EXTERNALLY FIRED GAS TURBINE CYCLE-0.64KWEL FOR ROTATIONAL SPEED OF 120000RPM

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Micro Gas Turbine and Biomass

To be feasible for small scale bio-energy power plants, MGTs are commercially available in the power range of 30-200 KWel [Matteo et al; 2008]. MGT cycles have also additional benefit of utilizing exhaust hot gases from MGT as a source of heat for the gasifier for assisting the endothermic reaction to take place.
Biomass gasification is an appropriate tool to use biomass in environmental friendly ways. Small scale gasification, employing downdraft gasifier, is a promising technology towards utilization of small and distributed potential of biomass with the ability to supply the producer gas with a low tar contents (0.015-0.5g/Nm3) [Matteo et al; 2008]. The authors have reported that small scale gasifiers (upto 500kW th) are usually based on downdraft technology with satisfactory conversion efficiency. When coupled with MGT also provides the opportunity for co-generation with available exhaust gas heat at medium temperature for biomass drying.
Rabou et al. (2007) have investigated the operation of standard Capstone 30kWel MGT without modifications, with varying composition of producer gas mixed with natural gas. The authors have reported the possibility of operation of MGT with 100% biomass producer gas with heating value of 6MJ/Nm3 (in our case producer gas heating value is around 5.22 MJ/Nm3). However, insufficient combustion stability was observed for operations longer than one hour when running with 100% producer gas. Stable combustion was obtained when a small amount of natural gas was mixed with producer gas with resulting heating value of 8MJ/Nm3. The authors have therefore concluded that the Capstone MGT would run reliably with a producer gas but with slightly higher heating value. A negligible effect has been observed on the gross efficiency of the MGT with reduction in fuel heating value, which is due to higher power demand for compression of larger fuel volume to get the same heat input. Figure 1.6 below presents the results of the experiments conducted; where the power output considerably decreases with increasing contribution of producer gas [Rabou et al; 2007].
Matteo et al. (2008) have investigated on a MGT fed by blends of biomass producer gas and natural gas. For pure producer gas case, the authors have reported an electrical output of 30% lower than for the nominal condition with natural gas feeding. This is due to higher power consumption for the producer gas compression, whereas an increase of 32.9% of thermal output is estimated which is mainly due to increase in exhaust gas temperature.
Barros and Pericles (2004) have analyzed the use of biomass in gas turbine and have reported that gas turbine cycles are most suitable for biomass fuels. Due to gas turbine engines being very much sensitive to fuel quality one of the option is to gasify the biomass. However, in order to utilize gas from the gasifier there is need for cleaning the gas to get rid of unwanted contents like tars, particulates, alkalies and heavy metals. Clogging in valves and filters occurs due to tars condensing on cool surfaces. Alkalis such as Na and K cause corrosion of turbine blades whereas particulates are more critical and even small amount of particulates cause erosion problems [Matteo et al; 2008].
For biomass producer gas having low heating contents, higher fuel flow rate is required which leads to larger size of turbine for the same power output as compared to natural gas based turbine. Also fuel compression is required in order to make it able to be injected in the combustor. Larger flow rate passing through the turbine leads to an increase of the pressure ratio with associated risk of surge.

Internally Fired Micro Gas Turbine

Direct and indirect firing techniques exist for utilizing biomass in the gas turbine. Under the direct firing techniques, one is Pressurized Fluidized Bed Combustion (PFBC) and the other is Biomass Integrated Gasification Gas Turbine (BIGGT). Figure 1.7, below shows the BIGGT technique wherein gasified gases are directly fired in the internal combustor.
Bohn and Lepers (2003) have investigated the effects of biogas combustion on the operation characteristics of an MGT and have reported that replacing the natural gas with biogas results in decreased turbine efficiency due to the following main reasons.
i) Lower gas temperature at combustor inlet due to higher biogas flow rate which is at lower temperature as compared to the air preheated in the recuperator.
ii) Large exhaust losses due to larger share of CO2 having higher specific heat.

Externally Fired Micro Gas Turbine

The indirectly fired gas turbine (IFGT) also called externally fired gas turbine (EFGT) is well suited for direct fuel burning techniques without employing the gasification systems. However, corrosive fuels put difficulties for high temperature heat exchangers. EFGT have a disadvantage of limitations on maximum TIT, due to heat exchanger material incapability of withstanding at the higher temperature, which results in decreased efficiency as reduced TIT has drastic impact on the power output and the efficiency. Figure 1.8 presents the schematic of simple EFGT.
Traverso et al. (2005) have reported that the inlet temperature for heat exchangers cannot go beyond 800 0C which results in limiting the TIT under 750 0C. Chiaramonti et al. (2004) have reported that, for heat exchangers, even with advanced materials the maximum allowable temperature for the entering flow is 1053 K. In order to get highly efficient EFGT cycles the use of super-alloys has not been recommended as 900 0C is the maximum temperature limit for the materials and therefore use of ceramics based heat exchangers have been seen as a most promising option [Barros and Pericles; 2004].
As regards the advantages of EFGT, it offers fuel flexibility with working fluid as clean air resulting in increased engine lifetime. Diesel, landfill gas, industrial off-gases, ethanol, and bio-based liquids and gases are among the other fuel options in addition to the natural gas [Pilavachi; 2002]. Indirect firing technique also allows the biomass to combust in external combustor with the advantage of biomass usage without employing the complex gasification systems. However, the cycle requires an addition of solid fuel burner and the heat exchanger as compared to standard natural gas based turbine. In this thesis work an external combustor is proposed to fire the producer gas and heat from the combustion gases is extracted by employing the heat exchanger thereby passing the air through it.
Chiaramonti et al. (2004) have investigated on EFGT (100kWel) with fuel sources of solid biomass and natural gas; fired in external and internal combustion systems respectively. The advantage of keeping the share of natural gas has been reported to maintain the maximum cycle temperature same as for the standard natural gas based turbine for getting good electrical efficiency. The authors have reported a decrease in both electrical power output and electrical efficiency, whereas the possibility of heat recovery at the exhaust increased for the externally fired mode as compared to reference gas turbine cycle.

Micro-Combustor Challenges

Like conventional gas turbine combustors, the functional requirements for micro scale combustors include efficient conversion of chemical energy to fluid thermal and kinetic energy with low pressure loss, reliable ignition and wide flammability limits. However, inadequate residence time for complete combustion and higher heat transfer rates from the combustor are the main constraints for micro combustors [Spadaccini et al; 2002]. Designing of recirculation zones for ignition and flame stability are important factors for micro combustors as the large recirculation zones or the designs which do not initiate reactions in all portions of the flow give reduced power density. It is important to complete the combustion process within shorter combustor through-flow time. Damköhler number provides the relation for residence time and the characteristic chemical reaction as under [Spadaccini et al; 2002]; τ reaction Dah τ residence (1.1).
For the complete combustion to take place the residence time must be greater than the chemical reaction time. Whereas the characteristic combustor residence time in terms of bulk flows through the combustor volume is presented by the relation [Spadaccini et al; 2002]; volumetric _ flow _ rate mRT τ residence≈ volume  VP (1.2).
The authors have reported that the residence time can be increased by increasing the size of the chamber, reducing the mass flow rate or increasing the operating pressure. Chemical reaction time is primarily a function of fuel properties and the mixture temperature and pressure. To achieve high power density in micro combustor, high mass flow rate is needed through small chamber volumes. Therefore, in order to get the complete combustion without compromising the power density of the combustor for a given operating pressure thus the density and assuming Damköhler number of unity, the only available option is to reduce the chemical reaction time thus the required residence time [Spadaccini et al; 2002].

Combustion of Low LHV Fuels

Due to increased interest towards integration of power generation systems with gasification processes for utilization of biomass, coal or tars, the need for robust, low emission gas turbine combustion systems for syngas is rapidly growing.
Syngas normally contains hydrogen, carbon monoxide and diluents such as nitrogen with varying chemical composition based on feedstock (e.g. biomass, coal or tar) and the employed gasification technique. The combustion properties of sysgas are generally dependent on H2 and CO contents. Sysgas fuels combustion using low emissions techniques like lean premixed combustion is, however, challenging due to the difficulty in achieving the sufficient mixing of fuel and air before the combustion [Daniele et al; 2008].
Also too low flame temperature and the excessive fuel flow rate impose the problems for combustion of low heating contents fuels. The alternative fuels have relatively more carbon contents than the natural gas which leads to an inefficient combustion development thus giving increased CO production [Tuccillo and Cameretti; 2005].
Daniele et al. (2008) have investigated on Lean Premixed combustion of undiluted syngas; NOx emissions and Lean Blow Out (LBO) limits. Their work focused on studying the properties of four various mixtures of syngas; derived from coal, refinery residues, biomass and co-firing of syngas with natural gas. The authors have reported about zero CO emissions apart from very lean conditions close to the LBO, which is due to low flame temperature with residence time too short for complete oxidation of CO at lower equivalence ratios close to LBO and higher probability of local extinction events. Towards NOx emissions pure syngas mixtures presents more NOx emissions as compared to the CH4 containing mixture. The residence time for pure syngas mixture has been found more than double as compared to the CH4 containing mixture. The results showed that higher the H2 contents in the fuel mixture the leaner the equivalence ratio for LBO flame extinction to occur. This difference is due to the higher laminar flame speeds of hydrogen (H2) compared to CH4. NOx emissions have been found exponentially dependent on the adiabatic flame temperature for all the fuel mixtures.

Gas Turbine Combustor Configurations

Today various combustion schemes exist to achieve an appropriate and wide range ignition with higher combustion efficiency and reduced soot and polluting gases. The new combustion techniques target at low thermal NOx with lower combustion temperature and low CO due to relatively large residence times and minimum cooling air requirements [Tuccillo and Cameretti; 2005]. Available combustion concepts include.

Combustion Separated Zones

In this combustion approach separate combustion zones, each with its own separate supply of well-mixed fuel and air is employed to optimize the combustion performance. For example “staged combustor” is employed (Figure 1.17) wherein primary combustion zone operates at low power and equivalence ratio of around 0.8 to get higher combustion efficiency and minimize the production of CO and unburned HC [Tuccillo and Cameretti; 2005]. The primary combustion zone is responsible for high temperature combustion and acts as a pilot source for the main combustion zone downstream, where premixed fuel-air mixture is supplied.

Catalytic Combustion

Towards low NOx emissions, catalytic combustor is an important scheme, wherein the fuel is pre-vaporized (if liquid) and premixed with air at low equivalence ratio and allowed to pass through a catalytic reactor bed. It results in lean mixture combustion with low NOx production due to lower temperature [Tuccillo and Cameretti; 2005]. One of the main developments related to the combustion is the catalytic combustor, which is effective way to reduce the pollutants emissions (NOx and CO) which also provides efficient use of lower heating contents biomass based fuels [Pilavachi; 2002]. In catalytic combustion which is a flameless process, the fuel oxidation takes place at temperature below 1700 0F thus causing low NOx production [Energy Nexus Group; 2002].

Table of contents :

ABSTRACT
ACKNOWLEDGEMENTS
TABLE OF CONTENTS
LIST OF FIGURES
LIST OF TABLES
NOMENCLATURE
1 BACKGROUND
1.1 MICRO GAS TURBINE
1.2 MICRO GAS TURBINE AND BIOMASS
1.3 INTERNALLY FIRED MICRO GAS TURBINE
1.4 EXTERNALLY FIRED MICRO GAS TURBINE
1.5 COMPRESSOR
1.6 REGENERATOR
1.7 COMBUSTION
1.7.1 Micro-Combustor Challenges
1.7.2 Combustion of Low LHV Fuels
1.7.3 Gas Turbine Combustor Configurations
1.7.3.1 Combustion Separated Zones
1.7.3.2 Catalytic Combustion
1.7.3.3 Engine size
1.7.3.4 Lean Premixed-Prevaporized (LPP) combustion
1.7.3.5 Lean Premixed Combustion
1.7.3.6 Rich Burn- Quick Mix-Lean Burn (RQL) Combustor
1.7.4 Micro Gas Turbine Combustors
1.7.4.1 The Annular Combustor Scaled for MGT applications
1.7.4.2 The Lean-Premixed Combustor
1.7.4.3 Rich Burn- Quick Quench-Lean Burn (RQL) Combustor
1.7.4.4 Comparison between the Combustors for MGT
1.8 EMISSIONS
1.9 GENERATOR AND POWER ELECTRONICS
1.10 MODELING TECHNIQUES
2 OBJECTIVES
3 METHOD OF ATTACK
4 BOUNDARY CONDITIONS
4.1 FUEL FOR THE GAS TURBINE
5 THERMODYNAMIC LAYOUT
5.1 INTERNALLY FIRED MICRO GAS TURBINE CYCLE
5.1.1 Compressor
5.1.1.1 Conditions at inlet of impeller
5.1.2 Regenerator
5.1.3 Combustor
5.1.4 Turbine
5.1.4.1 Conditions at rotor inlet
5.1.4.2 Conditions at rotor outlet
5.1.5 Exhaust Heat Recovery
5.1.6 Fuel Compression
5.2 EXTERNALLY FIRED MICRO GAS TURBINE (EFGT) CYCLE
5.2.1 High Temperature Heat Exchanger
5.2.2 Exhaust Heat Recoveries
5.2.2.1 Gas-to-Water Heat Exchanger
5.2.2.2 Air-to-Water Heat Exchanger
5.2.3 Fuel Compression
5.2.4 Air Compression (External Flow Path)
Numerical Modeling and Analysis of Small Gas Turbine Engine
6 NUMERICAL STUDIES
6.1 CONVERGENCE STUDY-INFLUENCE OF MESH SIZE
6.2 CONVERGENCE STUDY-INFLUENCE OF RMS RESIDUAL TARGET
6.3 COMPRESSOR PERFORMANCE
6.4 STEADY STATE COMPRESSOR SIMULATIONS
6.4.1 Single blade passage (inlet, impeller and diffuser)
6.4.2 360 degree compressor model (inlet, impeller and diffuser)
6.4.3 360 degree compressor model (inlet, rotor and diffuser) connected with inlet straight pipe
6.4.4 Complete compressor-360 degree model (inlet, rotor and diffuser) connected with inlet degree bended pipe
6.5 TRANSIENT COMPRESSOR SIMULATIONS
6.5.1 Single blade passage (inducer, impeller and diffuser)
6.5.2 Complete compressor-360 degree model (inlet, rotor and diffuser)
6.5.3 Complete compressor-360 degree model (inlet, rotor and diffuser) connected with inlet straight pipe
6.5.4 Complete compressor-360 degree model (inlet, rotor and diffuser) with connected inlet bended pipe
6.6 ANALYSIS OF NUMERICAL SIMULATIONS
7 RESULTS AND DISCUSSIONS
7.1 INTERNALLY FIRED GAS TURBINE CYCLE-1.749KWEL FOR ROTATIONAL SPEED OF 120000RPM
7.2 INTERNALLY FIRED GAS TURBINE CYCLE-2.388KWEL FOR ROTATIONAL SPEED OF 130000RPM
7.3 EXTERNALLY FIRED GAS TURBINE CYCLE-0.64KWEL FOR ROTATIONAL SPEED OF 120000RPM
7.4 EXTERNALLY FIRED GAS TURBINE CYCLE-0.986KWEL FOR ROTATIONAL SPEED OF 130000RPM
7.5 PARAMETRIC STUDIES/ANALYSIS OF THE SYSTEM COMPONENTS
7.5.1 Compressor
7.5.2 Exhaust heat recovery heat exchangers
7.5.3 Turbine
8 CONCLUSIONS AND FUTURE WORK
9 REFERENCES
10 APPENDICES
10.1 APPENDIX-A
Derivation of relation for specific fuel consumption (β )
10.2 APPENDIX-B
Explore-Biomass based Polygeneration Flow Chart
10.3 APPENDIX-C
10.3.1 Compressor (COMPR-air compression)
10.3.2 Regenerator
10.3.3 Combustor
10.3.4 Turbine
10.3.5 Exhaust Heat Recovery-Heat Exchanger (HEATREC)
10.3.6 Compressor (BIOCOMP-fuel compression)
10.4 APPENDIX-D
10.4.1 Compressor (MAINCOMP-air compression in the internal flow path)
10.4.2 High Temperature Heat Exchanger
10.4.3 Turbine
10.4.4 Regenerator
10.4.5 Air-to-Water Heat Exchanger (HEATREC1)
10.4.6 Compressor (AIRCOMP-air compression in the external flow path)
10.4.7 Compressor (BIOCOMP-fuel compression)
10.4.8 Combustor
10.4.9 Gas-to-Water Heat Exchanger (HEATREC2)
10.5 APPENDIX-E
10.5.1 Excel calculations-Internally fire gas turbine cycle (2.388kWel)
10.6 APPENDIX-F
10.6.1 Matlab code- Compressor (2.388kWel machine-Internally fired gas turbine cycle)
10.6.2 Matlab code-Turbine (2.388kWel machine-Internally fired gas turbine cycle)
Nawaz Ahmad EGI-2009-001MSc EKV1128
10.7 APPENDIX-G
10.7.1 Excel calculations-Externally fired gas turbine cycle (0.986kWel)
10.8 APPENDIX-H
10.8.1 Matlab code-Compressor (0.986kWel machine-Externally fired gas turbine cycle)
10.8.2 Matlab code-Turbine (0.986kWel machine-Externally fired gas turbine cycle

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