifier, and that all gases from the pyrolysis step participate in the subsequent reactions.

The constituents are then sent from the RYIELD-reactor to a RGEBBS-reactor1 in which air is introduced as the gasifying medium. The oxygen supplied with the air, as well as the oxygen present in the fuel, is assumed to react with carbon according to (U), thereby providing the necessary heat for the gasification process. It is further assumed that reactions (DI)-(V) attain equilibrium at a prescribed operating temperature12 . A value of the carbon conversion, based on experimental data found in the literature [Liinanki and Karlsson, 1994; Stahl, 1997; Salo et al., 1998], is also given.

After this, part of the ash and unbumt solid carbon present in the gas is withdrawn in a SPLIT- block. A HEATER-block is used to cool the gas, representing a heat loss from the gasifier. Finally, the mixture of gases is sent to a second RGEBBS-reactor where the gas phase reactions (VT) and (VH) take place.

It is, however, not certain that the equilibrium composition at the operation temperature is attained in a real gasifier, mainly because the time required for reaching equilibrium may be longer than the actual residence time in the gasifier. Factors affecting how fast equilibrium can be attained are the operating temperature, the reactivity of the feedstock, the quality of the gas/solids contact and the type of reactor [Beukens and Schoeters, 1985; Reed, 1981].

In order to take this into account, the equilibria in the second reactor are restricted by temperature approaches, the values of which influence the gas composition. In table 3.4 the values used here are presented, and the resulting gas compositions are comparable with data presented in the literature [Liinanki and Karlsson, 1994; Stahl, 1997].

Reaction ATaDD co+h2o<^co2 + h2 -150°C CHa+H20&C0 + 3H 2 -150°C

Table 3.4. Temperature deviations for equilibrium calculations.

In addition to this, two reactions are included to take into account the formation of ammonia and hydrogen sulfide. It is assumed that these reactions reach equilibrium.

N2+3H2<^2NH3 (Vffl)

s +h2 &h2s (IX)

The amount of gasifying medium (air) required in the process is calculated from the overall energy balance and the condition that both reactors should operate at the same temperature.

1 RGIBBS calculates equilibrium by Gibbs energy minimisation. 2 This might not always be true in a real fluid-bed gasifier, since the temperature is somewhat low [Beukens and Schoeters, 1985].

26 This is found to give air-flows in fair agreement with reported experiments [Liinanki and Karlsson, 1994fStahl, 1997; Faaij, 1997].

Next the model used for near atmospheric gasification is presented. As was stated, above a reasonable gas composition is vital when calculating the work required to compress the fuel gas. Hence, this model is partly based on information from the Swedish manufacturer TPS Termiska Processer [Larsson, 1998]. The ASPEN PLUS™ model used is shown in figure 3.7.

Inert gas Ash + unbumt

Figure 3.7. Near atmospheric pressure gasifier model.

Also here the dried biomass is decomposed in a RYIELD-reactor into solid carbon, oxygen, nitrogen, hydrogen, water, sulphur and ash. These constituents are then mixed with air and nitrogen in a RGEBBS-reactor. Here, only one RGIBBS-reactor has been used, calculating the equilibrium composition based on eight independent chemical reactions (this approach could also have been used for the pressurised version). The reactions, and the temperature deviations from the prescribed gasifier operating temperature, for equilibrium calculations, are presented in table 3.5.

These reactions do not necessarily occur in Reaction ATaon a real gasifier, but they can be used to 2C + 02 <=>2C0 -210 calculate the equilibrium composition. C+02 002 -260 -310 With gasifiers operating at temperatures 2H2 + 02 <=> 2H20 around 900°C, some heat losses are S + h2 ■$=> H2S -200 inevitable. Furthermore, the gasifiers N2+3H2<^2NH3 -600 require some electricity, mainly for the C + 2H2 <=> CH4 -300 lock-hopper systems. Consonni and Larson 2C + 2H2<*C 2H4 -300 (1996a) assumed a set of values for these -650 factors, presented in table 3.6. 2C + 3#2<=>02#6

Table 3.5. Reactions and temperature deviations for gasification at near atmospheric pressure.

27 Gasifier pressure Heat loss Electricity requirement (fraction of biomass HHV) (kJ/kg dried fuel, 15% moisture) Near atmospheric 0.01 10 30 bar 0.006 80 48 bar 0.006 110

Table 3.6. Heat loss and power consumption for various gasifiers [ Consonni and Larson, 7996^7.

In a later paper, the same authors [Larson et ah, 1998] have modified their figures on heat losses to 0.0121 and 0.0097, for near atmospheric and pressurised gasification, respectively.

Neergaard (1998) states the heat losses for a pressurised plant of the size studied here is about 1 % or lower.

According to calculations performed at TPS Termiska Processer the corresponding heat loss from a gasifier of TPS design, with gasification and tar cracking in two separate vessels, would be around 3% [Larsson, 1998].

In this work, the values of 0.02 and 0.01 have been used for heat losses in near atmospheric and pressurised gasification, respectively. Regarding the electricity requirement, the values in the first two rows of table 3.6 are used.

Consonni and Larson (1996a) also estimated the power consumption for inert gas production to be 300 kJ/kg. This value is probably dependent on the type of inert gas used; in this work, the power consumption for inert gas production has not been included.

3.2.3 Gas cleaning

Cleaning of the fuel gas may be required for two reasons; to protect downstream equipment and/or to protect the environment. Starting with the equipment, the gas turbine is the most sensitive part. The main concerns are particles and alkali compounds potentially harmful to the hot parts of the gas turbine.

The amount of fuel, char and ash particles carried with the gas out of the gasifier vessel varies widely in different types of gasifiers. If the gas is to be burned in a gas turbine combustor, the particles have to be separated in order to protect the turbine blades from deposition and erosion. The alkali metals, present as vapours in the hot fuel gas, can cause high temperature corrosion if condensed on vanes and blades in the gas turbine.

Different gas turbine manufacturers present different requirements regarding permissible concentrations of particles and alkali in the gas. In table 3.7 some figures on permissible concentrations of contaminants valid for the exhaust gas stream in a GE LM2500 gas turbine are given [Faaij, 1997]. It should be noted that the LM 2500 is an aero-derivative gas turbine. Generally, industrial gas turbines are considered less sensitive to these contaminants.

28 Contaminant Maximum concentration Calculated maximum in gas flow to expander concentration in fuel gas (ppbw) (ppbw) Solid particles: d < 10 um 600 3000 10 13 |im 0.6 3 Alkalis (Na + K + Li) 4 20 Alkali-metal sulphates 12 60 Chlorides 500 2500 Condensable tars - 0.008 mg/Nm3

Table 3.7. Permissible concentrations of some contaminants. [Faaij, 1997]

Bridgwater (1995) also presents a set of gas turbine fuel specifications, partly reproduced in table 3.8. The higher values stated for permissible particle concentration indicate that table 3.8 is valid for industrial gas turbines.

Contaminant Permissible concentration in fuel gas Solid particles: 4 < d < 10 |im < 10.0 ppmw 10 20 um <0.1 ppmw Alkalis metals 20 - 1000 ppb Alkali metals + sulphur <0.1 ppm Sulphur (HzS + SO2 etc.) < 1 ppm HC1 < 0.5 ppm Tars All in vapour phase or none

Table 3.8. Gas turbine fuel specifications [Bridgwater, 1995],

In tables 3.7 and 3.8, chlorides (corrosive) and tars are also included. The chloride content of woody biomass is normally low, and in the case of dolomite present in the gasifier or tar cracker, much of the hydrogen chloride (HC1) produced by gasification will be adsorbed. Furthermore, HC1 reacts with particulates which are later removed. Hence, chlorides are normally not a problem in these applications [Faaij, 1997].

Tars are higher hydrocarbons, mainly benzene and naphthalene, produced in various amounts in gasification. The concentration of tars in the fuel gas depends on reactor type, feedstock and gasification temperature. At higher concentrations, these compounds may constitute a substantial part of the fuel gas heating value.

As long as the tars remain in the gas phase, they do not constitute a problem for either the gas cleaning equipment or the gas turbine. But in order to reduce particle and alkali concentrations, the gas has to be cooled in the gas cleaning system. When the gas is cooled, tars may condense, and this results in fouling of heat exchangers and blocking of filters. This

29 also constitutes a loss of chemical energy. Hence, it is of interest to reduce the tar concentration before gas cooling, preferably by converting the tars to non-condensable gas constituents, such as, CO, H2 and CH4, preserving most of the chemical energy.

Cleaning the fuel gas to protect the equipment is also beneficial for the environment. There is, however, one additional pollutant that has to be controlled exclusively for environmental reasons, NOx. Due to the low heating value of the fuel gas, resulting in a lower flame temperature, the formation of thermal NOx should be lower than that of natural gas. This has been confirmed in tests [Neilson et al., 1998] with low LEV gas in a modified LM2500 combustor, indicating a 50% reduction in thermal NOx compared to natural gas firing.

However, biomass contains various amounts of fuel-bound nitrogen, which during gasification forms mainly molecular nitrogen (N2) and ammonia (NH3), but also cyanides (HCN, CN) and nitrogen monoxide (NO) [Leppalahti et al., 1994; Faaij, 1997; Rosen et al., 1997]. The amount of ammonia in the gas is influenced by the type of gasifier and the operational parameters. As an example of the latter the concentration is reported to increase with pressure and decrease with increased freeboard temperature [Leppalahti et al., 1994]. Depending on fuel type and gasification conditions, 40 - 95% of the fuel bound nitrogen can be converted to ammonia, during pressurised gasification [Stahl, 1997; Salo et al., 1998]. During combustion, NH3 is partly converted to NOx. Tests have shown conversion from NH3 to NOx in the range 5 - 80%, the higher conversion rate found at lower concentrations of NH3 in the fuel gas [Liinanki and Karlsson, 1994; Stahl, 1997; Layne and Hoffman, 1998].

A crude estimation of the NOx emissions based on the fuel used for the calculations in this report, reveals that this could be a serious problem. The fuel has a nitrogen content of 0.0045 kg N/kg DS and, based on the references above, conversion factors of 0.70 from fuel nitrogen to NH3 and 0.50 from NH3 to NOx are assumed. The resulting NOx emission would then be around 300 mg N02/MJf. This is well above the current Swedish emission standards for this type of plant. Using fuels with lower nitrogen content, tuning gasifier operational parameters and introducing new gas turbine combustor technology, thereby reducing the conversion from NH3 to NOx, could lower this NOx emission substantially. Even so, compliance with the emission standards may require additional measures as discussed below.

Cleaning of the gas from the gasifier is generally performed as hot gas cleaning or cold gas cleaning, depending on the gasifier technology. When gasification is performed at near atmospheric pressure, the resulting gas has to be compressed before entering the gas turbine combustor. To limit the power for compression, it is desirable to cool the gas before the fuel compressor. Hence, cold gas cleaning is normally utilised. When, on the other hand, the gasifier is pressurised, the gas pressure is sufficient for direct injection into the gas turbine combustor, and hot gas cleaning is chosen in order to preserve more of the sensible energy in the gas.

It was stated above, that the amount of particles in the fuel gas must be small to avoid erosion and deposition in the gas turbine. Equipment used to separate particles are, for example, cyclones, filters and scrubbers.

Cyclones are used for the separation of fly ash and char particles from the gas stream exiting the gasifier. If the char is returned to the gasifier, the carbon conversion is improved. The collection efficiency of cyclones depends strongly on particle the size and is often not

30 sufficient for the smallest particles. Hence, the cyclone is often followed by a filter and/or a scrubber.

In hot gas cleaning, the gas is cooled to a temperature in the range of 400 - 550°C and then passed through a filter. At that temperature, most of the alkali metals in the gas are condensed on particles in the gas stream and withdrawn with the particles in the filter. The temperature is, however, high enough to prevent all but the heaviest tars to condense [Stahl, 1997; Salo et al., 1998].

Filters operating at these temperatures in a reducing environment and at elevated pressure are now being developed and tested in a number of pilot and demonstration plants. Westinghouse, for example, participate in a number of projects testing ceramic barrier filters in IGCC and PFBC plants at 370 - 900°C. Most of the plants use coal as fuel, but some results are also reported for gasified bagasse [Newby et al., 1998]. In the demonstration plant in Vamamo where forest residues are gasified, a ceramic barrier filter has been used to clean gas at 350 - 400°C [Stahl, 1997]. Presently (October 1998), however, metallic filter elements are used [Stahl, 1998].

Encouraging test results aside, the reliability and availability of these filters, especially for biomass, still have to be proven in continuous operation.

Filters operating at low temperature, sometimes used in cold gas cleaning, are well proven and commercially available. As the gas is typically below 200°C [Faaij, 1997] when entering these filters, a low tar content is important to avoid filter blockage.

One way to reduce tar levels is by thermal cracking, since it has been shown that higher temperature reduces the tar levels [Simell et al., 1994; Stahl, 1997]. Another option is to use a catalyst to aid the decomposition of the tars. Catalytic cracking can, in principle, be performed either inside the gasifier vessel by using a catalyst, e.g. dolomite or calcite, as bed material, or in a separate vessel downstream of the gasifier. The latter method is reported capable of achieving 99% tar conversion at 900°C, resulting in levels below 65 mg/Nm3 [Simell et al., 1994]. Rensfelt, (1997) reports similar results for the TPS near atmospheric gasifier pilot plant with a separate CFB tar cracker, using dolomite as a catalyst. Catalytic tar cracking inside the gasifier has so far been more difficult to achieve, and research is continuing. Resulting tar levels of 5 g/Nm3 are reported [Stahl, 1997].

In cold gas cleaning, a wet scrubber is often used as a final stage in the gas cleaning. Within the scrubber, small particles and some gaseous constituents like ammonia can be separated. At the same time, the gas is often further cooled in the scrubber.

Using a scrubber, the amount of ammonia in the gas fed to the gas turbine combustor, and hence, the NOx emissions from the gas turbine, can be significantly reduced. For hot gas cleaning, other measures must be taken to limit the NOx emissions. Possible actions are to use only fuels with low nitrogen content or to incorporate a high temperature Ni-catalyst in the fuel gas stream. The latter approach has proven successful in laboratory tests on peat gasification [Leppalahti et al., 1994]. A somewhat different concept is known as selective oxidation, where a small amount of air is added to the fuel gas. In laboratory scale tests on coal, a 90% reduction of ammonia was achieved [Dawes et al., 1997]. Yet another possibility is catalytic reduction of NOx in the HRSG.

31 In the present work, two models for gas cleaning have been used. For pressurised gasification, hot gas cleaning" has been provided, and for near atmospheric gasification, cold gas cleaning. The main purpose of these models is to take into account the influence of the gas cleaning equipment on the overall thermodynamic behaviour of the plant. The models are, hence, quite simple and do not give any information on real separation efficiencies. Input concerning pressure drops, etc., are presented in section 3.3.

The hot gas cleaning includes a cyclone and a ceramic filter. In ASPEN PLUS™, the cyclone has been modelled by means of a SSPLIT block, since the available cyclone block seems to work only at near atmospheric pressure. For this block, a separation efficiency is given as input. The ceramic filter is simply modelled as a HEATER block, where a pressure drop and a temperature drop are given. Before entering the filter, the gas is cooled in a heat exchanger producing saturated steam to the steam cycle.

In cold gas cleaning, the fuel gas passes through a cyclone and a wet scrubber. This means that the gas is cooled to a lower temperature, before it is compressed to the delivery pressure. Some of the cooling is performed in the scrubber, but in order to recover as much of the heat as possible, an evaporator, a clean gas heater and an air preheater are placed upstream from the scrubber. The model of the scrubber is very simple, but sufficient for the thermodynamic calculations, and consists of a HEATER block where the gas is cooled. It is assumed that the gas is saturated with water vapour when leaving the scrubber [Bain et al., 1996]. After the scrubber, the gas is compressed in an intercooled compressor.

3.2.4 Gas turbine

Traditionally, gas turbines have used high quality fuels for production of electricity, and therefore, high efficiency has been of great interest for lowering the cost of electricity. This has been the incentive for developing more advanced configurations, like the combined cycle, with higher efficiencies. At the same time, research is conducted to enable the use of cheaper fuels in the gas turbine. There are a number of different approaches to this, e.g. closed cycle gas turbines incorporating a high temperature heat exchanger, combustion of pulverised solid fuels in the gas turbine combustor, and also gasification of solid fuels and subsequent combustion of the gas in the gas turbine combustor. All these require some redesign of the gas turbines that are currently built for oil and/or natural gas.

In the case of biomass derived gas, the heating value is substantially lower than what is normal for natural gas, and the flammability and the burning velocity are different. The lower heating value means that for the same power, the fuel flow has to be increased. This affects not only the design of the combustor, but also the matching between the turbine and the compressor.

The use of biomass derived gas as a gas turbine fuel has been treated in numerous studies, e.g. by Chellini (1998), Liinanki and Karlsson (1994) and Neilson et al. (1998). Generally, some modifications in the combustors and fuel piping are required to handle the larger flow and differing physical properties. As an example, tests performed at the GE laboratory in Schenectady, N.Y. showed that a typical fuel from the gasification of biomass could be used in a modified Frame 6 B combustor. Stable combustion was reported for various points of operation, and the thermal NOx-emissions were less than 10 ppm (15% oxygen) for gas LHV

32 in the range of 4.5 to 6.0 MJ/Nm3 [Liinanki and Karlsson, 1994]. In later tests, encouraging results have been achieved also for the smaller LM2500 combustor [Neilson et al., 1998].

Another problem that arises, when trying to use a gas turbine, which was originally developed for burning natural gas, in an IGCC plant, is the increase in mass flow through the turbine compared with the flow through the compressor. Assuming choked flow conditions at the turbine inlet at design point this requires an increase in pressure or a decrease in temperature, since the non-dimensional mass flow, 03/PQ3, at turbine entry is then constant at its maximum value [Cohen et al., 1996]. Decreasing the turbine entry temperature, 7 o3, is disadvantageous, since it reduces specific power and efficiency. Increasing the pressure, on the other hand, increases the compressor delivery pressure, and thereby reduces the stall margin of the compressor. To avoid the stalling, the compressor mass flow can be decreased by adjusting the inlet guide vanes. Another possibility is to rebuild the turbine to accommodate the larger flow, while maintaining the compressor mass flow, pressure ratio and turbine entry temperature at design values.

In the case of pressurised gasification, air to the gasifier is often bled from the compressor outlet. This results in a smaller total flow to the combustor than in the case of the gasification without air extraction. There are, however, examples where low LHV gas is burnt in existing gas turbines, without any compressor bleed off or modifications of turbine geometry.

According to Jeffs (1998), the ABB GT11N2 equipped with a new compressor, an enlarged combustion chamber and modified burners, is able to bum blast furnace gas with heating values as low as 2.3 - 4.0 MJ/kg. Tests showed NOx-emissions less than 35 ppm (15% oxygen) dry. In order to stay within the capacity of the original turbine with a fuel flow 17 times greater than the fuel flow with natural gas, the air flow through the compressor is reduced accordingly by closing the inlet guide vanes.

Another example of large gas turbines running on low LHV gas, is the IGCC plant at a steel mill in Taranto, Italy. Here, three MS900IE gas turbines run on a gas mixture with a design LHV of 7.5 MJ/kg, occasionally dropping to 6.2 MJ/kg. Despite the larger volume flow, no modifications have been made to the turbine inlet section, and the resulting rise in back pressure is handled without compressor bleed off [Chellini, 1998].

Since it is difficult to know exactly which measures, in terms of de-rating or rebuilding, a specific gas turbine require in order to utilise gasified biomass as a fuel, the calculations in this work are performed at design point. This means that the gas turbines are able to operate at the same pressure ratios and firing temperatures, regardless of the gasification technology, i.e. the gas turbine models represent machines at the same technology level, but customised to different mass flows. The model used for the gas turbine is presented in figure 3.8, and the main input parameters are presented in table 3.9. The parameters that are varied are the compressor pressure ratio (PR) and the combustor outlet temperature (COT)1.

1 ’’Combustor outlet temperature” is equal to ’’turbine entry temperature”, and is defined as the flow weighted mean total temperature of the working fluid immediately upstream of the turbine stator vanes, i.e. before mixing with stator cooling air. This.temperature is higher than the ’’turbine reference inlet temperature”, often referred to as ISO-TIT, which is the calculated turbine inlet total temperature based upon the combustor inlet temperature, the net energy release in the combustor and the compressor inlet flow plus the fuel mass flow [ISO 11086,1996].

33 Parameter Value used in calculations Compressor polytropic efficiency, T|ac 0.89 Turbine stage polytropic efficiency, T|D.st 0.89' Mechanical efficiency, T|m 0.99 Combustion efficiency for H2 1.00 Combustion efficiency for CO 1.00 Combustion efficiency for CH4 1.00 Combustion efficiency for C2H4 1.00 Combustion efficiency for C2H6 1.00 Pressure loss, compressor intake, Ap/p 0.01 Pressure loss, combustion chamber, Ap/p 0.04 Heat loss, combustion chamber, as fraction of fuel LHV 0.005

Table 3.9. Parameters used in the gas turbine model.

VRSTOIC/

COMPR

Figure 3.8. Gas turbine model for parametric study.

The compressor intake pressure loss is given as input to a HEATER block (HI). The air is then compressed using a COMPR model, where the polytropic efficiency is used to calculate the polytropic coefficient n trough

/z — l v r ; (3.3) n rip where yis the heat capacity ratio Cp/cv . These properties are calculated at average conditions for the compression, resulting in an iterative calculation of the compressor. 1

1 Ideal efficiency, used when cooling flow to stage is zero.

34 The enthalpy change in the compressor is then given by ASPEN Technology (1995):

n-1

Pin-vin Ah = ■ -l (3.4) M-l \Pi>in y n

In a FSPLIT-block, the air stream from the compressor is split into three or four streams, feeding the gasifier (in the case of pressurised gasification), the combustion chamber and the two cooled turbine stages.

The pressure loss in the combustion chamber is modelled separately in a HEATER-block (H2), after the RSTOIC-block which calculates the heat release from the chemical reactions. Five reactions are taken into account, and it is assumed that the reactants H2, CO, CH4, C2H4 and C2H6 are completely consumed. A heat loss from the combustion chamber is also included.

The turbine model comprises four stages, all of them modelled as isentropic turbines 1 using the COMPR block in ASPEN PLUS™. The isentropic efficiency is defined through

hput fyn Vs = (3.5) Kut ~ \n where hsout is the outlet enthalpy assuming isentropic expansion to the specified outlet pressure [ASPEN Technology, 1995].

When calculations are performed over a range of pressure ratios, it is recommended to use the polytropic efficiency instead [Cohen et al., 1996]. Therefore, Fortran blocks are included to calculate the isentropic stage efficiency corresponding to a given polytropic efficiency. The relation between isentropic and polytropic efficiency [Cohen et al., 1996] is given in eq. (3.6)

y—1 frAT*" 1- v Pin j (3.6) 's,st 7-1 1- l Pin J

Moreover, the first two of the four turbine stages are cooled with air bled from the compressor outlet, as shown in figure 3.8. The amount of cooling air and its influence on stage efficiency are calculated in the following manner, based mainly on Rosen (1993).

The calculation of the required cooling flow is based on the concept of blade cooling effectiveness: 1

1 ASPEN PLUS™, version 9.3 does not comprise a turbine model using polytropic efficiency.

35 ~ Tmtrl (3.7) where T\ is the gas inlet temperature, Tmtr[ the mean surface temperature of a stage and Tcooi the inlet temperature of the cooling medium. Here, Tcooi is always the air temperature at compressor outlet, since no intermediate extraction from the compressor is considered.

The blade cooling effectiveness has been shown to be a function of a non-dimensional parameter, B, given by

g = (3.8) abi,i ’ Abi,i where mcoo[ i is the cooling flow to stage i, cpxoo! is the specific heat capacity of the cooling air at the inlet temperature, is the average heat transfer coefficient in stage i and Abu is the stage surface area [Hasselbacher, 1989]. It is also possible to introduce a specific stage surface area in the form

abi,i (3.9)

where mg I- is the mass flow of gas in to stage i, before mixing with cooling air. The introduction of abu makes the model more general, particularly if this value may be considered constant for turbines of different sizes. This simplification is used here.

In Hasselbacher (1989), curves are drawn showing the blade cooling effectiveness as a function of B for different cooling methods. Rosen (1993) translated these into polynomials, and the result is shown below for two methods.

For convective cooling:

£c = 3.0000 • 10~3 + 0.63777 • B - 0.25554 • B2 + 3.1808 • 10~2 • B3 (3.10) and for film cooling:

£c = 0.15984 + 0.39620• B -4.869 -10~3 • B2 - 2.9461 • 10'2 • B3 (3.11)

Given a maximum allowable blade material temperature as input, it is now possible to calculate the required cooling flows.

In film cooling, the cooling air is always mixed with the main stream, and very often air used for convective cooling is also injected into the main stream. This causes disturbances in the flow through the expander, resulting in a lower stage efficiency than for an uncooled stage. This is taken into account by means of a semi-empirical expression [Traupel, 1988], giving the resulting polytropic efficiency for stage i as

36 g mcool4 Pi4n Pp4 Pp 4,uncooled (3.12) P\4n ~ Pc,out where pis the pressure at the inlet of stage i, piiin is the pressure at the inlet of the first stage and pc,out the pressure after the last cooled stage in the turbine. Furthermore Traupel (1988) proposed

( P\4n ^ K = •In (5 + 05) (3.13) %-cool V Pc,out) where zcooi is the number of cooled stages in the turbine and 5 is a constant relating the decrease in polytropic efficiency to the amount of injected cooling air. Based on experience, 5 is in the range of 0.5 -1.0. The last term in equation (3.13), 0.5, is included to compensate for the error made when introducing and mixing the cooling air with the main flow before the stage1 [Traupel, 1988].

As in Rosen (1993), it has been assumed here that the pressure ratios for the stages, njr, are equal according to

\ P\,in ust = (3.14) \Pn,out j where zt0t is the total number of stages in the turbine and pn.om is the pressure after the last stage in the turbine.

In the ASPEN PLUS™ calculations, equation (3.14) is used in a Fortran block to calculate the pressure ratio over each stage. Equations (3.7) - (3.11) are included in other Fortran blocks to calculate the mass flow of cooling air to the first two stages. Then equations (3.12) and (3.13) are used in two Fortran blocks to correct the polytropic efficiency. Finally, the corresponding isentropic stage efficiency is calculated by means of equation (3.6).

The final outcome of this cooling flow model is strongly dependent on the values of the heat transfer coefficient, the surface area of the stages, and the constant 5. This information is also rather difficult to obtain from the literature. In these calculations, the following values have been used: abU = 2000 W/m2K abU = 0.07 m2s/kg aWi2 = 0.09 m2s/kg 5 = 0.1

A deficiency in this model is that rotor cooling is not included, leading to somewhat lower cooling flow requirements than in reality. Nevertheless, it is possible to obtain reasonable

1 This theory was presented by Traupel back in the 1950's, and since then the development in the area of gas turbine cooling has been substantial. Hence, the value of (5+0.5) should probably be lowered to reflect turbines of today.

37 results for efficiency and turbine outlet temperature by ’’tuning” the variables S, and a#/, if one, for example, wants to model a specified gas turbine on the market.

It should also be noted that by this approach, the same level of cooling technology has been applied to different combustor outlet temperatures. This results in an optimal COT above which the decrease in efficiency due to increased cooling flows is greater than the increase due to higher inlet temperature. In reality, the COT would probably not be raised, unless the cooling technique was also improved. One exception is when the specific power is of great value, because this number still increases with increasing COT.

3.2.5 Boost compressor

In the case of pressurised gasification, air is bled from the gas turbine compressor, cooled, further pressurised in a boost compressor and then delivered to the gasifier. The boost compressor is required to overcome pressure losses in the gasifier and gas cleaning equipment, and to assure that the fuel gas is delivered at the gas turbine fuel control valve at the prescribed pressure.

Feed water

Air from gas turbine Air to gasifier compressor

Figure 3.9. Boost compressor for pressurised gasifier.

In figure 3.9 the cooling medium used to lower the compressor inlet temperature, is steam cycle feed water. Another possibility would be to use a recuperator to heat the air from the boost compressor, by means of the air from the gas turbine compressor. This would decrease the amount of air required for gasification and correspondingly increase the fuel gas heating value. It would, however, also require a final cooling by means of the feed water to reach a sufficiently low compressor inlet temperature. In the base case, no recuperator is included, and only feed water is used for cooling. The air temperature at the gasifier inlet then varies with the boost compressor pressure ratio, with the resulting value for the base case being around 200°C.

38 3.2.6 Fuel compressor

In systems with gasification at near atmospheric pressure, the fuel gas has to be compressed before it can be supplied to the gas turbine combustion chamber. Normally, the delivery pressure is 10 - 50% above the combustor pressure, mainly for control reasons.

The compression requires a significant amount of energy, and in order to decrease this amount, the fuel gas compressor is divided into three stages with intercoolers in between, as shown in figure 3.10.

In the intercooler, heat is transferred from the fuel gas to a part of the feed water flow, i.e. it works as a feed water preheater. To further decrease the gas temperature, additional heat exchangers with cooling water are used. This means less compression work, but also increased heat loss to the surroundings.

Feed Cooling Feed Cooling water water water water

Fuel gas from Gas to scrubber recuperator

Condensate Condensate

Figure 3.10. Fuel gas compressor with intercoolers.

It is assumed that the inlet temperature to each stage is the same and that some of the water in the fuel gas is condensed and withdrawn from the intercoolers.

After the fuel compressor, the gas is routed to a recuperator and heated before entering the gas turbine combustor. In that way, more of the sensible energy in the gas from the gasifier is utilised in the gas turbine.

3.2.7 Heat recovery steam generator

In the base case, steam for the steam turbine is produced at a single pressure level, resulting in a rather simple HRSG1. The models used for the calculations are presented in figures 3.11 and 3.12.

The HRSG presented in figure 3.11 has a separate evaporator for production of steam to a steam dryer. If instead, the steam to the dryer is extracted from the steam turbine, this evaporator is omitted, and the HRSG is of true one pressure design. In figure 3.12, the HRSG

1 When taking into account part load performance and plant dynamics, the layout probably has to be modified to ensure proper operation. This will presumably require a more complex HRSG.

39 used in systems with exhaust gas dryers is shown. Here, the exhaust gas leaving the economiser goes to the dryer, and then to a heat exchanger for district heating water (not shown in the figure) before it leaves through the stack.

1. Superheater 2. Evaporator 3. Economiser I 4. Heat exchanger for district heating 5. Evaporator, drying steam 6. Economiser H 1. Superheater To stack 2. Evaporator 3. Economiser

District heating water To dryer

Feed water

Steam to dryer

Feed water

Condensate Water to gas cooler from dryer

Water to gas cooler Steam from gas cooler Steam from gas cooler

Steam to turbine

Steam to turbine Gas turbine exhaust

Gas turbine exhaust Figure 3.12. Single pressure HRSGfor plant with exhaust gas dryer. Figure 3.11. HRSGfor plant with steam dryer.

40 3.2.8 Steam turbine

The thermal efficiency of the steam cycle depends on steam inlet and outlet data, and on the isentropic efficiency of the turbine itself. From a thermodynamic point of view, the inlet data, pressure and temperature should be as high as possible; while at the same time, the outlet pressure ought to be low. In this case, the outlet pressure is mainly limited by the required temperature levels in the district heating net. The inlet steam pressure and temperature are normally limited by material constraints. The pressure is sometimes further limited, especially for small turbines with a low output, by the desire to keep the first stage blades reasonably high. Furthermore, a high inlet pressure may lead to high steam wetness before the last stage, and this increases the risk of erosion.

The calculations performed in this report consider rather small plants with total outputs of 29 - 37 MWe and steam turbine outputs in the range of 7 - 18 MWe. Although several turbines of this size have been operated with steam pressures in the range of 100 - 160 bar [Mayer et al., 1998], the inlet steam pressures used in the parametric study are on the high side with respect to the first stage blade height, even if it is assumed that the turbines are geared and equipped with partial arc admission. As a result of this, the smaller turbines are likely to show lower efficiencies than the larger ones, but this has not been taken into account here. For larger plants, e.g. a plant with two gas turbines and a single steam turbine, yielding a plant with an electrical output of 60 - 80 MWe, the higher pressure levels could be more adequate.

The steam turbine has been modelled using three individual turbines in ASPEN PLUS™. One reason for this is that there is no model of a turbine with extraction in ASPEN PLUS™. The other reason is a desire to consider decreasing turbine efficiency once the two-phase region is entered. This results in a model presented in figure 3.13.

To feed To water tank condenser

Figure 3.15. Schematic steam turbine model.

The pressure after the high-pressure part of the turbine, HPST, is calculated by iteration, with the constraint that the liquid fraction should be in the interval between 0 - 0.001. Once this is done, the isentropic efficiency of the next part of the turbine, IPST, is recalculated using the following expression [Kehlhofer, 1978]:

(l-X,„) + (l-Q ^P ^p4ry o (3.15)

41 Here, r\p is the polytropic efficiency of a turbine section exposed to wet steam and rjp_dry the corresponding efficiency for the turbine section working in the superheated region. It should be noted here that equation (3.15) is used for isentropic efficiency, even though it was originally given for poly tropic efficiency.

Equation (3.15) may also be compared with the following expression, presented by Traupel (1988), for the isentropic efficiency of a steam turbine working with wet steam:

-7) (3-16) where % and T)s0 are the isentropic turbine efficiencies for wet and dry steam, respectively, a is the Baumann factor, often in the range of 1 - 1.2, and y is the average moisture fraction in the turbine section. When comparing equations (3.15) and (3.16), it can be seen that using equation (3.15) with r]p,dry = 0.85 for calculating an isentropic stage efficiency corresponds to using T]s0 = 0.85 and a Baumann factor of 1.18 in (3.16). Hence, using equation (3.15) for isentropic efficiency is similar to using equation (3.16).

The outlet pressure of the IPST is determined by the pressure in the feed water tank. The last part of the turbine, LPST, is also exposed to wet steam, and equation (3.15) is used once more. The outlet pressure of the LPST depends on the desired district heating water temperature out from the condenser and the pinch point.

In some plant layouts, it is possible to route the district heating water first to the condenser, giving a comparably low condenser pressure. In other cases, the final heating has to be performed in the condenser, and the pressure is then higher.

42 3.3 Parametric study

The parametric study aims to compare a number of different IGCC systems, and to investigate the influence of some parameters on the electrical efficiency and the fuel utilisation. Four base case configurations are formed by combining each of two technologies for gasification, pressurised and at near atmospheric pressure, with the two types of dryers, one using steam and the other using exhaust gas as the drying medium.

In this study, all calculations are performed at design point, meaning that, for example, component efficiencies and pressure drops do not depend on the mass flow through the component1. In other words, the components are adjusted to fit each system resulting from the parameter variation. This is a somewhat ’’academic ” approach, as opposed to the more practical questions about what happens when some parameters are varied with a given set of equipment. It is, however, believed that the approach used here is more general, comparing in a sense equally optimised systems. It certainly is less complicated, since it does not involve the calculation of the off-design behaviour of gas turbine compressors and other components.

In table 3.10, the parameters varied in the study are presented. Figures in extra bold type are considered as base case values. In addition, the system layouts are also varied in some cases.

Parameter Steam pressure at HP turbine inlet 80/100/120/140 bar Gas turbine combustor outlet 1200/1300 °C temperature 2,3

Gas turbine compressor pressure ratio 5-25 - Ambient temperature -5/5/1514 2 3 °C Moisture content wet fuel 30/40/50/60 % wet basis Moisture content dried fuel 10/15/20 % wet basis Fuel temperature in to gasifier 15°C/delivery temp, from dryer Fuel gas temperature at delivery to 380/530 °C combustion chamber Ratio of fuel delivery pressure to 1.1/1.4 pressure in combustion chamber Heating steam pressure in steam dryer 6/9 bar Evaporator pinch-point 5/10/15 °C Superheater outlet temperature 30/40/50 °C difference

Table 3.10. Parameters varied in IGCC study.

To perform these calculations, a number of component characteristics are fixed, and some assumptions that are made for efficiencies, temperature differences, etc., are presented in table 3.11. See also sections 3.2.1 - 3.2.8 for more information about the models used.

1 With the exception of cooled gas turbine stages. See also section 3.2.8 on the efficiency of the steam turbine. 2 Before mixing with cooling air to first stage. 3 Corresponding ISO turbine reference inlet temperatures are approximately 1090/1140°C [ISO 11086,1996]. 4 ISO conditions

43 Dryer Dried fuel outlet temperature = exhaust gas outlet dew temperature or steam saturation temperature Heat loss: 5% of dryer duty Gasifier Tgadief=900°C Carbon conversion efficiency: 0.99 Heat loss pressurised: 1% of fuel HHV Heat loss near atmospheric: 2% of fuel HHV Inert gas mass flow pressurised: 10% of dried fuel Inert gas mass flow near atmospheric: 5% of dried fuel Fuel compressor1 Polytropic efficiency: 0.80 Air compressor1 Polytropic efficiency: 0.75 Boost compressor Polytropic efficiency: 0.75 Gas turbine Maximum blade material temperature 2: Tmtrl=800°C HRSG Total pressure loss on gas side: Ap=3 kPa Superheater Maximum steam temperature: Tmax=560°C Minimum temperature difference: ATmin=30°C Pressure drop steam side: Ap/p=0.05 Evaporator Minimum temperature difference: ATmin=10°C Economiser Minimum temperature difference: ATmin=10°C Approach temperature: ATapp=10oC Water side pressure drop3: Ap/p=0.15 DH heat exchanger Minimum temperature difference: ATmin=10°C Gas cooler Gas side pressure drop: Ap/p=0.01 Other heat exchangers Min. temperature difference: ATmin=10oC (liq./liq. and liq./gas) Minimum temperature difference: ATmjn=30°C (gas/gas) Steam turbine condenser Pressure: 0.60 or 1.0 bar depending on configuration District heating water Retum/supply temperatures: 50/90°C Hot gas filter Temperature drop: 20°C Pressure drop: Ap=0.20 bar Scrubber Pressure drop: Ap=0.10 bar Gas saturated at outlet. Fuel compr. intercooler1 Gas outlet temperature: 35°C, saturated Generators Electrical efficiency: 0.99 Steam turbine Isentropic efficiency: 0.85 Mechanical efficiency: 0.99 Steam turbine gear Mechanical efficiency: 0.98 Feed water tank Pressure: 1.5 bar Pumps Overall efficiency4: 0.713

Table 3.11. Assumptions used for calculations ofIGCC systems.

1 Near atmospheric gasification. ' Used in cooling model with equation (3.11) for film cooling. 3 Total from feed water pump to steam drum. 4 fis*Tlm*T|e

44 Given these input data, it is possible to perform design point calculations by means of ASPEN PLUS™. In appendix I, examples are given of both the calculation sequence used, and the resulting process flow diagram.

The first investigation concerns the variation of electrical efficiency and fuel utilisation with the steam turbine inlet pressure and gas turbine compressor PR and COT. After that, the other parameters in table 3.11 are varied for a number of gas turbine compressor PR with the COT and live steam pressure held constant. In addition to this, the influence of final steam superheating in the raw gas cooler is investigated, as is the introduction of a second pressure level in the HRSG for some of the systems. Finally two additional systems, one with an uncooled gas turbine, and one utilising more advanced, large scale technology, are studied.

The results are given in a number of graphs enabling a comparison among the systems.

3.3.1 Pressurised gasification and exhaust gas dryer

This system consists of a pressurised gasifier and an exhaust gas dryer. A schematic picture of the plant layout is given in figure 3.14.

BC Booster compressor G Generator C Compressor GC Gas cooler com Condenser M Electrical motor DH District heating heat exchanger SH Superheater : Air and exhaust gas ECO Economiser ST Steam turbine : Biomass EV Evaporator T Turbine : Water/steam : Fuel gas

Figure 3.14. Process layout for IGCC with pressurised gasifier and exhaust gas dryer. (simplified)

45 The first calculations aim to investigate the influence of the gas turbine compressor PR and steam turbine inlet pressure on the electrical efficiency and total fuel utilisation. In figure 3.15, the electrical efficiency is shown as a function of the gas turbine compressor PR, with the steam turbine inlet pressure as a parameter. Although the curves are rather flat, there seems to be an optimum around the pressure ratio 17.5 for all four steam pressures when the gas turbine COT is 1300°C. Actually, the PR for maximum efficiency decreases somewhat with increasing steam pressure but the resulting maximum efficiency is very close to the calculated value at 17.5 regardless.

Turning to the fuel utilisation, this number decreases slightly with increasing PR. This is mainly because the flow of exhaust gas, and hence, the sensible energy, leaving the system at 60°C after the DH, increases.

Electrical efficiency (LHV) Fuel utilisation (LHV)

10 15 20 25 (PR) 25 (PR)

Figure 3.15. Electrical efficiency as a Figure 3.16. Fuel utilisation as a function function of gas turbine compressor of gas turbine compressor pressure ratio. pressure ratio.

In the exhaust gas dryer, the water evaporated from the fuel is mixed with the exhaust gas, and the resulting dew temperature is in the range of 56 - 60°C. This means that no water is condensed in the DH, even if the exhaust gas is cooled to 60°C, and hence, the latent heat in the fuel moisture cannot be recovered.

Calculations for a lower COT, 1200°C, have also been performed, and the result is compared with the results for the COT 1300°C. The live steam pressure in the presented cases is 100 bar, and as could be expected, the electrical efficiency is decreased, mainly due to the lower live steam temperature. The maximum for the efficiency is also moved towards a lower value of the gas turbine compressor PR, as can be seen in figure 3.17.

46 (%) Electrical efficiency (LHV)

-1200C

-1300C

10 15 20 25 (PR)

Figure 3.17. Electrical efficiency as a function of gas turbine compressor pressure ratio for two values of COT.

Regarding the fuel utilisation, the lower COT means greater mass flow of exhaust gas, and hence, the resulting fuel utilisation is 0.4 - 0.5 of a percentage point lower than when COT is 1300°C.

47 3.3.2 Pressurised gasification and steam dryer

Here, the wet biomass is dried by means of steam. The steam is either bled from the turbine or produced in a separate low-pressure evaporator in the HRSG. Figure 3.18 shows the configuration with an additional extraction from the gas turbine. When a second evaporator is used instead, the economiser is divided into two parts, with one part upstream from the drying steam evaporator in the exhaust gas stream and the other part downstream, utilising lower grade heat (see figure 3.11).

GASIFIER

Wet fuel

COND

Figure 3.18. Process layout for IGCC with pressurised gasifier and steam dryer (simplified, abbreviations as in figure 3.14).

First, the results for a system with separate low-pressure evaporator are presented. In figure 3.19, the electrical efficiency is given as a function of the gas turbine PR and with the steam turbine admission pressure as a parameter. The pressure of the drying steam is 9 bar, and the pressure of the conveying steam in the diyer is 4 bar, giving a temperature difference in the dryer of about 30°C.

Compared with the system with the exhaust gas dryer, the electrical efficiency is approximately 3 percentage points lower. This is mainly due to lower mass flow to the steam turbine, limited by the pinch-point in the dryer evaporator as discussed below. Also, the increased consumption of electricity in the dryer, decreases the electrical efficiency compared with the system using an exhaust gas dryer.

The fuel utilisation is given next in figure 3.20. The shape of the curve is similar to what was presented in figure 3.16, although the utilisation is significantly higher. This is mainly because

48 all of the water from the fuel is now condensed in a district heating heat exchanger, DH2. Hence, more latent heat is recovered within the system when a steam dryer is used.

(%) Electrical efficiency (LHV) (%) Fuel utilisation (LHV)

80 ba 80 ba 100 ba 100 ba 120 bcr 120 ba '140 bcr 140 ba

25 (PR) 25 (PR)

Figure 3.19. Electrical efficiency as a Figure 3.20. Fuel utilisation as a function function of gas turbine compressor of gas turbine compressor pressure ratio. pressure ratio.

With the pressure levels used here, the temperature profiles in the HRSG indicate excessive exergy losses in this component. The limiting pinch-point of 10°C is in all cases at the drying steam evaporator, giving pinch-points in the range of 54 - 127°C at the high-pressure evaporator. In an attempt to make better use of the exergy in the exhaust gas the low-pressure evaporator in figure 3.11 is omitted. The HRSG is now of a single pressure design. The steam to the dryer is extracted from the steam turbine at 9 bar and the condensate is returned to the economiser as indicated in figure 3.18 above. The resulting electrical efficiency is shown in figure 3.21.

(%) Fuel utilisation (LHV)

80 bcr 100 bo 120 ba* 140 bcr

25 (PR)

Figure 3.21. Electrical efficiency as a Figure 3.22. Fuel utilisation as a function function of gas turbine compressor of gas turbine compressor pressure ratio. pressure ratio. Drying steam bled from Drying steam bled from steam turbine. steam turbine.

It can be seen that for all but the highest gas turbine compressor pressure ratios, the electrical efficiency is higher when drying steam is bled from the turbine. The reason for this is that more of the energy in the exhaust gas is recovered at a higher temperature level (less exergy is lost). At higher pressure ratios, however, the temperature of the exhaust gas leaving the economiser increases. For example, at PR 25, this temperature is in the range of 240 - 250°C, depending on the steam pressure; while at PR 10, the corresponding figures are 165 - 172°C.

49 The resulting fuel utilisation, shown in figure 3.22, is comparable with what was shown for the system with-a separate evaporator for drying steam. Comparing figures 3.20 and 3.22, it can be seen, however, that the decline with increasing PR is less pronounced now, owing to better cooling of both the exhaust gas and the dryer condensate.

To convert more of the sensible energy in the exhaust gas into electricity, an additional low pressure evaporator can be added to the HRSG. The low-pressure steam is then injected into the steam turbine through a second admission. In figure 3.23 are shown the electrical efficiencies of the original steam cycle with evaporation at one pressure level (100 bar) only and of a system with two evaporation pressures (100/5 bar).

(%) Electrical efficiency (LHV) 1 vs 2 pressure HRSG

25 (PR)

Figure 3.23. Electrical efficiency for one and two pressure levels in the HRSG.

For gas turbine compressor PR 25, the increase in electrical efficiency is almost 1.3 percentage points. The difference between the systems then decreases with the PR since less energy is available for evaporation of low-pressure steam. It is also worth noting that the difference between the maximum efficiencies is only about 0.5 of a percentage point, with the maximum at a higher gas turbine compressor PR for the two-pressure HRSG. In practice, the rather modest increase in efficiency would probably not compensate for the increase in investment cost caused by the additional heat exchangers and the second steam admission.

The lower pressure level, 5 bar, has been chosen quite arbitrarily. A sensitivity analysis, however, indicates that there is not very much to gain by optimisation of this pressure.

Finally, the influence of the COT on the electrical efficiencies of these systems is shown in figure 3.24. It can be seen that the system with the turbine extraction is more sensitive to the COT.

50 (%) _ Electrical efficiency (LHV)

1200C

1300C

25 (PR)

Figure 3.24. Electrical efficiency as a function of gas turbine compressor pressure ratio for two values of COT.

Here a comment on the practical difference between the turbine extraction and the separate evaporator for drying steam may be appropriate. The flow bled from the steam turbine here is a significant part of the total flow through the turbine, and hence, this extraction would probably disturb the flow and decrease the turbine efficiency, unless it is placed between two separate turbine modules. On the other hand, the flow in the high pressure part of the steam turbine is now increased compared with the case without extraction, and that is beneficial for the efficiency of the first high pressure stage. This is especially important for small turbines, where the first stage blades tend to be very short. However, these effects have not been taken into account in the calculations presented above.

51 3.3.3 Near atmospheric gasification and exhaust gas dryer

Here, the gasification is performed at near atmospheric pressure. This means that there is less of a problem introducing the fuel into the gasifier, but also that the produced gas has to be compressed before it is delivered to the gas turbine combustor. The system is depicted in figure 3.25 below.

Air pre­ heater

COND

SC: Scrubber RH: Clean gas recuperator

Figure 3.25. Process layout for IGCC with near atmospheric gasifier and exhaust gas dryer, (simplified, abbreviations same as in figure 3.14)

Before the fuel gas is compressed, it has to be cooled in order to decrease the compression work. Hence, cold gas cleaning is performed. Leaving the gasifier, the gas is first cooled in an evaporator, similar to the one in the pressurised system, and then in a recuperator heating the clean compressed gas. In a third heat exchanger, heat is transferred to the gasification air1, and then the gas enters a wet scrubber. Here it is assumed that the gas leaves the scrubber at 35°C, saturated with water. The gas is then compressed in a three-stage intercooled compressor, as described in section 3.2.6, then heated and delivered to the gas turbine combustor.

1 The feasibility of this heat exchanger could be questioned. With a direct heat exchanger with air on one side and fuel gas on the other, there may be a risk of fire or explosion in the case of a leakage. A safer, but probably also more expensive, solution is to use water as an intermediate medium as depicted in figure 3.25.

52 Both in the scrubber and in the intercoolers, water in the fuel gas is condensed and withdrawn. Hence, the heating value of the compressed gas is slightly higher here than in the case with pressurised gasification. The lower water content also gives less water in the exhaust gas, thereby lowering the dew temperature to 53 - 58°C.

The electrical efficiencies obtainable with this system, with a gas turbine combustor outlet temperature of 1300°C, are presented in figure 3.26. For all steam pressures, the maximum electrical efficiency is achieved at a PR close to 17.5.

(%) Electrical efficiency (LHV) (%) Fuel utilisation (LHV)

■80 ba •80 ba 100 ba 100 ba 120 bcr 120 bo 140 bo

25 (PR) 25 (RR)

Figure 3.26. Electrical efficiency as a Figure 3.27. Fuel utilisation as a function function of gas turbine compressor of gas turbine compressor pressure ratio. pressure ratio.

Compared with the pressurised gasification, the electrical efficiency is 5 percentage points lower. The main reason for this, is the electricity requirement of the fuel gas compressor, amounting to 13 - 18% of the gross electric output.

The fuel utilisation is then presented in figure 3.27. Also in this system, the fuel utilisation remains virtually constant, since the sensible energy in the exhaust gas not recovered into the steam cycle, is used for heating water instead. The low dew temperature of the exhaust gas means that no condensation occurs in the DH and the fuel utilisation is now 7 percentage points lower, compared with the pressurised system. The lower fuel utilisation is due to the higher heat loss from the gasifier and the additional losses in the scrubber and fuel compressor intercoolers, where heat is rejected to scrubbing/cooling water.

Again, the slight decrease in fuel utilisation with increasing pressure ratio follows from the increased mass flow of the exhaust gas. In addition to this, the electrical and mechanical losses in the fuel compressor, as well as the heat losses in the intercoolers, increase.

The influence of the COT, as presented in figure 3.28, is very similar to the result for the system with pressurised gasification (figure 3.17). When the combustor outlet temperature is 1200°C, the PR for maximum efficiency is shifted towards lower values. The maximum efficiency is decreased by about 0.4 of a percentage point.

53 (%) _ Electrical efficiency (LHV)

1200C

1300C

25 (PR)

Figure 3.28. Electrical efficiency as a function of gas turbine compressor pressure ratio for two values of COT.

54 3.3.4 Near atmospheric gasification and steam dryer

This system consists of an atmospheric gasifier and a steam dryer. A schematic picture of the plant layout is given in figure 3.29.

Air pre­ heater 0—O

Gasifier COND

Figure 3.29. Process layout for IGCC with near atmospheric gasifier and steam dryer, (simplified, abbreviations as in figure 3.25)

In this system steam for the dryer is extracted from the turbine at 9 bar. Hence, there is no separate evaporator for drying steam, making this a true one-pressure HRSG. As such, it has a rather high exhaust gas temperature out from the economiser, especially since the feed water entering the economiser is already preheated in the fuel compressor intercooler. Hence, there is a surplus of low temperature heat, which is difficult to recover into the cycle.

Figure 3.30 shows the electrical efficiency for this system when varying the gas turbine PR and the live steam pressure. It can be seen that the highest efficiencies are achieved at rather low pressure ratios. The main reason for this is better cooling of the exhaust gas in the economiser. At PR 10 the exhaust gas leaves the economiser at 151 - 163°C; whereas the corresponding temperatures at PR 25 are 232 - 240°C.

55 (%) Electrical efficiency (LHV) (%) Fuel utilisation (LHV)

■80 be 80 ber 100 ber 100 ber 120 ber 140 ber 140 ber

25 (PR) 25 (PR)

Figure 3.30. Electrical efficiency as a Figure 3.31. Fuel utilisation as a function function of gas turbine compressor of gas turbine compressor pressure ratio. pressure ratio.

Compared with the system with the exhaust gas dryer, the maximum electrical efficiency is slightly lower (0.5 - 1.0 percentage point), and the compressor pressure ratio for maximum efficiency is shifted towards lower values.

The fuel utilisation, figure 3.31, is virtually insensitive to the pressure level in the steam cycle, but decreases somewhat with increasing gas turbine PR as explained earlier. Again, the steam dryer enables total recovery of the latent heat from the evaporated fuel moisture.

To convert more of the sensible energy in the exhaust gas to electricity, particularly for the higher pressure ratios, a second pressure level can be introduced in the HRSG. This requires the addition of a low-pressure economiser and evaporator in the HRSG and a second steam admission on the steam turbine. Figure 3.32 shows the resulting electrical efficiencies for steam pressures 100/5 bar. It can be seen that the gain in efficiency is greatest for the highest pressure ratios. As in the case of the pressurised gasification, the difference is, however, modest, the increase in maximum electrical efficiency being about 0.3 of a percentage point.

(%) Electrical efficiency (LHV) 1 vs 2 pressure HRSG

38 -' -

37 - -

Figure 3.32. Electrical efficiencies for systems with evaporation at one and two pressure levels.

56 As can be seen in figure 3.33, the influence of the COT on the electrical efficiency follows the trend of the earlier systems. When the COT is lowered, the maximum efficiency decreases 0.8 of a percentage point.

(%) Electrical efficiency (LHV)

1200C

1300C

25 (PR)

Figure 3.33. Electrical efficiency as a function of gas turbine compressor pressure ratio for two values of COT.

This concludes the initial investigation of IGCC systems. Next, the influence of the remaining parameters listed in table 3.10 will be treated.

57 3.3.5 Ambient temperature design value

The value of ambient temperature used in sections 3.3.1 - 3.3.4 above, 15°C, is part of the ISO conditions often used in thermodynamic calculations of gas turbine power plant systems. Using this value then facilitates comparison with results from other investigations. When calculating the thermodynamic performance of cogeneration plants located in Sweden, a lower value of design ambient temperature should normally be used, since these plants are run mainly when the outside temperature is low and there is a demand for district heating. Hence, the influence of lower ambient temperature on design point performance has been investigated for the systems described above.

In these calculations, it has been assumed that only the temperatures of the air and of the fuel fed to the process is changed. The temperature levels in the district heating network are kept constant, as are the temperatures of any other external cooling flows utilised in the process. It is further assumed that the COT and the compressor pressure ratio are kept at design values and that the polytropic efficiencies of the gas turbine compressor and expander are not changed. The fraction of the compressor flow used for cooling, however, is assumed to vary according equation (3.17), used for example by Palmer and Erbes (1994).

(3.17)

The result is a slight increase in mass flow of cooling air for lower ambient temperature.

Other pieces of equipment, such as the steam turbine and heat exchangers, are varied to fit each particular case.

In figures 3.34 and 3.35, the electrical efficiency and the fuel utilisation, respectively, for four IGCC systems are shown for three design values of ambient temperature. To identify the systems, the following notation is used:

1. IGCC with pressurised gasifier and exhaust gas dryer. 2. IGCC with pressurised gasifier and steam dryer1. 3. IGCC with near atmospheric gasifier and exhaust gas dryer. 4. IGCC with near atmospheric gasifier and steam dryer1.

It can be seen that the electrical efficiency generally decreases with ambient temperature. The lower air temperature results in increased gas turbine efficiency, due to the decreased work absorbed by the gas turbine compressor. This is, however, counteracted by a simultaneous decrease in steam turbine output, because useful heat is taken from the steam cycle to cover the increased demand in the dryer. When the temperature of the raw fuel is below 0°C, the water content is present in the form of ice, and an additional amount of heat corresponding to the latent heat of fusion is required in the dryer. Hence, the decrease in steam cycle output is more marked when the ambient temperature is -5°C.

1 Heating steam bled from steam turbine.

58 For systems with exhaust gas dryer the PR for maximum electric efficiency increases when the ambient temperature decreases. This trend is not seen for the systems with steam dryer.

(%) Electrical efficiency (LHV)

25 (PR)

Figure 3.34. Electrical efficiency of studied IGCC systems, with design value of ambient temperature (°C) as a parameter.

Turning to the fuel utilisation, the four systems behave very similarly, as can be seen in figure 3.37. When the temperature of the fuel fed to the dryer is decreased, more sensible energy must be supplied to heat the material, and when the temperature is below 0°C, the latent heat of fusion for the moisture also has to be added. Using 15°C as a base case, this results in a decrease in fuel utilisation of approximately 1 percentage point when the ambient temperature is 5°C. For -5°C, the decrease is almost 4 percentage points.

(%) Fuel utilisation (LHV)

25 (PR)

Figure 3.35. Fuel utilisation of studied IGCC systems, with design value of ambient temperature (°C) as a parameter.

Once these general trends have been evaluated, and site-specific values of the design operating temperature and temperature range have been defined, equipment specifications and performance should be fixed, and the performance over the operating temperature range be re­ evaluated. This re-evaluation is beyond the scope of the present investigation, and thus, is not reported here.

59 3.3.6 Fuel moisture content

As was stated above, the moisture content of fresh woody biomass is usually in the range of 40 - 60%, calling for some drying before feeding to the gasifier. In the following discussion, the influence of fuel moisture content, before and after drying, on electrical efficiency and fuel utilisation will be examined. The moisture content of the fuel delivered to the dryer is varied in the range of 30 - 60% and the moisture content after drying between 10 and 20%.

For this study, a gas turbine with compressor PR 15 and COT 1300°C is used. The steam turbine admission pressure is 100 bar, and the pressure of heating steam, in the case of the steam dryer, is 9 bar. Other assumptions are as in the base case.

Varying the moisture content of the fuel fed to the gasifier, requires that some assumption is made about how this influences the resulting gas composition. Here, the following approach, similar to what was presented by Hughes and Larson (1997), is used.

In the base case, the moisture content of the fuel fed to the gasifier is 15%. When varying the moisture content, the flow of dry substance is fixed. The air flow to the gasifier is then varied to achieve the pre-set operating temperature, and the gas composition is calculated based on the same equilibrium assumptions as were used earlier for 15% moisture (section 3.2.2). Resulting gas compositions (mole fractions) and LHV for the clean gas delivered to the gas turbine combustor1 are presented in table 3.12. It should be noted that also other gas compositions, closing the mass and energy balances, are possible. Earlier investigations [Hughes and Larson, 1997; Newby et al., 1998], however, indicate that the exact gas composition has limited impact on the calculated plant thermal performance.

Pressurised gasification Near atmospheric gasification Moisture content 10% 15% 20% 10% 15% 20% C02 0.1234 0.1289 0.1326 0.1303 0.1366 0.1419 Ar 0.0044 0.0045 0.0045 0.0051 0.0053 0.0054 CO 0.1788 0.1557 0.1353 0.2056 0.1876 0.1711 n 2 0.4056 0.4085 0.4114 0.4465 0.4592 0.4729 H2 0.1442 0.1518 0.1564 0.1268 0.1367 0.1448 CH4 0.0668 0.0538 0.0419 0.0616 0.0542 0.0464 h 2o 0.0764 0.0965 0.1176 0.0076 0.0076 0.0076 c2h 6 - - - 0.0139 0.0100 0.0069

LHV (MJ/kg) 5.631 5.048 4.496 6.106 5.580 5.070

Table 3.12. Composition (mole fractions) and lower heating value of clean gas when moisture content of fuel fed to the gasifier is varied. Some minor constituents, like NH3, are not presented, causing the sum of fractions to differ from 1.0.

Figures 3.36a - e show how the electrical efficiency varies with the moisture content of the fuel fed to the dryer. The parameter in the graphs is the moisture content after drying, before

1 For near atmospheric gasification, the major part of the H20 present in the gas leaving the gasifier is condensed and withdrawn in the scrubber and in the fuel compressor intercoolers.

60 feeding to the gasifier. The graphs give the electrical efficiency based on the higher heating value of the biomass delivered to the dryer. The reason for this is that the lower heating value varies considerably with the fuel moisture content, thereby altering the basis for the efficiency calculation.

(%) Electrical efficiency (HHV) (%) Electrical efficiency (HHV)

Figure 3.36a. Pressurised gasification and Figure 3.36b. Near atmospheric exhaust gas dryer. gasification and exhaust gas dryer.

Electrical efficiency (HHV) (%) Electrical efficiency (HHV)

i'.-w

f-:...•*- “ 30 40 50 60 (%) 60 (%)

Figure 3.36c. Pressurised gasification and Figure 3.36d. Near atmospheric steam dryer fed from turbine extraction. gasification and steam dryer fed from turbine extraction.

(%) Electrical efficiency(HHV)

60 (%)

Figure 3.36e. Pressurised gasification and steam dryer fed from separate evaporator.

61 Starting with the moisture content after the dryer, a general trend in all studied systems is that a low moisture "content is advantageous in terms of electrical efficiency. This is as expected, since the dryer uses low-grade heat to evaporate water; while in the gasifier, chemical energy in the fuel is used to evaporate any remaining water.

For a constant flow of dry substance to the gasifier, a higher moisture content means a greater total mass flow of fuel to the gasifier. More moisture also means more air must be fed to the gasifier to reach the operating temperature, and the result is an increase in gas flow from the gasifier. For a near atmospheric gasifier, more compression work is then required, and this is why the decrease in efficiency with increasing moisture content in the dried fuel is more pronounced for the near atmospheric gasifiers.

Turning to the moisture content of the raw fuel fed to the dryer, there is a marked difference between the two types of dryers.

Looking at the systems utilising exhaust gas dryers, there is a certain moisture content above which useful heat is taken from the steam cycle and used in the dryer instead. In figures 3.36a and b, this limit is between 40 and 50 % moisture. At lower initial moisture contents, a surplus of low temperature heat exists. Due to the low temperature, this energy cannot be converted into electricity in the steam cycle, and the electrical efficiency remains constant. The low temperature heat is then used for the heating of district heating water instead. Another possibility could be to include a second pressure level in the HRSG to recover some of this heat into the steam cycle.

For the systems with steam dryers, figures 3.36c, d and e, the efficiency increases with decreasing moisture content in the entire interval studied. Comparing figures 3.36c and d, the system with steam bled from the turbine yields higher efficiency at high moisture contents. At lower moisture content, this difference decreases.

When drying steam is bled from the steam turbine, it is obvious that the steam used for drying, the amount of which increases with the raw fuel moisture content, could have been used to produce more power in the steam turbine instead.

When steam for the dryer is instead produced in a separate evaporator, the production of high- pressure steam to the turbine is limited by the heat requirement of the drying steam evaporator. The resulting pinch-point at the high-pressure evaporator varies between 25°C and 155°C, when the fuel fed to the dryer has a moisture content of 30% and 60% respectively. This means the exergy losses in the HRSG are excessive for higher moisture contents, resulting in a low electrical efficiency. The decrease in pinch-point at the high-pressure evaporator with decreasing moisture content, indicates that the electrical efficiency for this system also levels off at a certain moisture content. Then this pinch-point limits the production of steam to the turbine, as in the systems with an exhaust gas dryer, and there is a surplus of low temperature heat.

Next, in figures 3.37a - e, the fuel utilisation for these systems is given, again based on the higher heating value.

62 (%) Fuel utilisation (HHV) (%) Fuel utilisation (HHV)

Figure 3.37a. Pressurised gasification and Figure 3.37b. Near atmospheric exhaust gas dryer. gasification and exhaust gas dryer.

(%) Fuel utilisation (HHV) (%) Fuel utilisation (HHV)

i— * q /til L 69 - . 74 - 72 • 70 68 - 66 - 30 40 50 60 (%) 30 40 50 60 (%)

Figure 3.37c. Pressurised gasification and Figure 3.37d. Near atmospheric steam dryer fed from turbine extraction. gasification and steam dryer fed from turbine extraction.

(%) Fuel utilisation (HHV)

30 40 50 60 (%)

Figure 3.37 e. Pressurised gasification and steam dryer fed from separate evaporator.

For the systems with exhaust gas dryers, figures 3.37a and b, the fuel utilisation decreases significantly with the raw fuel moisture content, while the influence of the moisture content after the dryer is very limited. The only exception from this trend is when very moist material is fed to the dryer in a system with pressurised gasification, figure 3.37a. At moisture contents above 50%, some of the water in the exhaust gas leaving the dryer is condensed in the district heating heat exchanger heating the cold return water. Hence, the slope of the curve is changed. It can also be seen that the higher the moisture content after the dryer, the higher the fuel utilisation. The main reason for this is the smaller flue gas flow leaving the system, since more

63 water into the gasifier results in a lower fuel gas heating value, and hence, less combustion air to reach the same COT.

With near atmospheric gasification, the dew point of the exhaust gas is lower, since some water is withdrawn in the gas cleaning and no condensation occurs even at 60% fuel moisture. The extraction of water in the scrubber and fuel gas compressor intercoolers, is also the reason why fuel utilisation decreases slightly with increasing moisture content after the dryer, figure 3.37b.

The systems utilising steam dryers show fuel utilisations virtually independent of raw fuel moisture. Here, the heat used to evaporate the fuel moisture is completely recovered (with the exception of heat losses) in a condenser heating the district heating water.

Any water left in the fuel fed to the gasifier, on the other hand, is withdrawn in the fuel gas compressor intercoolers (in the case of near atmospheric gasification) or is released from the system with the flue gas. This results in a loss of latent heat, as can be seen in figures 3.37c - e, where the fuel utilisation decreases with increasing moisture content after the dryer.

3.3.7 Heat loss between dryer and gasifier

In the base case investigation above, it was assumed that the dried fuel was cooled to ambient temperature (15°C) before feeding to the gasifier. This cooling constitutes a loss of sensible energy. Leaving aside practical considerations, such as, heat losses during transportation from dryer to gasifier and the possible need for an intermediate storage of dried fuel, a theoretical case can be calculated setting the heat loss to zero. The resulting theoretical gains in electrical efficiency are presented in figure 3.38.

(%) Electrical efficiency (LHV)

Press, gasif. Press, gasif. Atm. gasif. Atm. gasif. Exh.gas dryer Steam dryer Exh.gas dryer Steam dryer

Figure 3.38. Influence of heat loss from fuel between dryer and gasifier.

Since the fuel is at a significantly higher temperature when leaving the steam dryer (steam saturation temperature at 4 bar) than when leaving the exhaust gas dryer (exhaust gas dew temperature), the theoretical gain is also greatest for systems with a steam dryer. Naturally this holds also for the fuel utilisation, figure 3.39.

64 These figures present a substantial potential for increased efficiency if this heat loss is eliminated. Hovv much of this gain can be utilised in practice, calls for more detailed investigations.

Press, gasif. Press, gasif. Atm. gasif. Atm. gasif. Exh.gasdryer Steam dryer Exh.gas dryer Steam dryer

Figure 3.39. Influence of heat loss from fuel between dryer and gasifier on fuel utilisation.

3.3.8 Temperature and pressure of fuel gas

Turning to the influence of temperature and pressure of the gas delivered to the gas turbine fuel control valve, it must be noted that the base case values used above may be somewhat on the conservative side. Hence, the influence of raising the gas temperature from 380°C to 530°C, as well as lowering the pressure ratio over the fuel feeding system from 1.4 to 1.1, has been investigated.

Starting with the gas temperature, the idea is that more of the sensible energy should be utilised in the gas turbine combustor instead of being recovered in the steam cycle only. In practice, the gas temperature could be limited by the ability of the fuel gas feed system to withstand the higher temperature. Another constraint is the gas cleaning requirement, which for hot gas cleaning means the gas has to be sufficiently cooled to condense the alkali metals before the filter. Newby et al. (1998), Salo et al. (1998) and Liinanki and Karlsson (1994) indicate that cooling to below 600°C should be sufficient for hot gas filtering of the alkali species, and Newby et al. (1998), Salo et al. (1998) and Kurkela and Solantausta (1994) also claim the fuel gas feeding and control system after some modification can withstand that temperature. Hence, the choice made here seems plausible and has been used for both pressurised and near atmospheric gasification, using clean gas reheat in the latter case.

The results are given in figures 3.40a and b for pressurised and near atmospheric gasification, respectively. The gas turbine has a COT of 1300°C, and the live steam pressure is 100 bar. The steam dryer is fed with 9 bar steam.

65 (%) Electrical efficiency (LHV) (%) Electrical efficiency (LHV)

■530C

10 15 20 25 (PR) 25 (PR)

Figure 3.40a. Effect of fuel gas delivery Figure 3.40b. Effect of fuel gas delivery temperature on electrical efficiency of temperature on electrical efficiency of IGCC with pressurised gasification. IGCC with near atmospheric gasification.

In figure 3.40a, it is shown that the maximum electrical efficiency is raised by approximately 0.4 of a percentage point for the system with an exhaust gas dryer. The maximum is also shifted towards a lower PR due to changed temperature profiles in the HRSG (increased exhaust gas flow and evaporation duty).

The steam dryer here is fed from a separate evaporator, and increasing the fuel gas temperature results in an increase in electrical efficiency of 0.6 of a percentage point. Here the exergy loss in the HRSG is decreased, due to the increased evaporation duty of the high- pressure evaporator. The limiting pinch-point is, however, still at the drying steam evaporator (compare with section 3.3.2).

For near atmospheric gasification, the increase in fuel gas temperature results in 0.6 of a percentage point higher efficiency when an exhaust gas dryer is used.

When a steam dryer, this time utilising heating steam bled from the turbine, is used, the result is, however, completely different. Here, an increase in efficiency is achieved only for the pressure ratio 7.5. At this PR, the pinch-point at the high-pressure evaporator is decreased when the fuel gas temperature is increased, lowering the exergy losses in the HRSG without raising the exhaust gas temperature leaving the economiser. For the other pressure ratios, the output from the gas turbine is also increased but this is counteracted by an increase in both exhaust gas flow and temperature after the economiser in the HRSG, with the pinch-point remaining at the limiting value of 10°C. Taking full advantage of the increased fuel gas temperature then would require a second pressure level in the HRSG.

Considering the fuel utilisation, the values for the higher fuel gas temperature are in all cases 0.2 - 0.4 of a percentage point lower than the fuel utilisation values with the original gas temperature. The main reason is the increased loss of sensible energy to the surroundings, due to the increased flow of the flue gas leaving the system after the district heating heat exchanger in the HRSG.

In order to reduce the work required for the compression of gasification air or fuel gas, the pressure drop over the gas turbine fuel valve should be as small as possible. In the base case, it has been assumed here that the pressure before the fuel valve is 1.4 times the pressure in the combustion chamber. This is similar to what is used at the Vamamo demonstration plant [Stahl, 1997]. However, it seems possible to reduce this value [Neergaard, 1998], and

66 recalculating with a pressure ratio (PRvaive) of 1.1, results in the electrical efficiencies presented in figure 3.41 below.

(%) Electrical efficiency (LHV)

Press, gasif. Press, gasif. Atm. gasif. Atm. gasif. Exh.gas dryer Steam dryer Exh.gas dryer Steam dryer

Figure 3.41. Influence of fuel gas delivery pressure on electrical efficiency.

The increase in efficiency for the systems with near atmospheric gasification was 0.5 - 0.6 of a percentage point due to decreased compression work in the fuel gas compressor.

For the pressurised systems, no efficiency improvement resulted from the decrease in the fuel valve pressure ratio, although the electricity requirement of the boost compressor decreased by 50%. The reason for this is that the lower boost compressor pressure ratio resulted in lower air temperature in to the gasifier, since no recuperator was included, and consequently, a larger mass flow of air was required. To avoid this effect, the system with pressurised gasification could be modified through incorporation of a recuperative heat exchanger, keeping the air temperature constant when the fuel delivery pressure is decreased. The pressurised systems with an exhaust gas dryer then exhibits an increase in electrical efficiency of approximately 0.1 of a percentage point.

3.3.9 Pressure in steam dryer

For systems with a steam dryer, the pressure of the drying steam as well as of the conveying steam, are adjustable variables. In the base case calculations, the drying steam pressure was 9 bar and the conveying steam pressure was 4 bar. Here, the thermodynamic results of lowering these values to 6 and 2 bar, respectively, are presented for three systems. In the first system, drying steam is produced in a separate evaporator in the HRSG, and the gasifier is pressurised (section 3.3.2). In the second system, drying steam is instead extracted from the steam turbine (section 3.3.2). The third system also utilises steam from the turbine, and the gasifier works at near atmospheric pressure (section 3.3.4). Figure 3.42 shows that the lower pressure level results in 0.5 of a percentage point higher electrical efficiency for all systems.

A common feature for the three systems is that the temperatures of the fuel and of the evaporated moisture leaving the dryer decrease when the pressure of the conveying steam is decreased from 4 to 2 bar. As a consequence of this, the energy, and hence the amount of heating steam, required for drying, is also decreased.

67 □ 9/4 bar ■ 6/2 bar

Press, gasif. Press, gasif. Atm. gasif. Separate evap. Turbine extr. Turbine extr.

Figure 3.42. Electrical efficiency for systems with different steam pressure levels in the dryer.

Lowering the pressure of the drying steam in the case of a separate evaporator, means that heat for the dryer is taken from the exhaust gas at a lower temperature. Hence, more sensible energy at a high temperature is available for production of high-pressure steam in the HRSG, and the steam turbine output increases approximately 4 percent.

Also, when drying steam is bled from the turbine, the lower pressure level is advantageous. Firstly, the steam produces additional work during the expansion from 9 to 6 bar. The result is 3 percent higher output from the steam turbine. The colder condensate returned from the dryer and mixed with the feed water also enables better cooling of the exhaust gas in the HRSG economiser.

The fuel utilisation is also slightly improved, approximately 0.3 of a percentage point, as shown in figure 3.43. The main reason for this is the lower fuel temperature at the dryer outlet, decreasing the loss of sensible energy when cooled to the ambient temperature.

(%) Fuel utilisation (LHV)

Press, gasif. Press, gasif. Atm. gasif. Separate evap. Turbine extr. Turbine extr.

Figure 3.43. Fuel utilisation for systems with different steam pressure levels in the dryer.

Changing from 9 to 6 bar pressure of the drying steam with a corresponding decrease in pressure of the conveying steam from 4 to 2 bar means the driving force in terms of

68 temperature difference is slightly increased, implying the heat transfer area for the same duty would decrease. Even if higher volume flows resulting from the lower pressures presumably mean bulkier equipment, the specific investment cost is believed to be about the same for the two dryers. Hence, the increase in efficiency presented here would also result in better economy. This can, however, be very case specific and should be investigated more thoroughly before any decision be made.

Wimmerstedt and Linde (1998) showed that lowering the pressure level of the heating steam without a corresponding decrease in pressure of the conveying steam, would have resulted in an increase of heat transfer area and cost, for the same duty.

3.3.10 Temperature differences in heat recovery steam generator

In earlier sections, values were chosen for the temperature differences in the HRSG, based on current practice in conventional combined cycle boilers. To increase the thermal efficiency of the plant, these temperature differences could be reduced. This, however, results in increased heat transfer area, and hence, investment cost, i.e. it is a matter of economical optimisation.

Here, only the influence on the electrical efficiency of changing the pinch-point in the evaporators (ATPP) and of the minimum temperature difference in the superheater (ATsh) will be shown. This information could then be combined with a calculation of the change in heat transfer area and used as input in economical calculations.

In figures 3.44 and 3.45, the results for three systems are shown. The systems were chosen somewhat arbitrarily among those where the studied temperature differences influenced the efficiency. The numbers in the graphs refer to the following systems with the COT 1300°C, the live steam pressure 100 bar and certain values of the compressor PR.

1. Pressurised gasification and exhaust gas dryer, PR=25 2. Pressurised gasification and steam dryer (separate evaporator), PR=17.5 3. Pressurised gasification and steam dryer (turbine extraction), PR=12.5

(%) Electrical efficiency (LHV) (%) Electrical efficiency (LHV) 45

44

43 42 .i* 41 • ••• 40 -?--- 0 5 10 15 (C) 20 30 40 50 60 (C)

Figure 3.44. Influence of evaporator Figure 3.45. Influence of final temperature pinch-point on electrical efficiency. difference in superheater on electrical efficiency.

It can be seen that the evaporator pinch-point does influence the efficiency, whereas the final steam temperature out of the superheater is less important.

69 Not shown here, there are also cases where lowering the pinch-point does not increase efficiency because steam production is limited by the heat requirement of the dryer, rather than the evaporator pinch-point.

It should be noted that for lower gas turbine compressor pressure ratios, the live steam temperature is limited (in these calculations to 560°C) by material constraints rather than by the minimum temperature difference in the superheater. This can, for example, be seen for system 3 in figure 3.45, where the maximum live steam temperature is reached at ATsh=40°C.

3.3.11 Superheating in raw gas cooler

In the base case, the heat recovered in the gas cooler is used only for evaporating water, in order to limit the heat exchanger material temperature. Since heat is available at a temperature as high as 900°C, this means the exergy loss in the gas cooler is high. To reduce this loss, heat should be recovered at a higher temperature level, for example, by steam superheating. The superheater part of the gas cooler then becomes a gas/gas heat exchanger with a material temperature higher than the corresponding evaporator, calling for more heat resistant materials.

Here, the possible gain of superheating in the gas cooler is quantified, assuming proper materials can be found. This information could then be used to evaluate the economic result of such a process change, comparing the increased income resulting from higher efficiency with the increased cost of the heat exchanger. This comparison will, however, not be conducted here.

When some of the heat in the fuel gas is used for superheating, less energy is available for evaporation in the gas cooler. Hence, the duty of the evaporator in the HRSG will increase if the steam flow is not decreased. This implies that systems with a pinch-point at the high- pressure evaporator, higher than the limiting value, would be most interesting in this context.

The original parametric study revealed that the systems with exhaust gas dryers, for both near atmospheric and pressurised gasification, had quite large pinch-points, especially for the lowest gas turbine compressor pressure ratios. Also for the system with pressurised gasification and a steam dryer fed from a separate evaporator, the pinch-point at the high- pressure evaporator is high, the limiting temperature difference being at the evaporator producing steam for the dryer. Based on this, the influence of superheating in the gas cooler is investigated for the systems with pressurised gasification in combination with exhaust gas dryingor steam drying with a separate evaporator.

The calculations are performed over a range of gas turbine compressor pressure ratios. The COT is 1300°C, and the final steam temperature is 560°C. Since there is no thermodynamic advantage of shifting superheater duty from the HRSG to a gas cooler, the superheater in the HRSG is retained, and only final heating to 560°C is performed in the gas cooler. Transferring superheated steam from the HRSG to the gas cooler for further superheating before the steam turbine, most probably increases the pressure loss between the evaporator and the turbine.

70 Here this effect is, however, neglected 1 here, and both the evaporation pressure and the turbine admission pressure are kept at the same values as in the base case.

Figure 3.46 shows the results for the system with an exhaust gas dryer, for two live steam pressure levels. It can be seen that the efficiency gain from superheating in the gas cooler, denoted ”sh”, is modest, around 0.2 - 0.3 of a percentage point. At PR 10 the gas turbine outlet temperature is 640°C, making it possible to reach 560°C steam temperature in the HRSG. Hence, additional superheating in the gas cooler is not of interest in this case.

(%) Electrical efficiency (LHV)

80 bar

•80 barsh

140 bar

' "X—140 barsh

25 (PR)

Figure 3.46. Influence of final superheating in raw gas cooler (sh) on electrical efficiency of IGCC with exhaust gas dryer and pressurised gasification.

Figure 3.47 presents the corresponding graphs for a system with a steam dryer fed from a separate evaporator. Here, the efficiency increase is greater, and especially for the higher pressure ratios, additional superheating seems advantageous.

(%) Electrical efficiency (LHV)

•80 bar

80 barsh

140 bar

140 bar sh

25 (PR)

Figure 3.47. Influence of final superheating in raw gas cooler (sh) on electrical efficiency of IGCC with steam dryer fed from a separate evaporator and pressurised gasification.

1 A control calculation reveals that if the additional pressure drop is assumed to be 5%, the resulting decrease in electrical efficiency is < 0.05 of a percentage point.

71 3.3.12 Uncooled gas turbine

In many investigations, e.g. Larson et al. (1998), Salo et al. (1998) and Pruschek et al. (1998), it is assumed that present-day gas turbines, with advanced cooling techniques and the COT similar to or exceeding that which has been used above, can be utilised in IGCC systems for biomass. If this proves to be unfeasible, for example, due to problems in developing the required gas cleaning equipment, the use of an uncooled gas turbine might be a possibility. This calls for a significant decrease in the COT, causing a corresponding decrease in gas turbine efficiency.

For a system with pressurised gasification and an exhaust gas dryer, the use of an uncooled gas turbine with the COT 850°C has been investigated. Figure 3.48 shows the resulting electrical efficiencies as a function of the gas turbine compressor PR, with the live steam pressure as a parameter.

(%) Electrical efficiency (LHV)

40 bar 36--- 60 bar 80 bar 100 bar ' X' ' 120 bar 140 bar

14 (PR)

Figure 3.48. Electrical efficiency of IGCC with pressurised gasifier and exhaust gas dryer when COT is 850°C.

Compared to the corresponding system with the COT 1300°C, the efficiency is decreased by approximately 8 percentage points, and the PR for maximum efficiency is shifted towards a lower value. Furthermore, the advantage of a higher live steam pressure is now more limited, as can be seen from a comparison with figure 3.15.

The fuel utilisation, figure 3.49, decreases slightly with increasing PR, but is virtually independent of the live steam pressure. This is in accordance with what was shown in figure 3.16 for the system with the COT 1300°C. The absolute value of the fuel utilisation is, however, roughly 1 percentage point lower here.

72 (%) - Fuel utilisation (LHV)

•40 bar ’60 bar ‘80 bar -100 bar ■120 bar ‘140 bar

14 (PR)

Figure 3.49. Fuel utilisation ofIGCC with pressurised gasifier and exhaust gas dryer when COT is 850°C.

Due to the low COT, the exhaust temperature of this gas turbine is also relatively low, resulting in low steam temperatures and correspondingly high moisture contents in the low- pressure part of the steam turbine. Hence, the possibility of superheating in the gas cooler could constitute an interesting possibility to increase the efficiency of this system. The result of such a process alteration is shown in figure 3.50, where the live steam temperature is 560°C.

(%) Electrical efficiency (LHV)

♦ 40 bar —■—40 bar sh A—100 bar X—100 bar sh X 140 bar # -140 bar sh

14 (PR)

Figure 3.50. Comparison of electrical efficiencies for systems with (sh) and without final steam superheating in the raw gas cooler for COT 850°C.

It is now possible to utilise higher pressure ratios without exceeding 13% wetness at the steam turbine outlet. The efficiency increase due to additional superheating in the gas cooler amounts to 0.9 of a percentage point for the lowest live steam pressure, 40 bar. Also, for the live steam pressure 140 bar, the efficiency increases, but only 0.2 percentage points.

As in the corresponding investigation for the COT 1300°C, the additional pressure drop caused by the second superheater has been neglected, resulting in a slight overestimation of the potential efficiency increase.

73 3.3.13 Advanced gas turbine

As a comparison to the more conservative view taken above, a case will also be calculated where full advantage is taken of possible upcoming improvements in gas turbine technology. This can be seen as an estimation of applying the upper limit of possible technology for the next generation of combined cycles, utilised in large-scale biomass IGCC. Large-scale in this context typically means 300 - 400 MWe.

In Pruschek et al. (1998) the potential for improvements in a coal-based IGCC was investigated, and as a future option, a gas turbine with the following data was used:

ISO turbine reference inlet temperature: 1400°C Compressor pressure ratio: 30 Exhaust gas temperature: 633°C

Since the turbine inlet temperature is presented according to the ISO definition [ISO 11086, 1996], a simple uncooled model can be used. An isentropic turbine efficiency of 0.87 fits the model according to the data given. Here, this gas turbine will be used in a large-scale system where also some of the improvements found in sections 3.3.7 - 3.3.11 will be included. Again, the system will be based on pressurised gasification and exhaust gas drying, but now some assumptions differ from the base case used in the parameter study above. The new parameters are presented in table 3.13, where some of the improvements require technological development, and others are results from the larger scale.

Parameter New value Live steam pressure 180 bar Live steam temperature 600°C Steam turbine isentropic efficiency1 0.89 Fuel temperature in to gasifier 25°C Air temperature in to gasifier 400°C Inert gas mass flow 5% of dried fuel Fuel gas delivery temperature 530°C Fuel gas delivery system pressure ratio 1.1

Table 3.13. New parameters adopted for a large scale IGCC plant utilising advanced gas turbine technology.

Due to the larger scale, the steam turbine does not require a gear and this improves efficiency, as does a recuperative heat exchanger hearing the air after the boost compressorto 400°C, by cooling the air delivered from the gas turbine compressor.

The larger scale makes it difficult to find suitable heat loads for cogeneration. Hence, operation in a condensing mode could be a possibility for such a plant. This has been investigated using a condenser pressure of 0.040 bar and a reheat steam cycle with steam data 180 bar/600°C/600°C.

1 Dry steam.

74 In table 3.14, some results from these calculations are given. For comparison, the base case system with an exhaust gas dryer and pressurised gasification is also presented.

Base case Future large-scale Future large-scale system cogeneration condensing unit Gas turbine pressure ratio 17.5 30 30 ISO gas turbine inlet temperature 1140°C 1400°C MOOT Live steam data 100 bar/507°C 180 bar/600°C 180 bar/600°C/600°C Condenser pressure 1.0 bar 1.0 bar 0.040 bar Electrical efficiency (LHV) 0.448 0.515 0.576 Electrical efficiency (HHV) 0.364 0.418 0.468 Fuel utilisation (LHV) 0.856 0.877 0.576

Power to heat ratio 1.10 1.42 CO

Table 3.14. Possible future performance of large-scale biomass-based IGCC.

As a comparison, Pruschek et al. (1998) presented electrical efficiencies in the range of 0.50 - 0.54 (HHV) for a similar IGCC system (condensing unit) using coal. The higher figures, compared with the results in table 3.14, are partly due to better optimisation and lower condenser pressure, and partly due to the higher heat requirement for drying when using biomass.

75 3.4 Results and general discussion

3.4.1 Comparison ofIGCC systems

Here, results from the above sections are compiled and presented for discussion. Starting with the parametric study performed in sections 3.3.1 - 3.3.4, the aim was to compare a number of IGCC systems utilising different technologies for gasification and fuel drying. In the graphs below, results from this investigation have been gathered to facilitate comparisons.

Figures 3.51 and 3.52 show the electrical efficiency and fuel utilisation, respectively, as a function of the gas turbine compressor pressure ratio, for a live steam pressure of 100 bar. The configurations are denoted as follows:

1. IGCC with pressurised gasifier and exhaust gas dryer. 2. IGCC with pressurised gasifier and steam dryer. 2a. IGCC with pressurised gasifier and steam dryer, with separate evaporator in the HRSG. 3. IGCC with near atmospheric gasifier and exhaust gas dryer. 4. IGCC with near atmospheric gasifier and steam dryer.

(%) IGCC electrical efficiency (LHV)

25 (PR)

Figure 3.51. Electrical efficiency for various IGCC configurations, as a function of gas turbine compressor pressure ratio.

The calculations performed here indicate that systems utilising pressurised gasification have a potential efficiency advantage of 4 - 5 percentage points over systems with gasification at near atmospheric pressure. This difference is mainly caused by the electricity requirement of the fuel gas compressor in the latter system, but also by the greater heat loss from the gasifier and gas cleaning equipment.

Considering the type of dryer, the steam dryer generally gives lower electrical efficiency than the exhaust gas dryer. In figure 3.51, the difference is 1.5 percentage points for the systems

76 with pressurised gasification, and 0.7 of a percentage point for systems with gasification at near atmospheric pressure. The main reason is the greater exergy and/or energy losses in the HRSG due to unfavourable temperature profiles. In addition to this, the steam dryer has a higher fuel temperature at the dryer outlet, compared to that of an exhaust gas dryer, resulting in higher heat demand for drying.

In figure 3.52, it can be seen that the steam dryer enables higher fuel utilisation than the exhaust gas dryer. This is because all of the moisture evaporated from the fuel is condensed again in a district hearing heat exchanger. In the exhaust gas dryer, the evaporated moisture is mixed with the exhaust gas, and the resulting dew temperature is so low that none, or only a very small fraction, of the moisture will condense in the district heating heat exchanger.

IGCC fuel utilisation (LHV)

25 (PR)

Figure 3.52. Fuel utilisation for various IGCC configurations, as a function of gas turbine compressor pressure ratio.

Comparing pressurised and near atmospheric gasification, it is noted that the latter systems show lower fuel utilisation, the reason being higher heat losses from the gasifier and losses in the fuel compressor intercoolers, where some heat is lost to the surroundings.

In addition to this comparison of different IGCC concepts, a number of parameter variations have been performed. Firstly, the influence of design value of ambient temperature was studied in section 3.3.5. A decrease in temperature from 15 to 5°C resulted in 1 percentage point lower fuel utilisation, and virtually unchanged electrical efficiency. When the ambient temperature is instead -5°C, more energy has to be supplied to the dryer and the fuel utilisation is 4 percentage points lower compared with the base case. Also, the electrical efficiency is decreased, approximately 0.2 - 0.7 percentage points.

In section 3.3.6, it was shown that the moisture content of the fuel before and after drying, has a strong influence on both the electrical efficiency and the fuel utilisation. A lower moisture content after drying was advantageous in terms of electrical efficiency in all studied systems.

77 For systems with a steam dryer, there was also a significant increase in the fuel utilisation with decreasing moisture content after drying.

Considering the moisture content of the wet fuel fed to the dryer, the general trend is that lower moisture content results in higher electrical efficiency. Hence, air drying, for example in piles in the forest or at the power plant site, could be beneficial. For systems utilising exhaust gas drying the fuel utilisation is also strongly influenced by the initial moisture content, if instead a steam dryer is used, the latent heat of the fuel moisture is readily recovered in the district hearing network, and hence, the fuel utilisation is virtually unaffected by this parameter.

The parameter variations and process alterations described in sections 3.3.7 - 3.3.11, revealed some potential increases in electrical efficiency, compared with the maximum values presented in figure 3.51. The results are summarised in table 3.15, where the estimated efficiency increase in percentage points is given for various IGCC systems. If nothing else is noted the steam dryers are fed from a steam turbine extraction.

Press, gasif. Press, gasif. Near atm. gasif. Near atm. gasif. Exhaust gas dryer Steam dryer Exhaust gas dryer Steam dryer No heat loss between 0.3 1.1 0.3 1.1 dryer and gasifier. Higher fuel gas 0.4 0.6' 0.6 0 delivery temperature. Lower fuel gas 0 O' 0.5 0.6 delivery pressure. Lower pressure in - 0.5 " 0.5 steam dryer. Lower pinch point in 0.2 0.2 n.c. n.c. HRSG. Final superheating in 0.2 0.5' n.c. n.c. gas cooler. Two-pressure n.c. 0.5 n.c. 0.3 HRSG.12

Table 3.15. Possible improvements (percentage points) in electrical efficiency for studied systems. (n.c.: no calculation has been performed for this system, not applicable to this system)

The figures presented in table 3.15 require some further explanation. Starting with the heat loss between the dryer and the gasifier, it is important to note that the presented gain is the theoretical maximum if this heat loss could be completely eliminated. This is probably not possible in practice, and the real potential is consequently lower.

Higher fuel gas delivery temperature is generally advantageous, according to table 3.15. The exception is when a steam dryer with heating steam from a turbine extraction is used, since

1 Drying steam from separate evaporator. 2 See sections 3.3.2 and 3.3.4.

78 the transfer of evaporation duty from the gas cooler to the HRSG evaporator in that case decreases the steam cycle output.

Lowering the fuel gas delivery pressure in a system with gasification at near atmospheric pressure, decreases the compression work required in the fuel gas compressor, and the electrical efficiency increases correspondingly. For the pressurised systems, no such improvement was found, although the electricity requirement of the boost compressor decreased by 50%. The reason is that the lower boost compressor pressure ratio results in a lower air temperature in to the gasifier, and consequently, a larger mass flow of air was required. If a recuperative heat exchanger is added, keeping the air temperature constant when the fuel delivery pressure is decreased, the pressurised system exhibits a modest increase in electrical efficiency of approximately 0.1 of a percentage point.

For systems with steam as the drying medium, there is a possibility of decreasing the pressure levels in the dryer. Here, the estimated increase in electrical efficiency, when the heating steam pressure is decreased from 9 to 6 bar with a simultaneous decrease in conveying steam pressure from 4 to 2 bar, is 0.5 of a percentage point. It should be noted that this change decreases the potential improvement of reducing the heat loss between the dryer and the gasifier, since the fuel outlet temperature decreases with the conveying steam pressure. Hence, the figures presented in table 3.15 for heat loss and steam pressure cannot be added directly.

The influence of the minimum temperature difference in the HRSG is presented for the systems with pressurised gasification only, but the result for the near atmospheric gasification should be similar. The 0.2 of a percentage point increase is achieved through a decrease in the evaporator pinch point from 10°C to 5°C.

For the pressurised systems, also final superheating in the gas cooler has been investigated. Again, the results should be applicable also for the near atmospheric gasification.

Finally, the result of introducing a second pressure level in the HRSG is presented for the systems with steam dryers, where the exhaust temperature after the high-pressure economiser is occasionally rather high. In systems with an exhaust gas dryer, the temperature after the dryer is always too low for steam production, and the possibility of a second evaporator placed upstream from the dryer has not been investigated.

Based on the figures presented in table 3.15, there seems to be a potential for increased efficiency in the range of 1-3 percentage points for the systems studied. It is worth noting that if these measures are taken, the difference between systems based on steam or on exhaust gas drying diminishes.

Sections 3.3.1 - 3.3.4 illustrate also the influence of gas turbine COT. When there is a decrease in the COT from 1300°C to 1200°C with the same level of cooling technology, there is a rather modest decrease in the electrical efficiencies of all systems by 0.4 - 0.8 of a percentage point. Nevertheless, this parameter is probably the single most important one, as described in sections 3.3.12 and 3.3.13, where the influence of this parameter was further investigated.

When a simple uncooled gas turbine with a combustor outlet temperature of 850°C is used, the electrical efficiency decreased approximately 8 percentage points compared with the base

79 case where the COT is 1300°C. If, instead, the next generation of advanced large-scale gas turbines could be used for the biomass IGCC, an efficiency increase of more than 6 percentage points would be possible in large cogeneration plants. This figure, however, also reflects some of the improvements described in table 3.15, as well as some effects of scale.

3.4.2 Economy of IGCC

This section constitutes the exception mentioned earlier in the objectives (section 1.2), presenting estimates of the cost of electricity from IGCC systems.

Earlier investigations present numbers for present and future investment costs of IGCC systems of various sizes [Bridgwater, 1995; Larson and Marrison, 1997, Faaij, 1997; Rensfelt, 1997; Craig et al., 1994; Consonni and Larson, 1996b; Kurkela and Solantausta, 1994]. Since these systems are not yet commercially available, the cost estimates, especially for the ’’n* commercial plant”, are rather uncertain and are often given as intervals. Nevertheless, such information, together with assumptions regarding plant size, fuel price, annual operation time, interest rate and the price of heat for district heating, will be used here in an attempt to estimate the production cost of electricity for various configurations.

The IGCC systems studied above utilise different technologies for gasification and drying. Owing to this, also other parts of the integrated systems vary, e.g. boost or fuel gas compressor, hot or cold gas cleaning, number of heat exchangers, number of steam turbine extractions, etc. To make a fair comparison of the economics of these systems, the influence of such differences on the investment cost ought to be considered. Such detailed information has, however, not been obtained for this investigation. (It is usually difficult to get correct cost information even if the equipment is precisely specified, and here exact dimensions of equipment is not known.) Instead values of investment cost referring to the plant as a whole, have been taken from one of the aforementioned references, namely, Larson and Marrison (1997), who present curves for the investment cost of IGCC systems1, based on data fittings for various plant sizes. The resulting investment costs are stated to be valid for the first generation Bio-IGCC now under development, once these are commercially mature. It should be noted that the values reached asymptotically at large plant outputs, are in agreement with what was recently presented for the large-scale coal-based IGCC [Pruschek et al., 1998].

The following data fits are used to calculate the specific investment cost (USD/kW e)12 for IGCC plants in the size range of 20-100 MWe [Larson and Marrison, 1997].

Near atmospheric gasification: 1200 + 47198*MW e'156 (3.18) Pressurised gasification: 1100 + 110420*MW e'1 '42 (3.19)

Different correlations are presented for systems with pressurised gasification and for systems with near atmospheric gasification, and with gas cleaning equipment similar to what has been assumed earlier in this report, i.e. hot and cold gas cleaning, respectively. In both cases, an exhaust gas dryer is assumed for drying the wet biomass, and hence, no influence of different drying technologies is included in the presented curves.

1 Both condensing plants and cogeneration plants are included. 2 USD 1994.

80 Bridgwater (1995) presented a cost analysis for a small-scale IGCC plant where the dryer constitute approximately 7% of the total investment. Hence, the influence of different costs for different types of dryers may not be negligible. Furthermore, Wimmerstedt and Linde (1998) indicate there is indeed a difference in investment costs when comparing a rotary drum exhaust gas dryer with a steam dryer of pneumatic conveyer type. In this report, however, it is assumed that both the exhaust gas dryer and the steam dryer are of pneumatic conveyer type, and hence, it is believed satisfactory to assume roughly the same investment cost for both dryers. However, it should be noted that changing the exhaust gas dryer to a rotary drum dryer could decrease the investment cost compared to the systems using steam drying.

Using equations 3.18 and 3.19, after conversion to Swedish crowns (SEK)1, results in investment costs varying with plant size as shown in figure 3.53. The dashed lines indicate an interval for the investment cost of ± 20%, reflecting both the uncertainty of this kind of estimation and a possible influence of fluctuating market conditions.

As a comparison, the investment cost for conventional cogeneration plants, with solid fuel boiler and steam turbine, is said to be in the range 13000 - 15000 SEK/kW e [Jansson, 1997] and this is probably valid for plants with 20 - 40 MWe output. Also for that type of plant there is a trend of decreasing specific investment cost with increasing plant size.

(SEK/kWe) Investment cost vs plant size 30000

25000

20000

•Pressurised gasification 15000 •Near atmospheric gasification 10000

5000

(MWg)

Figure 3.53. Specific investment cost for IGCC systems.

In addition to the capital cost, also operation and maintenance (O&M) costs and fuel costs will be included. Based on numerous studies [Faaij, 1997; Gustavsson et al., 1995; Craig et al., 1994; Larson and Marrison, 1997; Kurkela and Solantausta, 1994; Larsson, 1998] the fixed O&M cost is estimated to 2.5 % of the investment cost per year, and the variable O&M cost to be 0.03 SEK/kWhf, regardless of the plant size.

1 Conversion factor 7.725 used to calculate 1998 SEK, based on 1994 exchange rate and Swedish consumer price index.

81 The final cost item is the fuel price. The cost of biomass depends largely on quality and transportation distance to the power plant. In recent years, the price of wood chips delivered to district heating plants has been around 110 - 115 SEK/MWhf, based on the lower heating value [Energimyndigheten, 1998]; thus, 110 SEK/MWhf will be used as the base case value in the following calculations.

In the calculations, a basic assumption is that the plant is used primarily to cover a heat demand in the district heating network. This heat demand is taken to be equal for all systems, and due to variations in the power to heat ratio, the resulting electricity production varies, as does the investment cost, since it is based on the power output [equations (3.18) and (3.19)]. In the base case, the heat output from the plant is 50 MW*.

It is assumed that this plant is operated as a base load plant in the district heating system, with an average utilisation time of 4500 hours/year at nominal capacity.

Furthermore, the income from produced heat must be assigned a value in order to calculate the cost of electricity. Here, the value of produced heat has been calculated as an avoided cost for heat production in a biomass-fired CFB boiler. This cost, in turn, is calculated based on assumptions made for investment cost, O&M cost and fuel cost for a boiler plant with the same heat output as the IGCC plant. For details, see appendix EL

Using the annuity method with an interest rate of 6% and a service life of 20 years, the annual capital cost can be calculated. Adding the O&M cost and the fuel cost gives as a result the cost of electricity, which is presented in the graphs below.

(SEK/MWhe) Cost of electricity vs plant size

-flue gas dryer, pressurised gasification

-steam diyer, pressurised gasification

-flue gas dryer, near atmospheric gasification

-steam dryer, near atmospheric gasification

0 20 40 60 80 100 (MWth)

Figure 3.54. Cost of electricity for IGCC systems of different sizes.

From figure 3.54, it can be seen that the higher capital cost at lower capacities results in a higher cost of electricity for the systems with pressurised gasification. It is also clear that the cost of electricity, even at larger scale, is not competitive under present Swedish conditions.

82 A higher fuel price results in a higher cost of electricity, as shown in figure 3.55, but the effect is weakened by'the fact that also the price of the produced heat increases with the fuel price. In this context it should be remembered that the result is presented for the base case plant size SOMWuv

(SEK/MWhe) Cost of electricity vs fuel price

-flue gas dryer, pressurised gasification -steam dryer, pressurised gasification flue gas dryer, near atmospheric gasification —X—steam dryer,near atmospheric gasification

70 80 90 100 110 120 130 140 150 160 (SEK/MWh,)

Figure 3.55. Influence of fuel price on cost of electricity for IGCC systems.

If, for example, due to problems with the availability of the plant, the annual operating time falls, then the cost of electricity increases markedly. This is shown in figure 3.56.

(SEK/MWhe) Cost of electricity vs annual full load hours

-flue gas dryer, pressurised gasification

-steam dryer, pressurised gasification

-flue gas dryer, near atmospheric gasification

-steam dryer, near atmospheric gasification

0 1000 2000 3000 4000 5000 6000 7000 8000 (h/year)

Figure 3.56. Influence of annual operating time on cost of electricity for IGCC systems.

83 Different investors, e.g. utilities, industries and municipalities, sometimes use different rates of interest when estimating the profitability of an investment. Furthermore the rate of interest changes over time and the choice made for this calculation was somewhat arbitrary. Hence, figure 3.57 is included to show the influence of the rate of interest on the calculated cost of electricity.

(SEK/MWhe; Cost of electricity vs interest rate

-Hue gas dryer, pressurised gasification

-steam dryer, pressurised gasification

-flue gas dryer, near atmospheric gasification

-steam dryer, near atmospheric gasification

20 (%)

Figure 3.57. Influence of interest rate on cost of electricity for IGCC systems.

Compared with the present production costs in existing Swedish power plants, the cost of electricity presented for the base case here is not competitive. In a possible future situation, where new production capacity is required, these results should instead be compared with the cost of electricity from alternative new plants.

3.4.3 Plant availability

For this technology to be competitive, not only thermal performance but also plant availability and reliability have to be competitive with or better than present-day technology.

The Bio-IGCC plant in Vamamo (section 3.1) has accumulated some 2000 hours of fully integrated operation on product gas [Sjunnesson, 1999], and this is more than any other IGCC plant based on biomass. Most of these hours have so far been collected during shorter test runs, and hence, it is almost impossible at this point to judge the possibilities of this technology in terms of availability.

Considering the system as a whole, there are great similarities between the IGCC using biomass and the IGCC fuelled by fossil fuels. IGCC plants for coal and heavy oil have been demonstrated for somewhat longer periods, as shown for example in table 3.1, and the operational results may be taken as an indicator of the availability of the corresponding IGCC for biomass.

84 Starting with the IGCC for coal, the plant in Buggenum, with a total of 11000 hours of operation since 1993, demonstrated 75% availability during 1997, including the yearly shutdown of three to four weeks [Stambler, 1997]. A later report indicates the same result for the first part of 1998 [Holt, 1998].

Availability of 75% is also reported for the Polk County plant during the last quarter of 1997 and the first quarter of 1998 [Holt, 1998]. After some initial problems, the Wabash River plant has demonstrated similar availability for the second half of 1997, and according to the project manager, the technology has the potential for 80 - 85% availability [Stambler, 1997].

The El Dorado Refinery project consists of a smaller IGCC plant, using a GE 6B gas turbine rated at 35 MWe, where petroleum coke and/or heavy oil is gasified. For this plant, the availability was reported to be close to 85% for the period March to December 1997. In other refinery projects for gasification of heavy oil availabilities in the range of 85 - 90% or even higher are projected, based on design improvements and operating experience from earlier plants [Stambler, 1997].

A common feature for all these plants is the initial period of low availability due to component failures. Examples of problematic equipment are heat exchangers (corrosion and tube leaks), hot gas filters (breakage and low collection efficiency), lock-hopper systems, ash and particulate removal systems and, in the Buggenum plant, the gas turbine (vibrations in combustors when burning coal gas).

These problems are not specific to the fossil fuel IGCC and could be expected also the for biomass-based IGCC. In fact, the lock-hopper systems for fuel feeding to the gasifier and for withdrawing ash and particles, have caused some problems at the Vamamo plant [Stahl, 1997].

There are, however, also problems specific to biomass. One such is the tar production in wood gasification, which is significantly greater than for coal [Bridgwater, 1995]. Furthermore the tars are often heavier and more stable, and hence, additional tar cracking, as described in section 3.2.3, may be required. Otherwise, filter blockage and heat exchanger fouling can occur, decreasing the plant availability.

Compared with the coal-based IGCC, the proposal for biomass IGCC presented in this work includes an additional piece of equipment, the dryer, which, of course, must also exhibit sufficient availability if plant availability is not to be decreased. Here, it could be worth mentioning that so far the steam dryer has problems reaching the same availability as the simpler rotary drum exhaust gas dryer [Wimmerstedt and Linde, 1998].

On the other hand, the IGCC plants for coal presented above have more or less integrated air separation units for oxygen production and gas cleaning steps for sulphur removal, equipment not required in the biomass IGCC.

All in all, there is no obvious reason why a biomass-based IGCC, once the fuel feeding system and the tar control methods have been sufficiently developed, should not achieve the same availability as a corresponding IGCC based on coal.

85 3.4.4 Development needs

It has been shown in earlier sections that the thermal performance of the IGCC could be improved by component development, e.g. through higher temperature in the fuel gas feeding system, superheating of steam in the raw gas cooler, higher gas turbine combustor outlet temperature, etc. There are, however, also other reasons for a continuous development of components and systems for the IGCC, and some of these are reviewed next.

Starting with fuel drying, the experience from the steam dryer is so far not entirely positive. One problematic issue is the fuel feeding and withdrawal when the dryer is operated at elevated pressure; another is the corrosion and erosion experienced in many types of dryers. Also, the treatment of the condensate from the dryer is sometimes a problem due to the low pH and high chemical oxygen demand (COD) [Wimmerstedt and Linde, 1998]. Hence, development is required before potential advantages of the steam dryer, e.g. high fuel utilisation in cogeneration mode, and decreased risk of fire and dust explosion in the dryer, can be fully utilised.

Also, for the pressurised gasification, the fuel feeding equipment could still be improved. Issues here are both reliability and consumption of inert gas. The lock-hopper systems used today often require large amounts of inert gas (usually N2) for pressurisation due to the low bulk density of the fuel. Alternative solutions utilising piston feeders [Stahl, 1997; van Ree, 1994] and screw-piston feeders [Liinanki and Karlsson, 1994] have been studied as possible ways to reduce the consumption of both inert gas and electricity. More testing of these concepts in biomass applications is required however. Smaller amounts of inert gas is beneficial not only through lower cost but also through the resulting increase in fuel gas heating value. Also, the use of C02 recycled from gas combustion as inert gas has been suggested, as a way to reduce the cost [Bridgwater, 1995].

For both pressurised and near atmospheric gasification, bed materials with a catalytic effect on tar cracking could be beneficial for the subsequent gas cooling and cleaning, reducing the risk of tar depositions on heat exchanger surfaces. In the case of cold gas cleaning, the cracking of tars also simplifies the waste water treatment after the scrubber. Promising results have been achieved, as presented in section 3.2.3, through the use of a second catalytic reactor. More test results from continuous operation for longer periods would increase the confidence in this technique, regarding, for example, possible catalyst deactivation. If it was possible to perform the cracking within the gasifier itself, the plant complexity and the investment cost could be decreased, and hence, further research in this area is justified.

Apart from tar cracking, also other parts of the gas cleaning equipment need to be further developed or proven reliable during extended test periods. One example is the hot gas filter used for particulate and alkali removal. As was stated in section 3.2.3, more experience from long-term testing on fuel gas from biomass is required.

The major concern in gas cleaning may prove to be hot gas cleaning of nitrogen compounds, in order to limit the NOx emissions. High temperature Ni-catalysts and selective oxidation of the fuel gas are two possibilities, but it seems that some development work still remains. Also, further development of gas turbine combustors could be useful in this respect. Tests with a rich-quench-lean (RQL) combustor have shown ammonia-to-NOx conversion as low as 5%, at combustor outlet temperatures in the range of 1150 - 1200°C, when burning fuel gas with high

86 concentrations of ammonia [Layne and Hoffman, 1998]. Another possibility, utilising proven technology, is catalytic reduction of NOx in the HRSG, but the investment cost and efficiency penalty make it undesirable.

Finally, the need for development of the gas turbine technology, and the potential efficiency improvements resulting from this development, should be emphasised.

In order to use the low heating value gas as a gas turbine fuel, some modifications of the fuel feeding system and the combustors are generally required. In addition to this, it would be desirable to design the turbine and/or compressor to comply with the increased fuel gas flow. This is especially vital for systems with near atmospheric gasification, where gasification air is not bled from the compressor. None of these changes requires any new technology, but presently the market for such gas turbines is probably too small to justify the cost of the development work.

A more fundamental issue is the interaction between fuel gas cleaning (gas quality) and gas turbine parameters, such as combustor outlet temperature, materials and coatings, cooling methods, interval between overhauls etc. It is likely that more research is required to better understand the requirements imposed by the fuel gas on the hot parts of the gas turbine, and vice versa.

Provided that the development of fuel gas cleaning methods is sufficiently successful, any further improvement in gas turbine technology, regarding, for example, material heat resistance or component efficiencies, should be applicable also to the biomass IGCC. Hence, there is a potential for even higher electrical efficiency, and this is probably the single most important advantage of the IGCC compared to other technologies utilising biomass for cogeneration, e.g. boiler and steam turbine plants.

87 4 Integrated gasification humid air turbine

4.1 Overview

In a combined cycle, the heat remaining in the gas turbine exhaust gas is utilised in a bottoming Rankine cycle, thereby raising the efficiency and work output. In the humid air turbine (HAT) concept, the Rankine cycle is omitted, and the exhaust gas heat is recovered and recycled into the gas turbine. Compared with the IGCC layout, the HRSG has been replaced by a recuperator and an economiser, resulting in an integrated gasification humid air turbine (IGHAT). A simplified schematic process layout is given in figure 4.1.

------®

Gasifier HUM

Make-up water

AC Aftercooler GC Gas cooler C Compressor HUM Humidifier DH District hearing heat exchanger Rec Recuperator ...... : Biomass ECO Economiser T Turbine ------: Water/steam G Generator ™ — — : Fuel gas

Figure 4.1. Process layout for IGHAT with exhaust gas dryer, (simplified)

In a standard HAT-cycle as presented by Rosen (1993), Agren (1997), Chiesa et al. (1995) and others, the economiser is used, together with the gas turbine compressor intercooler1 and aftercooler to heat water; thereby, low temperature sensible energy is recovered. The hot water is then routed to a humidification tower (humidifier), where some of the hot water is evaporated into the compressed air. When water is evaporated into a gas stream, it is the partial pressure of water vapour in the gas phase that determines the vaporisation temperature. At air inlet, in the bottom of the humidification tower, the partial pressure of the water vapour

1 Requires an interceded gas turbine. In this report the intercooler has been omitted.

88 is low, and hence, the vaporisation temperature is also low. As the air passes through the tower, meetingthe hot water in a counter-current flow, the humidity of the air increases, as does the vaporisation temperature. Due to the variation in the vaporisation temperature, this evaporation process has a smaller exergy loss than the corresponding evaporation in the HRSG boiler.

The heat required for vaporisation is taken as sensible heat from the hot water, and in some cases, also from the air, depending on the inlet air temperature. After being cooled in the humidification tower, the water is mixed with make-up water, corresponding to the amount evaporated, and re-routed to the heat recovery system (intercooler, aftercooler and economiser).

After the humidifier, the air temperature is rather low, and hence, it is possible to heat the humid air in a recuperator, even at gas turbine pressure ratios substantially higher than normal for recuperative cycles. Furthermore, the mass flow through the turbine is increased by the added water, without a corresponding increase in compressor work. The result of these features is an increase in thermal efficiency compared with the simple gas turbine cycle.

When the HAT-cycle is integrated with drying and gasification of biomass, an additional heat source, the gas cooler, is introduced. Sensible heat is transferred from the fuel gas to the water circuit and the humidifier. The amount of heat available from gas cooling depends on the fuel gas temperature required by the gas cleaning devices. Since the initial gas temperature is high (about 900°C) and the final water temperature is only about 200°C, the exergy loss in the gas cooler is substantial.

An additional heat sink, the dryer, is also added. When the exhaust gas is used as a drying medium, less energy is available in the economiser. In the case of a steam dryer, water has to be evaporated, and the heat for this may also be taken from the exhaust gas or from some other part of the cycle, decreasing the amount of heat recovered into the gas turbine.

Both in the humidifier and in the dryer, energy is used to evaporate water. In order to recover this energy, the water in the exhaust gas has to be condensed. Due to the high moisture content, this can be accomplished in a condensing heat exchanger that heats water for district heating.

89 4.2 Models

The dryer and gasifier models used are the same as presented in sections 3.2.1 and 3.2.2. The gas turbine model is also equivalent to the one used above. Hence, it is assumed that the gas turbine is customised to the HAT-cycle operation, allowing for a substantial flow increase in the turbine compared with the compressor. The recuperator and economiser replacing the HRSG in the IGCC, are ordinary heat exchangers, and the only new item to be modelled is the humidification tower.

4.2.1 Humidifier

The model used here is similar to the one presented in Agren (1997). From a SPLIT-block, part of the water recovering the sensible energy in the gas cooler, aftercooler and economiser is led to a MIXER-block and mixed with the compressor air. Two HEATER-blocks are also included to take into account the sensible energy transferred from the circulating water which is not evaporated in the tower. This is shown in figure 4.2, together with the notation used to explain the assumptions.

Heater Humid air to gas Compressed air from aftercooler turbine combustor

Mixer

H20 from aftercooler

H20 from gas cooler

H20 from economiser

Heater

Water to heat recovery system Make-up water

Figure 4.2. Humidifier model in ASPEN PLUS™.

To verify the feasibility of the humidification process, a simplified method similar to Chiesa et al. (1995) is used. It is based on a temperature-enthalpy diagram of the humidifier as shown in figure 4.3. It comprises two curves; a saturation curve representing the enthalpy of saturated air, and a working line representing the enthalpy of the bulk phase air in contact with water at a given temperature. The distance between these two curves can be interpreted as a driving force for the humidification process. To ensure the feasibility of this process, the enthalpy

90 driving force must always be positive. (If the driving force goes to zero, the required gas- liquid contact surface in the humidifier goes to infinity.)

In the calculations, this is checked in the following manner. A straight line connecting the extremes of the working line is compared with the saturation curve 1. A minimum enthalpy difference between these curves and a certain amount of superheat12 at the air outlet are defined, and the evaporation rate is varied until these preconditions are fulfilled. Since the real working line is slightly curved, the real enthalpy difference is greater than the calculated one, ensuring a positive enthalpy driving force everywhere in the humidifier [Chiesa et al., 1995; Agren, 1997].

Air enthalpy

Saturation curve

Working line

Minimum enthalpy driving force

Temperature

Figure 4.3. Air saturation curve and humidifier working line.

In an ordinary HAT-cycle, the water temperature at the humidifier outlet should be as low as possible in order to recover also the low temperature heat in the intercooler and the economiser. To achieve this, both the temperature of the air from the aftercooler and the pinch-point3 at the water outlet should be kept at a minimum.

When the HAT-cycle is integrated with drying and gasification, the situation may change depending on the actual configuration. Firstly, if the wet biomass is dried in an exhaust gas dryer (figure 4.1), there is a limit on the exhaust gas temperature out of the economiser in

1 In ASPEN PLUS, the reference state for enthalpy calculations for water/steam is not the same as for gaseous constituents (25°C, 1 atm). Hence, it is necessary to redefine the reference state for water in order to enable calculations with humid air. At 25°C and 1 atm, the specific enthalpy for H20(1) is given as -15865.3 kJ/kg. By adding the absolute value of this number to the specific enthalpy of water, it is possible to calculate the specific enthalpy of humid air. This is done when calculating the saturation curve and the working line in order to assure physical feasibility of the humidification process. 2 A common approach is to assume a saturated condition at air outlet, as in Rosen (1993) and Chiesa et al. (1995). Others assume that the air is slightly superheated [Agren, 1997; Cavani, 1998]. 3 Rosen (1993) introduced a pinch-point as the temperature difference between the circulating water at the outlet from the humidifier and the lowest possible water temperature at this point. The lowest possible water temperature is, in theory, the saturation temperature of the compressed air entering the humidifier. Using the notation in figure 4.3, the pinch-point is then: ATWjOUt = Tw out - TaMsal)

91 order to satisfy the demand of the dryer downstream. Secondly, there is no intercooler, but a fuel gas cooler. In the case of pressurised gasification, the gas temperature out of the cooler is much higher than the water temperature. Both these features decrease the importance of a low water temperature supplied to these heat exchangers, and the cold make-up water is, therefore, mixed with the flow to the aftercooler. Thereby, less work is required in the boost compressor.

When a steam dryer is used, utilising heat from the gas cooler (figure 4.6), more sensible energy is available for recovery in the economiser, and the exhaust gas outlet temperature depends on the water inlet temperature and the heat exchanger temperature difference. Hence, the water inlet temperature to the economiser should be low, and the make-up water is routed to this heat exchanger to lower the exhaust gas temperature.

In the case of cold gas cleaning (near atmospheric gasification), a low water temperature is also of greater importance. Keeping the water temperature at the humidifier outlet low, makes it possible to recover sensible energy from the fuel gas that would otherwise be lost to the environment. The cold make-up water should be mixed with the flow to the gas cooler and the fuel gas compressor intercoolers in order to maximise the heat recovery (figure 4.9). When near atmospheric gasification and cold gas cleaning is combined with steam drying (figure 4.12), the make-up water is added before the flow is split among the economiser, the gas cooler and the fuel gas compressor intercoolers.

In the modelling of the gas turbine, it has been assumed feasible to use the humidified air as a cooling medium. For simplicity, the firing temperature and the parameters in the cooling model (section 3.2.4) are held constant, and the lower temperature of the humid air, as compared with the dry air in a gas turbine without an aftercooler, then results in lower mass flows of cooling air. Another possibility would have been to keep the volume flow of the cooling air constant, allowing for an increase in the firing temperature at a constant material temperature.

4.3 Parametric study

Similar to the above study on the IGCC, a number of systems, and the influence of various parameters on these systems, are investigated. For the IGHAT, the two types of gasification, pressurised and near atmospheric, are studied. Also, the two types of dryers presented earlier, using steam or exhaust gas as the drying medium, are investigated.

Due to problems reaching the pre-set district heating supply temperature, two different systems are studied. In the first one, the exhaust gas is cooled to 60°C in a condensing heat exchanger in order to recover some latent heat. Since condensation starts at about 75 - 80°C, a large flow of district heating water can be heated approximately to this temperature, recovering a significant amount of the energy used for evaporation in the humidifier. Further heating this flow to the pre-set supply temperature would require additional heat, and this would decrease the amount of heat available for recovery in the HAT cycle. Therefore, in the first system, the district heating supply temperature is allowed to decrease.

In the second system, 90°C supply temperature is required. To fulfil this requirement, the flow of district heating water is decreased, thereby increasing the flue gas temperature and

92 correspondingly the flue gas losses1. The reason for decreasing the water flow is to limit the amount of additional heat required to reach the final temperature, so that the electrical efficiency will not suffer too much. The decrease in fuel utilisation will, however, be significant. Apparently there is a ’’trade-off’ between heat and electricity.

The assumptions used in the calculations are presented in table 4.1. Only the values differing from table 3.11 in section 3.3 are presented.

Recuperator Minimum AT=30°C Air side pressure drop: Ap/p=0.02 Economiser Minimum AT=10°C Final water temperature 10°C below saturation Water side pressure drop: Ap/p=0.05 DH heat exchanger Minimum AT=10°C (except exhaust gas condenser) Minimum AT=5°C (exhaust gas condenser) Minimum flue gas outlet temperature: 60°C Gas cooler Final water temperature 10°C below saturation Water side pressure drop: Ap/p=0.05 Aftercooler Minimum AT=10°C Final water temperature 10°C below saturation Water side pressure drop: Ap/p=0.05 Air side pressure drop: Ap/p=0.01 Humidifier Water inlet pressure drop: Ap/p=0.02 Water inlet temperature: 10°C below saturation at humidifier operating pressure Minimum enthalpy driving force: 25 kJ/kg dry air Outlet air temperature: 5°C above saturation temperature Steam dryer High pressure steam 9/16 bar Others Total exhaust gas pressure drop12: Ap=3 kPa

Table 4.1. Assumptions used for calculations ofIGHAT systems.

Any additional assumptions made for the various configurations are presented in the following subsections.

1 See appendix HI. 2 Gas turbine outlet pressure = patm+Ap. 4.3.1 Pressurised gasification and exhaust gas dryer

This configuration was shown in figure 4.1. The exhaust gas dryer is placed downstream from the economiser in the exhaust gas stream, imposing a limit on the temperature out of the economiser.

The first parameters studied are the gas turbine compressor pressure ratio and the combustor outlet temperature. Figure 4.4 shows how the electrical efficiency varies with the PR and the COT. Compared with the IGCC, the PR for maximum electrical efficiency is significantly higher for the IGHAT.

(%) Electrical efficiency (LHV)

1300C

'1200C 43 - -

41 - -

Figure 4.4. Electrical efficiency for IGHAT with pressurised gasification and exhaust gas dryer.

For the COT 1300°C, the highest electrical efficiency is 44.0% at a pressure ratio of 30. The curve is very flat in the range of 25 - 35. The corresponding temperature of the district heating water is, however, only 77°C. Replacing the economiser with a second heat exchanger and decreasing the flow of district heating water to reach 90°C supply temperature, results in a decrease in electrical efficiency of almost 2 percentage points. The maximum is also shifted towards a somewhat higher PR.

Figure 4.5 shows how the fuel utilisation for the two systems varies with the PR and the temperature. When the supply temperature for the district heating is allowed to be lower than 90°C, the exhaust gas can be cooled to 60°C in a condensing heat exchanger, giving values of fuel utilisation comparable to or better than the IGCC. The district heating supply temperature is then in the range of 72 - 77°C. To reach 90°C, the economiser is replaced by a second heat exchanger for the district heating water, and the flow of the district heating water is decreased. The result is a higher flue gas temperature, and correspondingly less condensation. The fuel utilisation then drops dramatically.

94 (%) _ Fuel utilisation (LHV)

35 (PR)

Figure 4.5. Fuel utilisation for IGHAT with pressurised gasification and exhaust gas dryer.

95 4.3.2 Pressurised gasification and steam dryer

In this system, the fuel is dried in a steam dryer similar to the one presented in section 3.2.1. The high-pressure steam supplying the heat is here produced in the gas cooler, as can be seen in figure 4.6.

Gasifier

HUM Make-up water

Figure 4.6. IGHAT with pressurised gasification and steam dryer, (simplified, abbreviations as in figure 4.1)

Since the outlet gas temperature is high, a relatively high steam pressure (here 16 bar) can be used to decrease the physical size of the dryer and to increase the temperature difference in the dryer.

To produce enough heating steam for the dryer, the fuel gas has to be cooled to a lower temperature than in the case of an exhaust gas dryer. Typically, the temperature after the gas cooler is around 300°C, instead of 400°C. This means less sensible energy is transferred to the gas turbine combustor, lowering the efficiency. Another, more practical, problem is the increased risk of tar condensation in the gas cooler. If this should prove to be a serious constraint, some of the heat required in the dryer has to be taken from the aftercooler or the economiser instead.

Heat is now recovered to the HAT-cycle only in the aftercooler and the economiser. District heating water is heated first in the exhaust gas condenser (DH1), and finally in the condensing heat exchanger after the dryer (DH2). The amount of heat possible to recover, again depends on the required temperature level in the district heating network.

96 Figure 4.7 shows how the electrical efficiency varies with the PR and the COT. Compared with the IGHAT with an exhaust gas dryer, the maximum electrical efficiency here is 0.8-0.9 of a percentage point lower. In this case, it is, however, possible to achieve the 90°C district heating water temperature without any decrease in electrical efficiency. Hence, if 90°C is required, this system gives electrical efficiencies 0.9-1.0 percentage point higher than the system with an exhaust gas dryer.

(%) Electrical efficiency (LHV)

Figure 4.7. Electrical efficiency for IGHAT with pressurised gasification and steam dryer.

In the case of the lower temperature level, the fuel utilisation for this system is very close to the former, as can be seen in figure 4.8. This is logical, since the flue gas leaves the system at the same temperature, 60°C. The differences occurring are due to the fuel moisture, here leaving the system through the condenser after the dryer, all in a liquid phase and at a different temperature. The attainable district heating temperature is somewhat higher for this system, in the range of 78 - 83°C. Decreasing the flow of district heating water to reach 90°C, again decreases the fuel utilisation significantly, but not as much as in the case of an exhaust gas dryer.

Fuel utilisation (LHV)

1200C

15 20 25 30 35 (PR)

Figure 4.8. Fuel utilisation for IGHAT with pressurised gasification and steam dryer.

97 4.3.3 Near atmospheric gasification and exhaust gas dryer

When the gasification is performed at near atmospheric pressure, cold gas cleaning is applied, resulting in a somewhat more extensive heat recovery system, as can be seen in figure 4.9.

----- 0

HUM

Make-up Gasifier water

SC: Scrubber RH: Clean gas reheater

Figure 4.9. IGHAT with near atmospheric gasification and exhaust gas dryer, (simplified, abbreviations as in figure 4.1)

The hot raw gas is first used to reheat the clean gas before the gas turbine combustor. After this, there is a gas cooler (GC) similar to the one in the pressurised systems. Due to material constraints, the depicted gas/gas heat exchanger could prove to be unfeasible. This could be solved by dividing the GC into two parts connected in series with the reheater in between, thereby allowing for efficient cooling at the higher temperature, as well as recovery of low temperature heat before the scrubber.

The saturated gas leaving the scrubber is then compressed in a three-stage intercooled compressor (see section 3.2.6). Some heat is recovered to the water circuit, and the gas is then further cooled by means of external coolers to decrease the compression work.

Also, for this system, it is difficult to obtain the 90°C district heating supply temperature, and in figure 4.9, the system with the economiser replaced by a second heat exchanger for district heating is shown. The resulting decrease in electrical efficiency can be seen in figure 4.10.

98 When 90°C district heating supply temperature is required, the electrical efficiency is almost 37% (LEV). If can also be seen that the difference between the two levels of the COT is small.

(%) Electrical efficiency (LHV)

.13000

12000

30 (PR)

Figure 4.10. Electrical efficiency for IGHAT with exhaust gas dryer and gasification at near atmospheric pressure.

Allowing for a lower supply temperature and replacing the DH2 in figure 4.9 by an economiser, results in approximately 2 percentage points higher efficiency. The increase in fuel utilisation is even more pronounced, as can be seen in figure 4.11.

(%) Fuel utilisation (LHV)

■i3ooti:

12000

30 (PR)

Figure 4.11. Fuel utilisation for IGHAT with exhaust gas dryer and gasification at near atmospheric pressure.

The trends in the graphs above are similar to what was presented earlier for pressurised gasification. Both the fuel utilisation and the electrical efficiency are, however, significantly lower when gasification is performed at near atmospheric pressure. The reasons for this are mainly the same as for the IGCC, i.e. higher heat losses and fuel gas compression work.

99 4.3.4 Near atmospheric gasification and steam dryer

The fourth system combines gasification at near atmospheric pressure with steam drying of the fuel. The studied system is presented in figure 4.12. When comparing it with the other systems, it is clearly the most complex one, recovering heat to the humidifier from four different sources. As can be seen, the production of steam for the dryer is here divided into two evaporators, instead of one, as in the pressurised case. The reason for this is that no single heat source of sufficient temperature is available when reheat of the fuel gas is performed in the gas cooling section.

HUM

Gasifier Make-up water

condensate

Figure 4.12. IGHAT with steam dryer and gasification at near atmospheric pressure, (simplified, abbreviations as in figure 4.9)

Figure 4.13 shows the electrical efficiency for this system. Since the district heating utilises only true waste heat, the electrical efficiency is independent of the supply temperature. The maximum efficiency at almost 38% is approximately 1 percentage point lower than the corresponding maximum efficiency for the system with an exhaust gas dryer. Compared with the system utilising pressurised gasification, the maximum efficiency is about 5 percentage points lower. It may also be noted that the PR at the optimum is significantly lower for near atmospheric gasification.

100 (%) Electrical efficiency (LHV)

1300C

1200C

30 (PR)

Figure 4.13. Electrical efficiency for IGHAT with steam dryer and gasification at near atmospheric pressure.

Turning to the fuel utilisation, figure 4.14, one sees that the shape of the curves is very similar to what was presented for the system with pressurised gasification. The higher values of fuel utilisation shown correspond to district heating temperatures in the range of 75 - 82°C.

(%) Fuel utilisation (LHV)

—-

30 (PR)

Figure 4.14. Fuel utilisation for IGHAT with steam dryer and gasification at near atmospheric pressure.

101 4.4 Results and discussion

4.4.1 Comparison of IGHAT systems

Here, some of the results given in the previous sections are summarised and discussed. Only results for the combustor outlet temperature 1300°C are given, but the tendencies are similar for the lower temperature. Figures 4.15 and 4.16 show the electrical efficiency and fuel utilisation, respectively, as a function of the gas turbine compressor PR.

The four configurations are numbered as follows:

5. IGHAT with pressurised gasifier and exhaust gas dryer. 6. IGHAT with pressurised gasifier and steam dryer. 7. IGHAT with near atmospheric gasifier and exhaust gas dryer. 8. IGHAT with near atmospheric gasifier and steam dryer.

The letter ”a” denotes systems with district heating supply temperature lower than 90°C.

(%) IGHAT electrical efficiency (LHV)

7a.

35 (PR)

Figure 4.15. Electrical efficiency for various IGHAT configurations, as a function of gas turbine compressor pressure ratio.

From figure 4.15 it can be concluded that the electrical efficiency obtainable with pressurised gasification is 5 percentage points higher than when near atmospheric gasification is used. It should also be noted that the optimal gas turbine PR is considerably higher with pressurised gasification. The curve is, however, very flat, and in order to keep the gasification at a reasonable pressure level, a PR somewhat lower than optimum could be chosen.

There is also a difference between the two types of dryers. When the supply temperature in the district heating network is allowed to be lower than 90°C, so that all the sensible energy in the exhaust gas down to 60°C can be utilised, the exhaust gas dryer has an advantage of about 1

102 percentage point. If a supply temperature of 90°C is required, then the steam dryer gives the highest efficiency, the difference again being approximately 1 percentage point.

In figure 4.16, the fuel utilisation is presented. When the district heating supply temperature is allowed to fall below 90°C, the flue gas dryer gives the highest fuel utilisation. The difference between the two types of dryers is, however, small. The resulting district heating supply temperatures are in all cases in the range of 73 - 83°C.

If 90°C district heating supply temperature is required, the situation changes, with the steam dryer now giving the higher fuel utilisation. The decrease is 15 - 30 percentage points as compared to the cases with lower supply temperatures, and hence, the temperature levels in the district heating network are of utmost importance for these systems.

(%) IGHAT fuel utilisation (LHV)

35 (PR)

Figure 4.16. Fuel utilisation for various IGHAT configurations, as a function of gas turbine compressor pressure ratio.

4.4.2 Development needs

The HAT cycle is presently under development, but before an IGHAT cycle can be realised substantial development work is required. Regarding the dryer, gasifier and gas cleaning equipment, the development needs are the same as presented for the IGCC in section 3.4.4.

As in the case of the IGCC, also here the gas turbine has to comply with a mismatch in mass flow between compressor and turbine. Due to the humidification of the air, this difference is now even larger and changes to the compressor or turbine, as well as to the combustor, are required. Moreover, the feasibility of using cooled, humidified air for cooling of vanes and blades in the turbine has to be investigated more thoroughly. This in turn may lead to redesign of cooling channels etc.

103 To achieve the efficiencies presented in figures 4.15 and 4.16, the highest possible amount of water has been -evaporated into the air, the limit being the amount of heat available in the cycle. Even though the oxygen concentration in the exhaust gas out of the combustor is rather low, typically in the range of 3 - 5 vol-% (dry gas) at the highest moisture contents, constraints in the form of combustion efficiency and stability have not been considered. In the case with maximum electrical efficiency, system 1 with PR 35, the air after the humidifier contains 0.36 kg H20/kg dry air. For the same system with the district heating supply temperature 90°C, the amount of heat available for evaporation of water in the humidifier is decreased, and the corresponding figure is 0.29 kg H20/kg dry air. Without further experimental research, it is difficult to predict if any, the lower one, or both of these levels are practically feasible.

Nakhamkin et al. (1994) studied combustion of natural gas and syn-gas (from coal gasification) in humid air at elevated pressure. The test results indicate that decreasing combustion efficiency might constrain the permissible amount of moisture in the combustion air. Furthermore, the air temperature at the combustor inlet showed to be a significant parameter, with a higher temperature raising the limit. Eventually, it was concluded that a careful combustor design should make it possible to move the limit into the region of 0.30 - 0.45 kg H20/kg dry air, when burning syn-gas from coal. Since the tested syn-gas differs markedly from the fuel gas from biomass gasification, these figures are not necessarily valid for the application described in this report.

The limited amount of excess oxygen also limits the possible further increase in the COT at constant humidity in the combustion air. Any future increase in permissible material temperature would then probably result in decreased cooling flows, rather than in higher firing temperatures.

In this context it should be mentioned that research in the field of combustion in humid air is presently conducted at the Lund Institute of Technology, within the framework of the Centre of Competence in Combustion Processes [Torisson, 1999].

104 5. Rankine cycle

Here, the present-day technology for biomass based cogeneration will be presented briefly. Some calculations are also performed, using parameters similar to those that have been used for the IGCC and the IGHAT. The results can then be used as a reference when judging the performance of the other technologies.

5.1 Overview

In recent years a number of small to medium size cogeneration plants utilising biomass have been built in Sweden. Most of these plants consist of a solid fuel boiler and a steam turbine.

One such plant began operation in Vaxjo in 1997. The boiler is of circulating fluidised bed (CFB) type, delivered by Foster Wheeler Energia Oy, Finland. In the boiler, 43 kg/s steam at 142 bar/540°C is produced. A geared VAX turbine from ABB STAL, Sweden, is used to expand the steam, and the generator is rated 38 MWe. The district heating water is first heated in an exhaust gas condenser, and then in two heat exchangers in series, fed from the low- pressure turbine outlet and extraction, respectively [VEAB, 1998].

The exhaust gas condenser is very suitable for this type of plant, since the wet fuel in combination with a small amount of excess air give rather high exhaust gas dew temperature. Hence, a substantial amount of the latent heat for evaporation of the fuel moisture can be recovered in a condensing heat exchanger.

Through humidification of the combustion air, the dew temperature of the exhaust gas can be raised further, and this is also a method to lower the flue gas temperature to the stack.

Being a mature technology, the possibility of further development in terms of thermal performance is rather limited. Higher live steam pressure and temperature, and better turbine efficiency could, however, give some improvement.

5.2 Models

The configuration of the system studied here is shown in figure 5.1. Wet fuel is introduced into a CFB boiler. The combustion air is humidified and preheated to an extent dependent on the amount of sensible energy available in the exhaust gas. Water is evaporated in the boiler walls, and the remaining sensible energy in the exhaust gas is utilised in the superheater, in the economiser, in the aforementioned air preheater, in the condensing heat exchanger heating the cold return water from the district heating network, and finally in the humidifier.

The boiler is modelled using a combination of a RYlELD-reactor and a RGIBBS-reactor, similar to the model used for near atmospheric gasification (section 3.2.2). Here, the air flow for stoichiometric combustion is first calculated, and then an air factor is used to give the real air flow.

105 COND Air preheater

Humidifier : Air and exhaust gas : Biomass : Water/steam Generator Superheater ST Steam turbine ECO Economiser COND Condenser DH District heating heat exchanger

Figure 5.1. System layout of a cogeneration plant utilising a CFB boiler and a steam turbine.

The superheated steam is expanded in a steam turbine with three extractions for regenerative feed water heating in one open (feed water tank) and two closed preheaters. Final heating of district heating water is performed in two condensing heat exchangers connected in series.

The closed feed water heaters are modelled as regular heat exchangers without condensate coolers. A minimum temperature difference is also defined for these components.

5.3 Parametric study

Here, the influence of the live steam pressure on electrical efficiency and fuel utilisation is presented. Furthermore, the influence of the steam turbine isentropic efficiency is presented, since this parameter can benefit from the continuous development of more efficient blading.

In table 5.1, the most important parameters used in the calculations are presented. The values are similar to those that were used for the IGCC and the IGHAT, the major difference being the higher isentropic steam turbine efficiency. This is an attempt to take into consideration

106 that the steam turbine now is approximately twice as large as in the IGCC systems1. Also, the condenser pressure is now somewhat lower, resulting in increased electrical efficiency.

Boiler Fuel moisture content: 0.50 kg HaO/kg wet fuel Air factor: 1.20 Carbon conversion efficiency: 0.99 Heat loss: 1% of fuel HHV Air fan Poly tropic efficiency: 0.80 Superheater Steam temperature: 560°C Pressure drop steam side: Ap/p=0.05 Economiser Minimum temperature difference: ATmin=10°C Approach temperature: ATapp=10oC Water side pressure drop: Ap/p=0.15 Humidifier Air saturated at 50°C District heating water Supplytemperature: 90°C Return temperature: 50°C DH heat exchanger Minimum temperature difference: ATmin=10°C (exhaust gas condenser) DH heat exchanger Pressure: 0.85 and 0.50 bar respectively (steam condensers) 50% of total steam flow to each condenser. Closed feed water Minimum temperature difference: ATmjn=5°C preheaters Other heat exchangers Minimum temperature difference: ATmin=10°C (liquid/liquid and liquid/gas) Minimum temperature difference: ATmin=30°C (gas/gas) Steam turbine Isentropic efficiency12: 0.87 Mechanical efficiency: 0.99 Feed water tank Pressure: 1.5 bar Generator Electrical efficiency: 0.99 Steam turbine gear Mechanical efficiency: 0.98 Pumps Overall efficiency3: 0.713

Table 5.1. Assumptions used for calculation of cogeneration system based on CFB boiler and steam turbine.

In the calculations, the net electrical output is taken as the steam turbine generator output, minus the consumption in the air fan and the pumps shown in figure 5.1. No other consumption, e.g. for flue gas treatment and fuel handling, is included. Furthermore, no pressure drops have been included in the calculations of the feed water preheaters.

1 For the same fuel input. 2 Dry steam. 3Pis*T|m*TU

107 5.4 Results and discussion

5.4.1 Results of parametric study

Starting with the electrical efficiency, figure 5.2 shows an increase with increasing live steam pressure, as for the IGCC presented earlier. Compared with the IGCC, the absolute value of the electrical efficiency is, however, significantly lower.

(%) Electrical efficiency (LHV)

31 --

28 -- -

160 (bar)

Figure 5.2. Electrical efficiency for a Rankine cycle, as a function of live steam pressure.

In this system, the exhaust gas dew temperature is approximately 74°C. Hence, much of the latent heat in the exhaust gas moisture can be recovered in the condensing district heating heat exchanger (DH in figure 5.1) and in the air humidifier. The resulting fuel utilisation is presented in figure 5.3, and based on the lower heating value, a utilisation exceeding 100% is achieved. Here, it should be noted that the flue gas temperature leaving the system is 47°C. This is lower than in the IGCC and IGHAT cases, because the district heating water return temperature no longer limits the flue gas temperature.

(%) Fuel utilisation (LHV) 115

114

113

112

111

110 ■i------* 80 100 120 140 160 (bar)

Figure 5.3. Fuel utilisation for Rankine cycle, as a function of live steam pressure.

108 A brief investigation of the influence of steam turbine isentropic efficiency was also performed. Using a live steam pressure of 140 bar, the electrical efficiency for different values of isentropic efficiency is presented in figure 5.4.

Electrical efficiency (LHV) vs turbine isentropic efficiency

Figure 5.4. Electrical efficiency for the Rankine cycle, as a function of steam turbine isentropic efficiency.

5.4.2 Development needs

This system, representing a fairly mature technology, is included merely as a reference case. Hence, remaining development needs and potential performance improvements have not been studied in detail. Nevertheless, some brief comments on this subject are given below.

The live steam data used in these calculations are rather advanced, considering the relatively small plant size. This is especially true for the steam temperature, as will be discussed below. Despite this, the electrical efficiency is not very impressive. The fuel utilisation, on the other hand, is very high due to the condensation of a significant part of the water in the exhaust gas. At present, the price difference between electricity delivered to the grid and heat delivered to a district heating network is rather small (and sometimes even negative), and hence, a high fuel utilisation is often more important than a high electrical efficiency. In a possible future scenario when electricity is more highly valued, it may, however, be a problem that only a comparably small amount of electricity can be produced based on a given heat demand. In that case, a higher power to heat ratio would be desirable.

The steam temperature used in these calculations, 560°C, is higher than what is presently used in plants burning biomass. The main reason why high steam temperature is considered difficult in biomass boilers is the problem encountered with corrosion of superheater surfaces. Hjalmarsson and Kjork (1998) in their investigation concluded that the superheater material temperature and the ash melting point are of major importance for this phenomenon.

The material temperature depends not only on the steam temperature, but also, through the gas temperature, on the construction of the boiler. Examples are given of CFB boilers with steam temperatures approaching 540°C without severe corrosion problems. Better materials, careful design and placement of superheater surfaces, and better control of the combustion to avoid

109 hot spots and final combustion in the superheater, are suggested measures to control the corrosion problem. The choice of 560°C as the live steam temperature could then be interpreted as an attempt to include also some development of the more mature technology, when using it as a reference.

Also, the aforementioned development of more efficient blading must be emphasised, as a means to improve the electrical efficiency of this system.

no 6 Comparison of presented technologies

In the objectives of this report it was stated that similar assumptions and input data should be used for all studied concepts, in order to make a comparison among the presented technologies more reliable. Thus, for subsystems and pieces of equipment utilised in more than one system, the same calculation model and the same input data have been used.

Another objective was to calculate systems of comparable complexity, to assure that no concept is unduly favoured. However, since no definite measure of system complexity was defined, it is not possible to say with any certainty that this has been achieved. Starting from very simple layouts, bearing in mind that the scale for this type of plant is normally rather small, additional heat exchangers have in some cases been added, in order to decrease the heat losses. The limit for these adjustments is merely set by common sense, with the aim to create systems of reasonable overall complexity, when compared with competing concepts. To, at least, give a conception of the differences among the studied concepts, the number of heat exchangers and flow splits utilised in each system are presented in table 6.1. This could be interpreted as a crude estimate of the system complexity, although a lot of other important factors are not considered.

Concept Number of heat Number of exchangers flow splits IGCC, PG + ED 7 3 IGCC, PG + SD 8 5 IGCC, AG + ED 13 3 IGCC, AG + SD 14 5 IGHAT, PG + ED 5 4 IGHAT, PG + SD 6 3 IGHAT, AG + ED 10 5 IGHAT, AG+ SD 13 6 Rankine cycle + solid fuel boiler 9 4

Table 6.1. Heat exchangers and flow splits in various concepts (PG: pressurised gasification, AG: gasification at near atmospheric pressure, ED: exhaust gas dryer, SD: teamdryer).

Table 6.1 implies that the systems with near atmospheric gasification are more complex than the pressurised systems. This is also true when considering only the heat recovery system. However, the pressurised gasification itself is more complicated, due to the higher operating pressure and the extraction of air from the gas turbine compressor. Hence, the entire power plant systems are believed to be of comparable complexity. When comparing the IGCC with the IGHAT and the Rankine cycle, table 6.1 reveals no clear difference. There is, however, an inherent difference between the IGCC on one hand, and the IGHAT and Rankine cycle on the other; the number of gas and steam turbines. The IGCC, utilising both a gas and a steam turbine, is in this respect more advanced than the others. In addition to this, there are differences in the type of heat exchangers (media and temperature) and flow splits (media and type of equipment) listed in table 6.1. Hence, even if an effort has been made not to favour any concept more than others in this respect, there is still more work to do in the area of defining equal systems for a comparison like this one. Given these comments on the modelling of the various concepts, it is now time to turn to the result of the actual comparison.

Ill 6.1 Thermodynamic comparison

In chapters 3-5, the influence of a number of parameters were investigated for the various concepts. At the end of each chapter, the results from the parametric study were presented and discussed. For the IGCC and IGHAT, also the differences among the investigated variants of each technology (type of gasifier and dryer) were shown. For details, the reader is referred to these sections.

The results given below have already been presented in earlier sections, and the idea of this section is merely to facilitate the comparison of the different technologies. The electrical efficiency and fuel utilisation are shown in figures 6.1 - 6.4. The graphs for the IGCC represent the cases with combustor outlet temperature 1300°C and live steam pressure 100 bar. The steam dryers in both cases utilise steam from a turbine extraction. The IGHAT systems are also for COT 1300°C and the Rankine cycle results are for live steam pressure 140 bar. The district heating water supply temperature is in all cases 90°C.

(%) Electrical efficiency (LHV) IGCC and Rankine

45 --

SD + AG 35 - -- Rankine \ : 30 —

25 (PR)

Figure 6.1. Electrical efficiency for IGCC systems, as a function of gas turbine compressor pressure ratio. Result for Rankine cycle included as reference. Abbreviations as in table 6.1.

It can be seen in figures 6.1 and 6.2, that compared with the Rankine cycle (solid fuel boiler and steam turbine) the new technologies are superior in terms of electrical efficiency. This is a result of the incorporation of a gas turbine in the power plant concept. Comparing the IGCC and the IGHAT, the former shows slightly higher efficiencies, and also more pronounced optima of gas turbine compressor pressure ratio.

112 -Electrical efficiency (LHV) IGHAT and Rankine

ED + AG

35 (PR)

Figure 6.2. Electrical efficiency for IGHAT systems, as a junction of gas turbine compressor pressure ratio. Result for Rankine cycle included as reference. Abbreviations as in table 6.1.

Considering the fuel utilisation, the conventional Rankine cycle shows the best result, due to high exhaust gas dew temperature and extensive condensation.

(%) Fuel utilisation (LHV) IGCC and Rankine

Rankine';.;

SD + AG ;

: ED^PG ---

25 (PR)

Figure 6.3. Fuel utilisation for IGCC systems as a function of gas turbine compressor pressure ratio. Result for Rankine cycle included as reference. Abbreviations as in table 6.1.

Comparing figures 6.3 and 6.4, the IGHAT systems exhibit lower values of fuel utilisation than do the IGCC systems. In this context it should be emphasised that the fuel utilisation of the IGHAT is very sensitive to the district heating supply temperature. Allowing for lower

113 supply temperature results in values of fuel utilisation comparable to the IGCC systems, as presented in sections 4.3.1 - 4.3.4.

(%) Fuel utilisation (LHV) IGHAT and Rankine

110 - - - 1 oo - -

90 --

70 --

ED+AG

35 (PR)

Figure 6.4. Fuel utilisation for IGHAT systems as a function of gas turbine compressor pressure ratio. Result for Rankine cycle included as reference. Abbreviations as in table 6.1.

To conclude, the new technologies offer a potential for better electrical efficiency, at the expense of lower fuel utilisation. When comparing cogeneration plants, the mutual importance of electrical efficiency and fuel utilisation depends on the valuation of electricity and heat at the time of comparison. Here such an economic evaluation is not performed. It is, however, believed that the system descriptions and the resulting figures of electrical efficiency and fuel utilisation presented here can be used for such an evaluation, if proper cost data is at hand.

6.2 Technological comparison

Aside from the differences in thermodynamic performance, there are also other differences among the studied technologies. One issue is the need for technological development of key components in the various concepts. This has been treated in earlier chapters, and here only brief comments on the development needs of two major subsystems, present in both the IGCC and the IGHAT, will be given.

A common feature of the IGCC and the IGHAT is the gas cleaning equipment. The possibility of cleaning the relatively small flow of fuel gas, instead of the larger flue gas flow, is considered as a major advantage of the gasification concept compared with direct combustion of biomass. Cold gas cleaning, utilising a wet scrubber as a final stage, is fairly well proven and particularly suited to near atmospheric gasification. The remaining problematic issue then, is the tar cracking, required in order to avoid filter clogging and withdrawal of tars (and hence loss of chemical energy).

114 To fully utilise the potential advantages of pressurised gasification, hot gas cleaning should be provided instead. Development of both ceramic and metallic hot gas filters is underway, and the results so far are promising. A more difficult problem could be the NOx emissions, resulting from the nitrogen in the fuel. Alternative concepts for solving this problem have been reviewed in section 3.2.3, but further research and development seems necessary. If no other technique works, flue gas treatment, e.g. SCR, could be used. This is an established but costly method.

Combustion of low LHV gas in modified gas turbine combustors has been tested successfully and should not constitute a major problem. To achieve the efficiencies presented in this report, some changes have to be made to existing gas turbines, to accommodate the larger mass flow through the turbine. This is especially important for the systems with gasification at near atmospheric pressure, and for all IGHAT systems. For the IGHAT, there is also the question of stable combustion in humid air. This could limit the amount of water evaporated into the combustion air and, hence, lower the obtainable electrical efficiency.

Besides the development of critical components, also development of reliable systems, enabling high availability, is of great importance. As for all new technologies, satisfactory plant availability still has to be proved for systems based on biomass gasification. For the IGCC, recent results from coal based plants indicate that reasonable availability can be achieved. The IGHAT systems are in many parts similar and should be able to reach the same values.

Another important issue is the suitability of these cogeneration technologies for smaller scales, enabling the utilisation of smaller district heating networks as heat sinks. The general trend is for all technologies, that the specific investment cost increases with decreasing plant size, and generally this is the limiting factor. In this respect, the IGHAT may have an advantage over the IGCC, since the investment cost should decrease when the steam cycle is excluded. This still remains to be shown in practice.

The main technical constraint for moving towards smaller scale, is the efficiency of the steam cycle. Decreased steam flow means lower turbine efficiency, unless the steam pressure is also decreased to increase the volume flow. Decreasing the steam pressure, however, decreases the thermodynamic efficiency of the Rankine cycle. Comparing an IGCC and a Rankine cycle, producing the same amount of heat for district heating, the electrical output of the latter is much smaller, and the resulting steam turbine output is only 10 - 20 % larger than for the IGCC. Hence, both technologies are almost equally suited for smaller scales in this respect. The IGHAT does not use a steam turbine and hence does not suffer from this constraint.

6.3 Concluding remarks

Having performed this comparison, what are the conclusions? Starting with the thermodynamic performance, the systems utilising gasification exhibits higher electrical efficiency and lower fuel utilisation, compared with the more conventional cogeneration plant based on CFB boiler, steam turbine and exhaust gas condensation. Hence, the new technologies would favour from an increased difference in price between electricity and heat.

115 An economic comparison of the studied technologies is outside the scope of this work, but the estimated cost of electricity presented for the IGCC (section 3.4.2) is well above the present market price of electricity in Sweden. Hence, there is probably no economic incentive for building cogeneration plants based on gasified biomass, unless the price of electricity is increased or the capital cost of the plants is decreased. Furthermore, there are, as presented above, still some technological uncertainties regarding components in both the IGCC and the IGHAT, which have to be handled before these technologies are ready to enter the market.

Nevertheless, if one believes that there will, eventually, be a market for these technologies, where should the research and development efforts be directed? Firstly, there is more to be done on the system level, such as calculation of dynamic behaviour and part load performance. This would complete the design point calculations performed in this work. Also, further sensitivity analyses are required, to identify critical components with significant influence on plant performance.

Once this is done, further development of the critical components is required. Examples of required component development have already been presented in sections 3.4.4 and 4.4.2 for IGCC and IGHAT, respectively. It should be noted here that development may be required not only to improve the thermodynamic performance, but also for environmental or economical reasons.

In the research and development work, both theoretical and experimental methods will be used. As an example of the former, computational fluid dynamics (CFD) will be useful in designing components with low pressure drops and/or good mixing. Similarly, new findings in material science, often based on experimental tests, are probable to be of great value when developing various components.

Besides the technical development of systems and components, development of manufacturing methods is vital, in order to lower the capital cost of the plants.

Furthermore, the method of judging the complexity of different technologies has to be developed further, to ascertain fair comparisons. The course taken here, focusing on the number of certain pieces of equipment, is a first attempt but probably not sufficiently accurate. Another approach could be to avoid the term ’’complexity” and instead focus on cost data for the studied technologies, comparing systems with approximately equal specific investment costs. There is, however, a risk of introducing additional sources of error when this approach is used, especially when new technologies are studied.

To conclude, there are a number of promising concepts for cogeneration, based on gasified biomass, which are likely to be developed further. It should, however, be remembered that also competing technologies may improve, as a result of research and development. Hence, the targets of the development may have to be changed to even higher efficiencies and/or lower investment costs than what present-day market studies indicate.

116 7. Acknowledgements

Many people have in different ways contributed to this report, and to all of them I am very grateful.

Especially I would like to thank my supervisor, professor Tord Torisson for his support and encouragement throughout this work. I also wish to thank professor Lars Sjunnesson, M.Sc. Eva K Larsson, M.Sc. Krister Stahl and M.Sc. Pontus Steinwall for sharing with me their knowledge in this field.

Financial support has been given from the Swedish National Energy Administration and this is greatly acknowledged.

Last, but not least, I would like to thank my family for always being there.

Fredrik Olsson References

1. Andersson, P. 1989. Drying of wood fuels in air and superheated steam, Licentiate Dissertation, Lund Institute of Technology, Lund, ISSN 0346-8054. 2. ASPEN Technology. 1995. Reference Manual, Release 9.2, ASPEN Technology Inc., Cambridge, Massachusetts. 3. Bain R.L., Overend R.P. and Craig K.R. 1996. Biomass-fired power generation, presented at Biomass usage for utility and industrial power, Snowbird resort and conference centre, Snowbird, UT, April 29 - May 3,1996. 4. Barbucci P., Neri G. and de Lange H. 1994. The feasibility of electricity production from biomass by gasification systems. Task 3: Italian case studies, JOULE II program, contract no JOU2-CT92-0226, Final Report. 5. Beenackers A.A.C.M. and Maniatis K. 1996. Gasification technologies for heat and power from biomass, pp. 228-259 in Biomass for Energy and the Environment, Proceedings of the 9th European Bioenergy Conference, Copenhagen, Denmark 24-27 June 1996, Volume 1, Pergamon, ISBN 0 08 0428 495. 6. Bergmann H., Chalupnik R. and Jacobs J. 1997. Entwicklungsstand von Kombi- Kraftwerken auf Kohlebasis, VGB Kraftwerkstechnik 77, Heft 3, pp. 166-173 (in German). 7. Bjerle I. et al. 1995. Final report - Bioforgasning etapp 2 LTH, Department of Chemical Engineering II, Lund Institute of Technology, Lund. 8. Bridgwater A.V. 1995. The technical and economic feasibility of biomass gasification for power generation, Fuel, Vol 74, No 5, pp 631-653. 9. Briesch M.S., Bannister R.L., Diakunchak I.S. and Huber D.J. 1995. A Combined Cycle Designed to Achieve Greater Than 60 Percent Efficiency, Journal of Engineering for Gas Turbines and Power, Vol. 117, October 1995, pp. 734-741. 10. Buekens A.G. and Schoeters J.G. 1985. Modelling of biomass gasification, pp 619-689, in Fundamentals of thermochemical biomass conversion, ed. by Overend R.P., Milne T.A. and Mudge L.K., Elsevier Applied Science Publishers Ltd., London, ISBN 0-85334- 306-3. 11. Bohm H. 1994. Fossil-fuelled power plants, VGB Kraftwerkstechik 74 (1994), Number 3, pp 151-164. 12. Cavani A. 1998. Modified humid air turbine cycle for biomass gasification, 98-GT-233, presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Stockholm, Sweden, June 2-June 5, 1998. 13. Chellini R. 1998. Low BTU Gas for Combined-Cycle Plant, Diesel & Gas Turbine Worldwide, June 1998, pp. 12-15. 14. Chiesa P., Lozza G., Macchi E. and Consonni S. 1995. An Assessment of the Thermodynamic Performance of Mixed Gas-Steam Cycles: Part B - Water-Injected and HAT Cycles, Journal of Engineering for Gas Turbines and Power, Vol. 117, July 1995, pp. 499-508. 15. Cohen H., Rogers C.F.C. and Saravanamuttoo H.I.H. 1996. Gas Turbine Theory, 4th edition, Longman Group Ltd., Harlow, ISBN 0-582-23632-0. 16. Consonni S. and Larson E.D. 1996a. Biomass-gasifier/aeroderivative gas turbine combined cycles: Part A - Technologies and performance modelling, Journal of Engineering for Gas Turbines and Power, July 1996, vol. 118, pp 507-515.

118 17. Consonni S. and Larson E.D. 1996b. Biomass-gasifier/aeroderivative gas turbine combined cycles: Part B - Performance calculations and economic assessment, Journal of Engineering for Gas Turbines and Power, July 1996, vol. 118, pp 516-525. 18. Coulson J.M. and Richardson J.F. 1991. Chemical engineering, vol. 2, fourth edition, Pergamon Press pic., Oxford, ISBN 0-08-037956-7. 19. Craig K.R., Mann M.K. and Bain R.L. 1994. Cost and performance potential of advanced integrated biomass gasification combined cycle power systems, pp 641-654, IGTI-Vol. 9, ASME COGEN-TURBO 1994. 20. Dawes S.G., Mordecai M., Welford G.B. and Otter N.R. 1997. Recent design and cost studies for air blown gasification, pp. 347-358 in Proceedings of the 14th international conference on fluidized bed combustion, edited by Preto F.D.S., the American Society of Mechanical Engineers (ASME), New York, ISBN 0-7918-1557-9. 21. DeMoss T. 1996. They’re he-e-re (almost): the 60% efficient combined cycle, Power Engineering, July 1996, pp. 17-21. 22. Double J.M., Smith E.L. and Bridgwater A.V. 1989. Computer modelling of fluidized bed gasification, pp 651-655, in Pyrolysis and gasification, Ferrero et al., Elsevier Applied Science, London, ISBN 1-85166-449-1. 23. Energimyndigheten. 1998. Prisbladfor biobranslen, torvmm, information leaflet from the Swedish National Energy Administration (in Swedish). 24. Faaij A.P.C. 1997. Energy from biomass and waste, Ph.D. thesis, Utrecht University, Utrecht, ISBN 90-393-1592-2. 25. Fyhr C. 1996, Superheated steam drying of wood chips in pneumatic conveying dryers, Doctoral Thesis, Chalmers University of Technology, Goteborg, ISBN 91-7197-336-2. 26. Gustavsson L., Borjesson P., Johansson B. and Svenningsson P. 1995. Reducing CO2 emissions by substituting biomass for fossil fuels, Energy, vol. 20, no. 11, pp. 1097-1113. 27. Gustavsson L. and Johansson B. 1994. Cogeneration: one way to use biomass efficiently, Heat Recovery Systems & CHP, vol. 14, no. 2, pp. 117-127. 28. Hasselbacher H. 1989. Gas Turbines, chapter 2.5, pp 2.141-2.143 in Standard handbook of power plant engineering, edited by Elliot T.C., McGraw-Hill, New York, ISBN 0-07- 019106-9. 29. Hermansson M. 1992. Steam drying - an experimental and theoretical study, Licentiate Dissertation, Lund Institute of Technology, Lund, ISSN 0346-8054. 30. Hjalmarsson L. and Kjork A. 1998. Korrosion i overhettare vid biobrdnsleeldning. (Corrosion in superheater when firing biomass fuels.), Report No. 623, Varmeforsk, Stockholm, ISSN 0282-3772 (in Swedish). 31. Holt N.A.H. 1998. Commercial IGCC demonstration plants in the USA - status report May 1998, Presented to the VGB Technical Committee on Fuel Conversion, Madrid, Spain, June 2-3 1998. 32. Hughes W.E.M and Larson E.D. 1997. Effect of fuel moisture content on biomass-lGCC performance, 97-GT-4, presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Orlando, Florida, June 2 - June 5, 1997. 33. Hulkkonen S. 1994. The feasibility of electricity production from biomass by gasification systems. Task 4: Finnish case studies, JOULE II program, contract no JOU2-CT92-0226, Final Report. 34. Ihren N. 1991. Torkning av skogsbranslen - teknisk jamforelse av torkmetoder vid koppling tillfdrgasning och elproduktionfran kombicykel, FUD-Rapport 1991:30, Vattenfall Utveckling AB, Vallingby, ISSN 1100-5130 (in Swedish). 35. ISO 11086. 1996. Gas turbines - vocabulary, Reference number ISO 11086: 1996(E/F).

119 36. Jansson L-E. 1997. Metoderfor okat elutbyte i befintliga hetvattenanldggningar. (Methods for increasing power generation in existing hot-water boiler plants.) Report No. 619, Varmeforsk, Stockholm, ISSN 0282-3772 (in Swedish). 37. Jeffs E. 1996. Korean combined cycle is commercial launch ofGT24, Turbomachinery International, March/April 1996, pp 23-26. 38. Jeffs E. 1998. Baoshan combined cycle running on blastfurnace gas, Turbomachinery International, March/April 1998, pp 18-20. 39. Johansson T.B., Williams R.H., Ishitani H. and Edmonds J.A. 1996. Options for reducing CO2 emissions from the energy supply sector, Energy Policy, Vol. 24 Nos. 10/11, pp. 985-1003. 40. Kehlhofer R. 1978. Calculation for Part-Load Operation of Combined Gas/Steam Turbine Plants, Brown Boveri Review, 10-78, pp 672-679. 41. Kjaer S. 1993. Die zuktinftigen 400-MW-ELSAM-Bldcke in Aalborg und Skaerbaeck. VGB Kraftwerkstechik 73 (1993), Heft 11, pp 933-940 (in German). 42. Kristiansen A. 1996. Understanding coal gasification, IEA Coal Research, London, ISSN 92-9029-267-9. 43. Kurkela E. and Solantausta Y. 1994. The feasibility of electricity production from biomass by gasification systems. Task 2: General process evaluation, JOULE II program, contract no JOU2-CT92-0226, Final Report. 44. Larson E.D. and Marrison, C.I. 1997. Economic Scales for First-Generation Biomass- Gasifier/Gas Turbine Combined Cycles Fueled From Energy Plantations, Journal of Engineering for Gas Turbines and Power, Vol. 119, April 1997, pp. 285-290. 45. Larson E.D., Kreutz T.G. and Consonni S. 1998. Combined biomass and black liquor gasifier/gas turbine cogeneration at pulp and paper mills. 98-GT-339, Presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Stockholm, Sweden, 46. June 2-June 5, 1998. 47. Larsson E.K. 1998. TPS, Studsvik, personal communication. 48. Layne A.W. and Hoffman P.A. 1998. The U.S. department of energy's advanced turbine systems program, 98-GT-141, Presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Stockholm, Sweden, 49. June 2-June 5, 1998. 50. Leppalahti J., Kurkela E., Simell P. and Stahlberg P. 1994. Formation and removal of nitrogen compounds in gasification processes, in Advances in Thermochemical Biomass Conversion, vol. 1, edited by Bridgwater A.V., Chapman & Hall, London, ISBN 0 7514 0172 2. 51. Liinanki, L. and Karlsson, G. 1994. VEGA test & verifikation - trycksatt forgasning av biobrdnslen, (VEGA test & verification - pressurised gasification of biomass) Rapport 1994/12, Vattenfall Utveckling AB, Stockholm, ISSN 1100-5130 (in Swedish). 52. Lindhardt S. and Lind-Hansen A. 1994. I/S Vestkraft, Block 3, Esbjerg. VGB Kraftwerkstechik 74 (1994), Heft 1, pp 15-24 (in German). 53. Magnusson, L. 1993. Uppfoljning av erfarenheterfran anlaggningarfor torkning av biobrdnslen, Stiftelsen for varmeteknisk forskning, Stockholm (in Swedish). 54. Mayer K.H., Bendick W., Husemann R.-U., Kem T. and Scarlin R.B. 1998. New materials for improving the efficiency of fossil-fired thermal power stations. VGB PowerTech, Volume 78/1998, Number 1/98, pp 22-27, ISSN 0372-5715. 55. Mujumdar A.S. 1995. Handbook of industrial drying, vol. 1, Marcel Deker Inc., New York, ISBN 0-8247-8996-2.

120 56. Nakhamkin M., Swensen E., Paprotna H., Yankowich P., Marchionna N. and Cohn A. 1994. Combustion studies of natural gas and syn-gaswith humid air, Electric Power Research Institute, Palo Alto, CA, TR-103382. 57. Neergaard M. 1998. Sydkraft, Malmo, personal communication. 58. Neilson C.E., Shafer D.G. and Carpentieri E. 1998. LM 2500 Gas Turbine Fuel Nozzle Design and Combustion Test Evaluation & Emissions Results with Simulated Gasified Wood Product Fuels, 98-GT-339, presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Stockholm, Sweden, June 2-June 5,1998. 59. Newby R.A., Lippert T.E., Alvin M.A., Brack G.J. and Sanjana Z.N. 1998. Status of Westinghouse hot gas filters for coal and biomass power systems. 98-GT-340, presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Stockholm, Sweden, June 2-June 5,1998. 60. Nilsson P-O., Hektor B., Lonner G. and Parikka M. 1996. Tradbrdnslepotential i sodra Sverige (Wood fuel potential in southern Sweden), Rapport 1995/10, Vattenfall Utveckling AB, Vallingby, ISSN 1100-5130 (in Swedish). 61. NUTEK. 1997. Energilaget 1997. (Energy in Sweden 1997). Info 437-97, Narings- och teknikutvecklingsverket, Stockholm (in Swedish). 62. Nyren C. 1992. Miljdejfekter vid biobransletorkning. (Environmental effects from drying of biomass.), Report from Vattenfall Research, Vallingby, ISSN 1100-5130 (in Swedish). 63. Praschek R., Haupt G. and Ullrich N. 1998. Are 1GCC power plants competitive ?, presented at the VGB-KongreB ’’Kraftwerke 1998”, StraBburg, 12.-17.10.1998. 64. Reed T.B. 1981. Biomass Gasification, Principles and Technology, Energy Technology Review No. 67, Noyes Data Corporation, Park Ridge, New Jersey, ISBN 0-8155-0852-2. 65. Rensfelt E.K.W. 1997. Atmospheric CFB gasification - the Greve plant and beyond. presented at the International Conference on Gasification and Pyrolysis of Biomass, 9-11 April 1997, Stuttgart, Germany. 66. Rosen C., Bjombom E., Yu Q. and Sjostrom K. 1997. Fundamentals of pressurised gasification of biomass, pp. 817-827 in Developments in thermochemical biomass conversion, edited by Bridgwater A.V. and Boocock D.G.B., Chapman & Hall, London, ISBN 0 7514 0350 4. 67. Rosen P. 1993. Evaporative gas turbines, Licentiate Thesis, Lund Institute of Technology, Lund, ISSN 0282-1990. 68. Ryding B. 1993. Biobransleeldad krafivarme i ett stortfjdrrvarmesystem: systemstudier med energisystemmodellen Martes, Rapport fran Vattenfall Research, 1993:3, Vattenfall, Vallingby (in Swedish). 69. Salo K., Horvath A. and Patel J. 1998. Pressurised gasification of biomass, 98-GT-348, presented at the International Gas Turbine & Aeroengine Congress & Exhibition, Stockholm, Sweden, June 2-June 5, 1998. 70. Simell P., Kurkela E. and Stahlberg P. 1994. Formation and catalytic decomposition of tars from fluidised bed gasification, in Advances in Thermochemical Biomass Conversion, vol. 1, edited by Bridgewater A.V., Chapman & Hall, London, ISBN 07514 0172 2. 71. Sjunnesson L. 1999. Sydkraft AB, Malmo, personal communication. 72. SOU 1992:90. 1992. Biobranslenforframtiden. Allmanna forlaget, Stockholm, ISBN 91- 38-13162-5. 73. Stambler 1.1997. Oil refinery IGCCs turning waste feedstock into product and power, Gas Turbine World, November-December 1997, pp. 30-34.

121 74. Steinwall P. 1995. Integrering av torkning ochforgasning av biobrdnslen med den evaporativa gasturbincykeln BIO-IGHAT, Department of Heat and Power Engineering, Lund Institute of Technology, Lund, ISSN 0282-1990 (in Swedish). 75. Stahl K. and Neergaard M. 1996. IGCC power plant for biomass utilisation Vamamo, Sweden, VGB Kraftwerkstechik 76 (1996), Number 4, pp 306-308. 76. Stahl K. 1997. Vamamoverket - en demonstrationsanlaggning for el- och varmeproduktion ur biobrdnsle, baserad pa trycksattforgasning, report from Sydkraft, Mahno (in Swedish). 77. Stahl K. 1998. Sydkraft AB, Malmo, personal communication. 78. Traupel W. 1988. Thermische Turbomaschinen, vol. 1, 3rd ed., Springer-Verlag, Berlin, ISBN 3-540-07939-4 (in German). 79. Torisson T. 1999. Dept, of Heat and Power Engineering, Lund Institute of Technology, Lund. Personal communication. 80. van Ree R. 1994. Biomass gasification: a ‘new’ technology to produce renewable power, an inventory of technologies, Netherlands Energy Research Foundation ECN, ECN-CX— 94-057, Petten. 81. VEAB. 1998. Information brochure from Vaxjo Energi AB (in Swedish). 82. Weinzierl K. 1994. Kohlekraftwerke derZukunft, VGB Kraftwerkstechik 74 (1994), Heft 2, pp 109-114 (in German). 83. Wimmerstedt R. and Hallstrom A. 1984. Torkning av torv och biobrdnslen. Teknik, ekonomi, utvecklingsbehov. (Drying of peat and biofuels. Techniques, economy and development needs.) Report LUTKDH/(TKKA-3002)/1-117/(1984), Department of chemical engineering, Lund Institute of Technology, Lund, ISSN 0346-8054 (in Swedish). 84. Wimmerstedt R. 1995. Drying of peat and biofuels, in Handbook of industrial drying, vol. 2, edited by Mujumdar A.S., Marcel Deker Inc., New York, ISBN 0-8247-9644-6. 85. Wimmerstedt R. and Linde B. 1998. Analys av det tekniska och ekonomiska lagetfor torkning av biobrdnslen. (Assessment of Technique and Economy of Biofuel Drying). (in Swedish). Report No. 637, Varmeforsk, Stockholm, ISSN 0282-3772 (in Swedish). 86. Yan J. 1998. Biomass gasification power generation technologies, Draft, Department of Chemical Engineering and Technology, Royal Institute of Technology, Stockholm, ISSN 1104-3466. 87. Agren N. 1997. Simulation and design of advanced air-water mixture gas turbine cycles, Department of Chemical Engineering and Technology, Royal Institute of Technology, Stockholm, ISSN 1104-3466.

122 Appendix I

Calculation sequence and process flow diagram

In the following pages the calculation sequence and the process flow diagram, as it appears in ASPEN PLUS™, for an IGCC system with pressurised gasification and steam dryer are presented.

The calculation sequence shows the number of nested iteration loops required in the calculation. In short these iteration loops serve the following purposes:

CONT: Converges a number of tear streams to ensure overall mass and energy balances are satisfied. GTPOWER: Converges the power output from the gas turbine. AIRFLOW: Converges a design specification where the air flow trough the gas turbine compressor is varied to achieve a pre-set combustor outlet temperature. AIRFEED: Converges a design specification where the amount of air fed to the gasifier is varied to achieve a pre-set operating temperature. DRUMFLOW: Converges the flow of water/steam in the HRSG. DRYFLOW: Converges a design specification where the amount of steam fed to the dryer is varied to achieve a pre-set dryer condensate outlet temperature. TURBPR: Converges a design specification where the HPST model outlet pressure is varied until steam is saturated at outlet. EXTR: Converges a design specification where the amount of steam bled to the feed water tank is varied until the feed-water leaving the tank is saturated. FVFLOW: Converges a design specification where the flow of district heating water to the condenser (CONDENS) is varied to achieve the pre-set supply temperature after heat exchanger B6. FVSPLIT: Converges a design specification where the fraction of total district heating water flow led to the fuel moisture condenser (B20) is varied until the pre­ set supply temperature after B20 is achieved.

A bold F denotes a FORTRAN block used for specifying or calculating input data to a unit operation model (capital letters). The bold T denotes a transfer block transferring stream data (flow, temperature and/or pressure) from one stream to another.

The process flow diagram shows the system layout and some results from the calculations. Other results, not shown, are for example electrical efficiency and fuel utilisation. These are calculated in the FORTRAN block F:Rep and presented separately.

1(3) Calculation sequence.

F:Deltap-l F:Fan FrHeatval F:Water DRY-REAC Begin CONT Bll STORAGE LOCKHOPP F:Hlossgas HPPUMP Begin GTPOWER Begin AIRFLOW B1 GTKOMP Begin AIRFEED FrCoolairl F:Coolair2 AIRSPLIT BIO F:Gasifpre BOOSTCOM FrGasify DECOMP Filnertflo F:Coaleff HETRREAC B5 HEATLOSS GASREAC PARTSEP Return AIRFEED GASEVAP FILTER F:Pbk BK Return AIRFLOW B2 MDC1F FrExppart F:Etaisl STEG1 MDC2F F:Etais2 STEG2 F:Etais3 STEG3 F:Etais4 STEG4 Return GTPOWER FzPumppres FzHpsttemp B12 Begin DRUMFLOW HPEC02 B3 GASEVAP HPEV HPDRUM HPSH Return DRUMFLOW HPT Begin DRYFLOW B7 B8 DRYHX Return DRYFLOW DRY-FLSH B9 B15 DRYPUMP Begin TURBPR HPST Return TURBPR DPST Begin EXTR B23 FzSteff LPST Begin FVFLOW CONDENS Begin FVSPLIT FVSPLI B20 Return FVSPLIT B6 Return FVFLOW KONDPUMP FEEDTANK Return EXTR TzFeedwatr Return CONT GB6 F:Rep

2(3) w UJ Appendix II

Cost of heat from boiler plant

To estimate the value of the heat produced in the IGCC cogeneration plant (section 3.4.2) the cost of producing the same amount of heat in a biomass fuelled CEB boiler is calculated. Here also the capital cost is included, i.e. two possible investments, a cogeneration plant or a boiler plant, are compared. This corresponds to a situation where the heat load is increased beyond the capacity of the existing plants or when an old plant is to be taken out of service and replaced.

If there instead was additional capacity available in an existing boiler, the marginal cost of producing additional heat in that plant could be used to estimate the value of the heat in the IGCC calculations.

As in the case of IGCC there is also for these boiler plants a large variation in investment and O&M costs reported in the literature. The following assumptions, mainly based on [Gustavsson et al., 1995; Gustavsson and Johansson, 1994; Ryding, 1993] have been used for the boiler plant, regardless of size:

Fuel utilisation (LEV): 1.0 (with flue gas condensation) Specific investment cost: 3000 SEK/kWth O&M cost, fixed: 2% of investment per year variable: 0.02 SEK/kWhth Fuel cost: same as for cogeneration plant Interest rate: same as for cogeneration plant Plant life: same as for cogeneration plant

As for the IGCC plant the annuity method is used to calculate the annual capital cost.

KD Appendix III

T,Q-diagram for IGHAT systems with different district heating supply temperatures

Here the difference between the two IGHAT systems is shown in T,Q-diagrams. The first constraint, valid in both cases, is a minimum temperature difference in the exhaust gas condenser, ATmin. In the case of maximum heat recovery the exhaust gas temperature out of the heat exchanger is set to Tout = 60°C. The district heating water return temperature is TDH,m = 50°C and for a given exhaust gas inlet temperature T;n the resulting district heating supply temperature TDH,out can be calculated. The resulting T,Q-diagram is shown in the left part of figure 1.

Figure 1. T,Q-diagrams for IGHAT systems with different district heating water supply temperature.

In order to raise the district heating supply temperature Ton,out to 90°C, the economiser (ECO) is replaced by a second heat exchanger for district heating water (DH2). The amount of sensible energy in the exhaust gas is however limited and hence the flow of district heating water must be decreased simultaneously in order to achieve 90°C supply temperature. This results in higher exhaust gas temperature, less condensation of water in the exhaust gas and correspondingly smaller heat output. The change in temperature levels and heat recovery can be seen in the right part of figure 1.

KD