**Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion**

Ene Barbu, Romulus Petcu, Valeriu Vilag, Valentin Silivestru, Tudor Prisecaru, Jeni Popescu, Cleopatra Cuciumita and Sorin Tomescu

Additional information is available at the end of the chapter

http://dx.doi.org/10.5772/54404

#### **1. Introduction**

The gas turbine installations represent one of the most dynamic fields related to the applica‐ bility area and total installed power. The gas turbines have been developed particularly as aviation engines but they find their applicability in many areas, one of which being simulta‐ neously obtaining electric and thermal energy in gas turbine cogeneration plants. The gas turbine cogeneration plants may be classified based on the constructive technology of the gas turbine in [1]: aeroderivative gas turbines plants (up to 10 MW); industrial gas turbines plants, specifically designed for obtaining energy (from 10 up to hundreds MW). An avia‐ tion gas turbine with expired flying resource is still functional due to the fact that the flight time is limited as a consequence of the specific safety normatives requirements. Therefore the aeroderivative gas turbine is defined as a gas turbine, derived from an aviation gas tur‐ bine, dedicated to ground applications. According to the initial destination, these gas tur‐ bines have been designed for maximum efficiency considering the limited fuel quantity available for an aircraft flying large distances. The basic idea in developing the aeroderiva‐ tive gas turbine has been to transfer all the scientific and technologic knowledge ensuring a high degree of energy utilization (design concepts, materials, technologies, etc.) from avia‐ tion to ground [2]. Therefore the obtained gas turbines are lighter, with smaller size, in‐ creased reliability, reduced maintenance costs and high efficiency. The remaining resource for ground applications is proportional with the flight resource, being able to reach up to 30,000 hours considering the lower operating regimes. From the point of view of the actual application, the free power turbine groups are the most recommended [3]. Unlike the aero‐ derivative turbine power units, the industrial power units are built by the original producer

© 2013 Barbu et al.; licensee InTech. This is an open access article distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited. © 2013 Barbu et al.; licensee InTech. This is a paper distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

with the necessary changes for actual industrial application. The development of aeroderi‐ vative and industrial gas turbines has been affected by the progress of the aviation gas tur‐ bines in military and civilian fields. Many aeroderivative gas turbines ensure compression rates of 30:1 [4]. The industrial gas turbines are cumbersome but they are more adaptable for long running and allow longer periods between maintenance controls. The base fuel for gas turbine cogeneration groups is the natural gas (with a possible liquid fuel as alternative) but the diversification of the gas turbines users and the increase in fuels price has pushed the large producers to consider alternative solutions. Nowadays the most utilized fuels in gas‐ eous turbines are the liquid and gas ones (classic and alternative). The high temperature of the exhausted gases, approximately 590 °C on some gas turbines, allows the valorization of the heat resulted in a heat recovery steam generator. Due to the fact that the oxygen concen‐ tration in the exhausted gases is 11-16% (volume), a supplementary fuel burning may be ap‐ plied (afterburning) in order to increase the steam flow rate, compared to the case of the heat recovery steam generator [5]. The afterburning leads to an increase in flexibility and global efficiency of the cogeneration group, allowing the possibility to burn a large variety of fuels, both classic and alternative. Nitrogen oxides usually represent the maine source of emissions from gas turbines. The NOx emissions produced by the afterburning installation of the cogeneration group are different according to the system, but they are usually small and in some cases the installation even contributes to their reduction [6]. The usual methods for NOx emissions reduction, water or steam injection for flame temperature decrease, affect the gas turbine performances, particularly to high operating regimes, leading to CO emis‐ sions increase. It must be noted that the load of the gas turbine also affects the emissions, the gas turbine being designed to operate at high loads. The general theme of the chapter is giv‐ en by the technological aspects that must be considered when aiming to design a gas turbine cogeneration plant flexible from the points of view of the utilized fuel and the qualitative and quantitative results concerning some classic and alternative gas fuels. Based on the spe‐ cific literature in the field and the experience of National Research and Development Insti‐ tute for Gas Turbines COMOTI Bucharest, there are approached theoretic and experimental researches concerning the utilization of natural gas, as classic fuel, and respectively dime‐ thylether (DME), biogas (landfill gas) and syngas, as alternative fuels, in gas turbine cogen‐ eration groups, the interference between flexibility and emissions. It is particularly analysed the issue of reutilization of aviation gas turbines in industrial purposes by their conversion from liquid fuel to gas fuels operation. There is further presented the actual method of con‐ version for an aviation gas turbine in order to be used in cogeneration groups.

loads. Internationally, many companies with top performance in aviation gas turbines are involved in aeroderivative programs in response to market demands for energy producing installations. The best known among these are: Rolls-Royce, Pratt & Whitney, General Elec‐ tric, Motor Sich, Turbomeca, MTU, etc. Rolls-Royce has developed the RB 211-H63 gas tur‐ bine starting from the aviation RB 211 which, through novel constructive and technologic transformations has been pushed to efficiency up to 41.5%. A 38 MW version will be availa‐ ble in 2013 with the possibility of upgrade to 50 MW in future years [7]. Many gas turbine producers aim to reach the full load in ten minutes from the start. A Japanese project of Mit‐ subishi Heavy Industries Ltd. (MHI) aims to manufacture a gas turbine operating at 1700 °C inlet temperature and 62 % efficiency. Pratt & Whitney, starting from the PW 100 turboprop, have developed the ST aeroderivative gas turbine family (ST 18, ST 40). The researches con‐ ducted at National Research and Development Institute for Gas Turbines COMOTI Buchar‐ est have allowed obtaining aeroderivative gas turbines in the 20 – 2,000 kW range, through valorisation of the aviation gas turbines with exhausted flight resource, obsolete or dam‐ aged. Therefore the AI 20 GM (figure 1, right) aeroderivative turboshaft, operating on natu‐ ral gas, is based on the AI 20 turboprop (figure 1, left). The AI 20 GM is used in power groups driving the backup compressors in the natural gas pumping stations on the main line at SC TRANSGAZ SA. The aeroderivative GTC 1000 (figure 2, right), based on TURMO IV C (figure 2, left), operating on natural gas, is used in a power group driving two serial centrifugal compressors for the compression of the associated drill gas, in one SC OMV PET‐ ROM SA oil exploitation, at Ţicleni – Gorj. Researches have also been conducted regarding the valorisation of the landfill gas in a aeroderivative gas turbine applicable to cogeneration groups [2]. A project for a cogeneration application using the GTE 2000 aeroderivative gas turbine has been started in 2000. The result of the project is a cogeneration plant, with two independent lines, producing electric and thermal (hot water) energy, located in the munici‐ pality of Botosani, with SC TERMICA SA as beneficiary (figure 3, left). The experience ac‐ quired from the GTE 2000 cogeneration plant has been used in a new project for a medium power aeroderivative gas turbine cogeneration plant, the application using the ST 18 A aero‐ derivative gas turbine, manufactured by Pratt & Whitney. The ST 18 A aeroderivative gas turbine has been derived from the aviation PW 100 through redesigning a series of compo‐ nents of which are distinguished the combustion chamber, the case and the intake. Further‐ more, the ST 18 A has been designed and manufactured to operate with water injection in the combustion chamber (duplex burners), method that ensures the reduction of NOx emis‐ sions. The application consists in a cogeneration plant, with two independent cogeneration lines, producing electric and thermal (superheated steam used in the oil extraction techno‐ logic process) energy. The beneficiary of the application is SC OMV PETROM SA, Suplacu de Barcău, Bihor County (figure 3, right) [8]. What makes the difference between aviation and aeroderivative gas turbines are operating conditions and reliability. Thus, aviation gas turbines have over the period of their useful life so many ordered starts and stops (associat‐ ed with aircraft flight), short operation between starting and stopping (of hours), short peri‐ ods between revisions (after each stop) and overhauls (after more than 1,000 hours of operation), the lifespan of about 12,000 cumulative hours of operation. Aeroderivative gas turbines can operate up to 8,000 hours continuously without ordered stop, overhauls are

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

http://dx.doi.org/10.5772/54404

81

#### **2. The aeroderivative gas turbine – A solution for gas turbine cogenerative groups flexibility on gas fuels**

The flexibility of the gas turbine cogeneration plants implies reaching an important number of requirements: operating on classic and alternative fuels; capability of fast start; capability to pass easily from full load to partial loads and back; maintaining the efficiency at full load and partial loads; maintaining the emission to a low level even when operating on partial with the necessary changes for actual industrial application. The development of aeroderi‐ vative and industrial gas turbines has been affected by the progress of the aviation gas tur‐ bines in military and civilian fields. Many aeroderivative gas turbines ensure compression rates of 30:1 [4]. The industrial gas turbines are cumbersome but they are more adaptable for long running and allow longer periods between maintenance controls. The base fuel for gas turbine cogeneration groups is the natural gas (with a possible liquid fuel as alternative) but the diversification of the gas turbines users and the increase in fuels price has pushed the large producers to consider alternative solutions. Nowadays the most utilized fuels in gas‐ eous turbines are the liquid and gas ones (classic and alternative). The high temperature of the exhausted gases, approximately 590 °C on some gas turbines, allows the valorization of the heat resulted in a heat recovery steam generator. Due to the fact that the oxygen concen‐ tration in the exhausted gases is 11-16% (volume), a supplementary fuel burning may be ap‐ plied (afterburning) in order to increase the steam flow rate, compared to the case of the heat recovery steam generator [5]. The afterburning leads to an increase in flexibility and global efficiency of the cogeneration group, allowing the possibility to burn a large variety of fuels, both classic and alternative. Nitrogen oxides usually represent the maine source of emissions from gas turbines. The NOx emissions produced by the afterburning installation of the cogeneration group are different according to the system, but they are usually small and in some cases the installation even contributes to their reduction [6]. The usual methods for NOx emissions reduction, water or steam injection for flame temperature decrease, affect the gas turbine performances, particularly to high operating regimes, leading to CO emis‐ sions increase. It must be noted that the load of the gas turbine also affects the emissions, the gas turbine being designed to operate at high loads. The general theme of the chapter is giv‐ en by the technological aspects that must be considered when aiming to design a gas turbine cogeneration plant flexible from the points of view of the utilized fuel and the qualitative and quantitative results concerning some classic and alternative gas fuels. Based on the spe‐ cific literature in the field and the experience of National Research and Development Insti‐ tute for Gas Turbines COMOTI Bucharest, there are approached theoretic and experimental researches concerning the utilization of natural gas, as classic fuel, and respectively dime‐ thylether (DME), biogas (landfill gas) and syngas, as alternative fuels, in gas turbine cogen‐ eration groups, the interference between flexibility and emissions. It is particularly analysed the issue of reutilization of aviation gas turbines in industrial purposes by their conversion from liquid fuel to gas fuels operation. There is further presented the actual method of con‐

80 Progress in Gas Turbine Performance

version for an aviation gas turbine in order to be used in cogeneration groups.

**2. The aeroderivative gas turbine – A solution for gas turbine**

The flexibility of the gas turbine cogeneration plants implies reaching an important number of requirements: operating on classic and alternative fuels; capability of fast start; capability to pass easily from full load to partial loads and back; maintaining the efficiency at full load and partial loads; maintaining the emission to a low level even when operating on partial

**cogenerative groups flexibility on gas fuels**

loads. Internationally, many companies with top performance in aviation gas turbines are involved in aeroderivative programs in response to market demands for energy producing installations. The best known among these are: Rolls-Royce, Pratt & Whitney, General Elec‐ tric, Motor Sich, Turbomeca, MTU, etc. Rolls-Royce has developed the RB 211-H63 gas tur‐ bine starting from the aviation RB 211 which, through novel constructive and technologic transformations has been pushed to efficiency up to 41.5%. A 38 MW version will be availa‐ ble in 2013 with the possibility of upgrade to 50 MW in future years [7]. Many gas turbine producers aim to reach the full load in ten minutes from the start. A Japanese project of Mit‐ subishi Heavy Industries Ltd. (MHI) aims to manufacture a gas turbine operating at 1700 °C inlet temperature and 62 % efficiency. Pratt & Whitney, starting from the PW 100 turboprop, have developed the ST aeroderivative gas turbine family (ST 18, ST 40). The researches con‐ ducted at National Research and Development Institute for Gas Turbines COMOTI Buchar‐ est have allowed obtaining aeroderivative gas turbines in the 20 – 2,000 kW range, through valorisation of the aviation gas turbines with exhausted flight resource, obsolete or dam‐ aged. Therefore the AI 20 GM (figure 1, right) aeroderivative turboshaft, operating on natu‐ ral gas, is based on the AI 20 turboprop (figure 1, left). The AI 20 GM is used in power groups driving the backup compressors in the natural gas pumping stations on the main line at SC TRANSGAZ SA. The aeroderivative GTC 1000 (figure 2, right), based on TURMO IV C (figure 2, left), operating on natural gas, is used in a power group driving two serial centrifugal compressors for the compression of the associated drill gas, in one SC OMV PET‐ ROM SA oil exploitation, at Ţicleni – Gorj. Researches have also been conducted regarding the valorisation of the landfill gas in a aeroderivative gas turbine applicable to cogeneration groups [2]. A project for a cogeneration application using the GTE 2000 aeroderivative gas turbine has been started in 2000. The result of the project is a cogeneration plant, with two independent lines, producing electric and thermal (hot water) energy, located in the munici‐ pality of Botosani, with SC TERMICA SA as beneficiary (figure 3, left). The experience ac‐ quired from the GTE 2000 cogeneration plant has been used in a new project for a medium power aeroderivative gas turbine cogeneration plant, the application using the ST 18 A aero‐ derivative gas turbine, manufactured by Pratt & Whitney. The ST 18 A aeroderivative gas turbine has been derived from the aviation PW 100 through redesigning a series of compo‐ nents of which are distinguished the combustion chamber, the case and the intake. Further‐ more, the ST 18 A has been designed and manufactured to operate with water injection in the combustion chamber (duplex burners), method that ensures the reduction of NOx emis‐ sions. The application consists in a cogeneration plant, with two independent cogeneration lines, producing electric and thermal (superheated steam used in the oil extraction techno‐ logic process) energy. The beneficiary of the application is SC OMV PETROM SA, Suplacu de Barcău, Bihor County (figure 3, right) [8]. What makes the difference between aviation and aeroderivative gas turbines are operating conditions and reliability. Thus, aviation gas turbines have over the period of their useful life so many ordered starts and stops (associat‐ ed with aircraft flight), short operation between starting and stopping (of hours), short peri‐ ods between revisions (after each stop) and overhauls (after more than 1,000 hours of operation), the lifespan of about 12,000 cumulative hours of operation. Aeroderivative gas turbines can operate up to 8,000 hours continuously without ordered stop, overhauls are made at intervals up to 30,000 cumulative operating hours and, for some brands, the cumu‐ lative operating ranges may be even higher.

cal gas fuels can be classified as [9]: high heating value (LHV=45-190 MJ/Nm3

For the gas turbines used in cogeneration groups, for economic reasons, the most used fuels are heavy oil and waste products from various manufacturing or chemistry processes [3]. Using liquid fuels imposes: ensuring combustion without incandescent particles and depos‐ its on the firing tube and the turbine; decreasing the corrosive action of the burned gases caused by the aggressive compounds (sulphur, lead, sodium, vanadium, etc.); solving the pumping and atomization issues (filtration, heating, etc.). A series of fuels must be well pu‐ rified or filtrated for eliminating water, solid particles or some remiss substances. Heavy liq‐ uid fuels must be heated to a convenient temperature to allow their proper pumping and spraying. Coke number and tar number are of particular interest for burning in gas turbines. Coke number (carbon residue) represents the residue left by an oil product (fuel oil, diesel, etc.) when burned in special conditions (closed space, restricted air access, etc.), expressed in mass percent. Tar number indicates the presence of resins, aromatic hydrocarbons, etc. but it must be considered for information only. In order to define the combustion behaviour of a heavy liquid fuel (like oil) it would be indicated to consider as a criterion the product of the coke number and tar number [10]. In terms of reusing aviation gas turbines in industrial purposes, the possibilities of using liquid fuels are decreasing. For each application, the re‐ quirements of the beneficiary must be analysed related to the characteristics of the fuels af‐ fecting the combustion (density, molecular weight, evaporation limit, flammability temperature, volatility, viscosity, surface tension, latent heat of vaporization, calorific value, the tendency for soot, etc.). In terms of using gas fuels, the problem is less challenging due to their thermal stability, high heating value, lack of soot and tar. However, in order to en‐ sure the pressure level required by the gas turbine, afterburning, etc., the elimination of wa‐ ter and different impurities, a control – measurement station must be provided for the gas fuels to be used (natural gas at 2xST 18 plant – figure 4). Some alternative gas fuels (resulted

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

**Figure 4.** Control – measurement station for natural gas at 2xST 18 – Suplacu de Barcau plant (left) with booster

(right) 1 – cogeneration power plant; 2 – control – measurement station; 3 – booster

propane, refinery off-gas), medium heating value (LHV=11.2-30 MJ/Nm3

refinery gas, petrochemical gas, fuels resulted through gasification etc).

*2.1.1. General requirements regarding the utilization of fuels in gas turbines*

landfill gas, coke oven gas), low heating value (LHV<11.2 MJ/Nm3

; butane,

83

; weak natural gas,

; BFG - Blast Furnace Gas,

http://dx.doi.org/10.5772/54404

**Figure 1.** AI 20 turboprop (left) and AI 20 GM gas turbine (right) [2]

**Figure 2.** TURMO IV C turboshaft (left) and GTC 1000 gas turbine (right) [2]

**Figure 3.** GTE 2000 – Botoşani (left) and 2xST 18 – Suplacu de Barcau (right) plants

#### **2.1. Classic and alternative fuels for gas turbine cogeneration groups**

The performances of the gas turbine cogeneration groups (efficiency and emissions) depend in high degree of the type and physical and chemical properties of the used fuels. Depend‐ ing on the lower heating value (LHV), in relation to natural gas (LHV=30-45 MJ/Nm3 ), typi‐ cal gas fuels can be classified as [9]: high heating value (LHV=45-190 MJ/Nm3 ; butane, propane, refinery off-gas), medium heating value (LHV=11.2-30 MJ/Nm3 ; weak natural gas, landfill gas, coke oven gas), low heating value (LHV<11.2 MJ/Nm3 ; BFG - Blast Furnace Gas, refinery gas, petrochemical gas, fuels resulted through gasification etc).

#### *2.1.1. General requirements regarding the utilization of fuels in gas turbines*

made at intervals up to 30,000 cumulative operating hours and, for some brands, the cumu‐

lative operating ranges may be even higher.

82 Progress in Gas Turbine Performance

**Figure 1.** AI 20 turboprop (left) and AI 20 GM gas turbine (right) [2]

**Figure 2.** TURMO IV C turboshaft (left) and GTC 1000 gas turbine (right) [2]

**Figure 3.** GTE 2000 – Botoşani (left) and 2xST 18 – Suplacu de Barcau (right) plants

**2.1. Classic and alternative fuels for gas turbine cogeneration groups**

The performances of the gas turbine cogeneration groups (efficiency and emissions) depend in high degree of the type and physical and chemical properties of the used fuels. Depend‐ ing on the lower heating value (LHV), in relation to natural gas (LHV=30-45 MJ/Nm3

), typi‐

For the gas turbines used in cogeneration groups, for economic reasons, the most used fuels are heavy oil and waste products from various manufacturing or chemistry processes [3]. Using liquid fuels imposes: ensuring combustion without incandescent particles and depos‐ its on the firing tube and the turbine; decreasing the corrosive action of the burned gases caused by the aggressive compounds (sulphur, lead, sodium, vanadium, etc.); solving the pumping and atomization issues (filtration, heating, etc.). A series of fuels must be well pu‐ rified or filtrated for eliminating water, solid particles or some remiss substances. Heavy liq‐ uid fuels must be heated to a convenient temperature to allow their proper pumping and spraying. Coke number and tar number are of particular interest for burning in gas turbines. Coke number (carbon residue) represents the residue left by an oil product (fuel oil, diesel, etc.) when burned in special conditions (closed space, restricted air access, etc.), expressed in mass percent. Tar number indicates the presence of resins, aromatic hydrocarbons, etc. but it must be considered for information only. In order to define the combustion behaviour of a heavy liquid fuel (like oil) it would be indicated to consider as a criterion the product of the coke number and tar number [10]. In terms of reusing aviation gas turbines in industrial purposes, the possibilities of using liquid fuels are decreasing. For each application, the re‐ quirements of the beneficiary must be analysed related to the characteristics of the fuels af‐ fecting the combustion (density, molecular weight, evaporation limit, flammability temperature, volatility, viscosity, surface tension, latent heat of vaporization, calorific value, the tendency for soot, etc.). In terms of using gas fuels, the problem is less challenging due to their thermal stability, high heating value, lack of soot and tar. However, in order to en‐ sure the pressure level required by the gas turbine, afterburning, etc., the elimination of wa‐ ter and different impurities, a control – measurement station must be provided for the gas fuels to be used (natural gas at 2xST 18 plant – figure 4). Some alternative gas fuels (resulted

**Figure 4.** Control – measurement station for natural gas at 2xST 18 – Suplacu de Barcau plant (left) with booster (right) 1 – cogeneration power plant; 2 – control – measurement station; 3 – booster

through gasification and biomass pyrolysis), biogas, residual gases from industrial processes (rich in hydrogen) can play an important role in the operation of the gas turbine cogenera‐ tion groups, but they must reach some requirements regarding the calorific value and the composition [11]. Therefore there is necessary to eliminate the impurities, tar, to limit the sulphur and its compounds to 1 mg/Nm3 , respectively the alkaline metal compounds to 0.1 mg/Nm3 [12].

ent biomass categories and utilization of different gasification technologies, the composition of the resulted gas and the lower heating value (LHV) can vary according to tables 3 and 4 [12, 15]. Tables 1 and 3 show that the lower heating values for biogas and syngas are lower than for the natural gas, requiring, in their application in cogeneration groups, higher mass flow rates with minimum pressure losses. Therefore, the injection nozzles of the gas turbine and the burners of the afterburning installation must be designed for velocities allowing a homogenous mixture between fuel and oxid, as well as low pressure losses. The syngas con‐ tains high quantities of hydrogen which affect the combustion in gas turbine cogeneration groups in terms of flame stability, combustion efficiency, etc. Using hydrogen as fuel and introducing a component with dilution role (steam, nitrogen, etc.) the operation of the gas

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85

**No. Name Natural gas Biogas**

1 CH4 [vol %] 91.0 55-70

2 CnH2n [vol %] 8.09 0

3 CO2 [vol %] 0.61 30-45

4 N2 [vol %] 0.3 0 - 2

5 Lower heating value [MJ/Nm3] 39.2 23.3

6 Density [kg/Nm3] 0.809 1.16

**No. Name Natural gas (Ardeal) Dimethylether**

1 Theoretical combustion temperature [0C] 1,900 2,000

2 Autoignition temperature [0C] 650-750 350

3 Lower heating value [MJ/Nm3] 35.772 59.230

4 Explosion limit [% gas in air] 5 - 15.4 3 - 18.6

5 Density [kg/Nm3] 0.716 2.052

**Table 2.** Physical and chemical characteristics for natural gas (Ardeal) and dimethylether [14]

**Table 1.** Composition, physical and chemical proprieties for natural gas and biogas [13]

turbine is affected [16].

#### *2.1.2. Alternative fuels – Characteristics and consequences regarding their use in gas turbine cogeneration groups*

The biogas produced through anaerobic fermentation is cheap and constitutes a renewable energy source producing, from burning, neutral carbon dioxide (CO2) and offering the pos‐ sibility of treatment and recycling for residues and secondary agricultural products, various biowaste, organic waste water from industry, sewage and sewage sludge. The properties and the composition of biogas are different depending on the raw material used, processing system, temperature, etc. The comparative compositions of natural gas and biogas are given in table 1 [13]. For both fuels the main component (giving the energetic value) is the meth‐ ane (CH4), the significant differences being given by the high content of CO2 and H2S (hy‐ drogen sulphide) in biogas. Technically, the main difference is given by the Wobbe index for natural gas (see chapter 2.2), two times higher than the index for biogas. This leads to a lim‐ ited possibility of replacing the natural gas with biogas because only gases with similar Wobbe index can substitute each other. The improvement of the biogas can be achieved by replacing CO2 with CH4 so as to approach the characteristics of natural gas. Furthermore, the water and hydrogen sulphide must be eliminated to avoid the harmful action of the re‐ sulted sulphuric acid on different components of the cogeneration group (gas turbine, after‐ burning installation, heat recovery steam generator, etc.). Landfill gas resulted from waste deposits represents a cheap energy source, with a composition similar to the biogas resulted from anaerobic fermentation (45-60 % methane, 40-55 % carbon dioxide) [2]. When it comes to using biogas in gas turbine cogeneration groups or introducing it in the natural gas net‐ work, special treatment is required (condensate separation, drying, adsorption of volatile substances, etc.). Dimethylether (DME, CH3-O-CH3) is a clean alternative fuel which can be produced from fossil fuels, namely coal or vegetal biomass gasification. It can be transport‐ ed and stored similar to liquefied petroleum gas (LPG), its physical and chemical character‐ istics, related to natural gas in Ardeal (99.8 % CH4 and 0.2 % CO2), being given in table 2 [14]. The flame produced by burning the dimethylether is very similar to the flame pro‐ duced by the natural gas (figure 5), which makes it suitable to be used as fuel in transporta‐ tion, cogeneration groups, etc.

Through biomass of coal gasification (with oxidant agents such as oxygen, air, steam, etc.) it can obtain synthesis gas (syngas) with main components hydrogen (H2) and carbon monox‐ ide (CO). The syngas can be used to obtain methanol, hydrogen, methane, etc. or can be used as fuel in gas turbine cogeneration groups. Since leaving the gas-producing installation the gas containes ash particles and various compounds of chlorine, fluorine, alkali metals, etc., which must be removed to protect the cogeneration line. Through gasification of differ‐ ent biomass categories and utilization of different gasification technologies, the composition of the resulted gas and the lower heating value (LHV) can vary according to tables 3 and 4 [12, 15]. Tables 1 and 3 show that the lower heating values for biogas and syngas are lower than for the natural gas, requiring, in their application in cogeneration groups, higher mass flow rates with minimum pressure losses. Therefore, the injection nozzles of the gas turbine and the burners of the afterburning installation must be designed for velocities allowing a homogenous mixture between fuel and oxid, as well as low pressure losses. The syngas con‐ tains high quantities of hydrogen which affect the combustion in gas turbine cogeneration groups in terms of flame stability, combustion efficiency, etc. Using hydrogen as fuel and introducing a component with dilution role (steam, nitrogen, etc.) the operation of the gas turbine is affected [16].

through gasification and biomass pyrolysis), biogas, residual gases from industrial processes (rich in hydrogen) can play an important role in the operation of the gas turbine cogenera‐ tion groups, but they must reach some requirements regarding the calorific value and the composition [11]. Therefore there is necessary to eliminate the impurities, tar, to limit the

*2.1.2. Alternative fuels – Characteristics and consequences regarding their use in gas turbine*

The biogas produced through anaerobic fermentation is cheap and constitutes a renewable energy source producing, from burning, neutral carbon dioxide (CO2) and offering the pos‐ sibility of treatment and recycling for residues and secondary agricultural products, various biowaste, organic waste water from industry, sewage and sewage sludge. The properties and the composition of biogas are different depending on the raw material used, processing system, temperature, etc. The comparative compositions of natural gas and biogas are given in table 1 [13]. For both fuels the main component (giving the energetic value) is the meth‐ ane (CH4), the significant differences being given by the high content of CO2 and H2S (hy‐ drogen sulphide) in biogas. Technically, the main difference is given by the Wobbe index for natural gas (see chapter 2.2), two times higher than the index for biogas. This leads to a lim‐ ited possibility of replacing the natural gas with biogas because only gases with similar Wobbe index can substitute each other. The improvement of the biogas can be achieved by replacing CO2 with CH4 so as to approach the characteristics of natural gas. Furthermore, the water and hydrogen sulphide must be eliminated to avoid the harmful action of the re‐ sulted sulphuric acid on different components of the cogeneration group (gas turbine, after‐ burning installation, heat recovery steam generator, etc.). Landfill gas resulted from waste deposits represents a cheap energy source, with a composition similar to the biogas resulted from anaerobic fermentation (45-60 % methane, 40-55 % carbon dioxide) [2]. When it comes to using biogas in gas turbine cogeneration groups or introducing it in the natural gas net‐ work, special treatment is required (condensate separation, drying, adsorption of volatile substances, etc.). Dimethylether (DME, CH3-O-CH3) is a clean alternative fuel which can be produced from fossil fuels, namely coal or vegetal biomass gasification. It can be transport‐ ed and stored similar to liquefied petroleum gas (LPG), its physical and chemical character‐ istics, related to natural gas in Ardeal (99.8 % CH4 and 0.2 % CO2), being given in table 2 [14]. The flame produced by burning the dimethylether is very similar to the flame pro‐ duced by the natural gas (figure 5), which makes it suitable to be used as fuel in transporta‐

Through biomass of coal gasification (with oxidant agents such as oxygen, air, steam, etc.) it can obtain synthesis gas (syngas) with main components hydrogen (H2) and carbon monox‐ ide (CO). The syngas can be used to obtain methanol, hydrogen, methane, etc. or can be used as fuel in gas turbine cogeneration groups. Since leaving the gas-producing installation the gas containes ash particles and various compounds of chlorine, fluorine, alkali metals, etc., which must be removed to protect the cogeneration line. Through gasification of differ‐

, respectively the alkaline metal compounds to 0.1

sulphur and its compounds to 1 mg/Nm3

mg/Nm3

[12].

84 Progress in Gas Turbine Performance

*cogeneration groups*

tion, cogeneration groups, etc.




**Table 2.** Physical and chemical characteristics for natural gas (Ardeal) and dimethylether [14]

have high or low heating value. Density and temperature of the used fuel, as well as the en‐ vironmental temperature, can affect the performances and lifespan of the equipments in the cogeneration group. According to these influence factors, the most important parameter for characterizing the interchangeability is the Wobbe index (named after engineer and mathe‐ matician John Wobbe), defined as ratio between the lower heating value (LHV) and the

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

/ *rel comb air d* = r

<sup>D</sup> æ öæ ö <sup>=</sup> ç ÷ç ÷ <sup>D</sup> è øè ø

**No. Gas name Wobbe index [(MJ/Nm3]**

 Natural gas 48.554 Liquefied petroleum gas 79.993 Methane 47.947 Ethane 62.513 Propane 74.584 Carbon monoxide 12.812 Biogas 27.3 Dimethylether 47.422 Hydrogen 38.3

2 1 1 1 22 *p Wo A*

 r

Therefore, two gas fuels, with different chemical compositions but the same Wobbe index, are interchangeable and the heat delivered to the equipment is equivalent for the same fuel pressure. Table 5 gives the values of Wobbe index for several gas fuels. In order to consider the temperature of the fuel, the Wobbe index can be corrected with the temperature. Accord‐

2 2

where Δp1 and Δp2 represent the overpressure of fuel 1, respectivelly 2, Wo1 and Wo2 – Wobbe indexes of fuel 1, respectively 2, A1 and A2 – injection nozzle area for the two fuels.

0.5 /( ) *Wo LHV drel* <sup>=</sup> (1)

*p Wo A* (3)

(2)

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87

sqare root of density of the fuel, relative to air density (drel):

ing to [17], two fuels are interchageable if they respect:

**Table 5.** Wobbe index for various gases [2, 13, 14]

**Figure 5.** Flame of Bunsen burner, with grid type flame stabilizer, on natural gas (left) and dimethylether (right) [14]


**Table 3.** Chemical composition of syngas and lower heating values resulted from biomass gasification [15]


**Table 4.** Chemical composition of syngas and lower heating values resulted from different methods of gasification [12]

Solving the fuels interchangeability issue for gas turbine cogeneration groups, by develop‐ ing high level combustion technologies for alternative fuels, particularly hydrogen, will have a major impact on system efficiency and environment.

#### **2.2. Fuels interchangeability and validation criteria**

Interchangeability in gas turbine cogeneration groups represents the capability to replace a gas fuel with another without affecting the application or the installation burning the gas fuel. The used gas fuels consist in mixtures of combustible gases (methane and other light hydrocarbons, hydrogen, carbon monoxide) and inert gas (mostly nitrogen, carbon dioxide, water vapor). Depending on the combustible gases ratio (usually methane), the gas fuels can have high or low heating value. Density and temperature of the used fuel, as well as the en‐ vironmental temperature, can affect the performances and lifespan of the equipments in the cogeneration group. According to these influence factors, the most important parameter for characterizing the interchangeability is the Wobbe index (named after engineer and mathe‐ matician John Wobbe), defined as ratio between the lower heating value (LHV) and the sqare root of density of the fuel, relative to air density (drel):

$$\text{Wo} = \text{LHV} \mid \left(d\_{rel}\right)^{0.5} \tag{1}$$

$$d\_{rel} = \rho\_{comb} \lor \rho\_{air} \tag{2}$$

Therefore, two gas fuels, with different chemical compositions but the same Wobbe index, are interchangeable and the heat delivered to the equipment is equivalent for the same fuel pressure. Table 5 gives the values of Wobbe index for several gas fuels. In order to consider the temperature of the fuel, the Wobbe index can be corrected with the temperature. Accord‐ ing to [17], two fuels are interchageable if they respect:

$$\frac{\Delta p\_2}{\Delta p\_1} = \left(\frac{\mathcal{W}\mathcal{O}\_1}{\mathcal{W}\mathcal{O}\_2}\right)^2 \left(\frac{A\_1}{A\_2}\right)^2\tag{3}$$

where Δp1 and Δp2 represent the overpressure of fuel 1, respectivelly 2, Wo1 and Wo2 – Wobbe indexes of fuel 1, respectively 2, A1 and A2 – injection nozzle area for the two fuels.


**Table 5.** Wobbe index for various gases [2, 13, 14]

**Figure 5.** Flame of Bunsen burner, with grid type flame stabilizer, on natural gas (left) and dimethylether (right) [14]

**Table 3.** Chemical composition of syngas and lower heating values resulted from biomass gasification [15]

**CO2 [%]**

16 6 4 56 18 - 4.1 Air gasification 16 6 4 56 15 3 4.1 Air gasification 40 13 15 3 - 29 11.826 Oxygen gasification

**Table 4.** Chemical composition of syngas and lower heating values resulted from different methods of gasification

Solving the fuels interchangeability issue for gas turbine cogeneration groups, by develop‐ ing high level combustion technologies for alternative fuels, particularly hydrogen, will

Interchangeability in gas turbine cogeneration groups represents the capability to replace a gas fuel with another without affecting the application or the installation burning the gas fuel. The used gas fuels consist in mixtures of combustible gases (methane and other light hydrocarbons, hydrogen, carbon monoxide) and inert gas (mostly nitrogen, carbon dioxide, water vapor). Depending on the combustible gases ratio (usually methane), the gas fuels can

**H2O [%]**

**CO [%]**

[12]

**H2 [%]**

86 Progress in Gas Turbine Performance

**CH4 [%]**

**N2 [%]**

have a major impact on system efficiency and environment.

**2.2. Fuels interchangeability and validation criteria**

**Name CO H ] <sup>2</sup> CH4 CnH2n CO2 N2** Dry oak 18.3 16.9 2.8 0.5 16.0 - 5.422 Dry beech 19.4 17.5 2.6 0.6 15.0 49.3 5.526 Dry fir 15.1 19.1 1.6 0.9 15.8 57.1 4.053 Wood coals 31.2 6.3 2.9 - 2.5 57.1 5.702

**Syngas chemical composition [%] Lower heating value**

**LHV [MJ/Nm3]** **[MJ/Nm3**

**Observations**

Therefore, the validation criteria for replacing a fuel with an equivalent one are given by: autoignition temperature, flame temperature (with higher influence on NOx formation), flame velocity, flashback, efficiency, NOx and CO emissions, flue gas dew point, etc. Autoig‐ nition temperature of gas fuel in mixture with air is the temperature on which the instanta‐ neous and explosive autoignition occurs, without the existence of an incandescent source of ignition. The turbulent flame is generally less stable than the laminar flame, the instability in flame front break-up field being emphasized by the increase in tube diameter. Free swirl tur‐ bulent flames are more prone to flame front break-up than the laminar ones due to the higer periferal jet velocity. For turbulence angles greater than 30°, the stability area is achieved on the contour of the burner only for rich mixtures [18]. In areas with poor mixture, due to the decrease in velocity, the backflow can occur without flame attachment on the burner edge. The velocity distribution in the swirl flow determines the stabilisation of the flame as a cen‐ tral suspended one. Components with rapid burning, such as hydrogen, accelerate the flame velocity with a tendency to backflow or extinguishment. The backflow tendency of the flames is proportional with the ignition velocity of the fuel gas, a high velocity leading to a high effect. It is also dependent of the primary air proportion and the components with re‐ duced burning velocity can lead to flame front break-up. In order to consider these factors, an empiric relation has been established for the flame front break-up index at interchangea‐ bility Iret [19]:

$$I\_{ret} = \frac{k\_i f\_i}{k\_b f\_b} \left(\frac{LHV\_i}{LHV\_b}\right)^{0.5} \tag{4}$$

**2.3. Converting the aviation gas turbines from liquid to gas fuels operation**

maintenance costs.

*operation*

The complexity of thermo-gas-dynamic processes defining the gas turbine operation in a cogeneration group require theoretical and experimental research activities on gas tur‐ bines in order to accomplish the conversion from liquid to gas fuels operation. For the gas turbines on market, in exploitation, the exploitation and maintenance technical speci‐ fications are generally known, being provided by the producer. When the object of the research is an existing gas turbine lacking the technical documentation which completely define the contructive solution, the issue must be approached through activities of exper‐ imentation, measurements, CAD 3D modelling, numerical simulation in CFD environ‐ ment, constructive modifications and renewed experimentation in order to validate the constructive solutions, permanently aiming the performances correlated with the maxi‐ mum effectiveness (thrust, power), minimum specific fuel consumption, maximum effi‐ ciency, versatility on fuel conversion, maximum availability, minimum operation and

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

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89

*2.3.1. General criteria – Researches concerning the modifications on a gas turbine for gas fuel*

The basic procedure for an aeroderivative gas turbine is to keep the rotor assembly, com‐ pressor – turbine, which is the "heart" of the gas turbine, form the aviation gas turbine and to redesign the combustion chamber in order to operate on a different fuel than the kerosene. Therefore, for the basic gas turbine in the turboshaft category, at least the com‐ bustion chamber must be designed for gas fuels operation. The shaft of the power tur‐ bine is mechanicaly connected to a driven load, mechanical work consumer, depending on the application involving the aero-derivative gas turbine (electric generator, compres‐ sor, pump, etc.). The command and automatic control system of the aero-derivative gas turbine are designed depending on the application. The bearings can be redesigned, ach‐ ieving a conversion from rolling bearings to slide bearings. For the basic turboprop (des‐ tined for propeller aircrafts), at least the combustion chamber and the reducing gear box and/or the gas generator's turbine must be redesigned, depending if the turboprop does or does not include free turbine. Usually, only the gas generator is used, eliminating the gear box. The issues concerning the automatic control system and the bearings are iden‐ tical to those of the turboshaft. For the basic turbojet (simple flow jet predominantely for military aircrafts) the redesigning of the combustion chamber and the designing of a power turbine gas-dynamicaly connected to the gas generator are necessary [2]. The is‐ sues concerning the automatic control system and the bearings are also identical to those of the turboshaft. Regarding the combustion chamber, is desired to constructively alter‐ ate it as little as possible, maybe only in terms of injection system. Due to the fact that the rest of the parameters characterizing the operating process remain unchanged, those regarding zero velocity and ground conditions of the basic gas turbine, the operation of the combustion chamber can be considered as in terms of gas-dynamic similarity. A first problem that must be studied when replacing the fuel is maintaining the combustion ef‐

where: k – constant concerning the flame front break-up limit; f – factor concerning pri‐ mary air; LHV – lower heating value; b and i – indexes regarding the control fuel, re‐ spectively the replacement fuel. A particular issue is raised by the fuels with reduced heating value. Therefore, the landfill gas contains over 40 % CO2, requiring a suitable fuel feeding in order to achieve combustion. The fuels with reduced heating value have a small range of flammability requiring, at partial loads or transient operating regimes, the utilization of a supplementary fuel (such as propane). The mass flow rates necessary for gas turbine operation on reduced heating value gas fuels are high (neglecting the wa‐ ter or steam injection in the gas turbine) compared with the operation on natural gas, fact that modifies the compressor's operating characteristic [20]. From biomass gasifica‐ tion with air, it is obtained syngas with LHV of 4-6 MJ/Nm3 , and from the gasification with steam or oxygen (see table 4) LHV of 9-13MJ/Nm3 . An alternative for increasing lower heating value is the mixing with natural gas. Therefore, if the landfill gas has a LHV of 17-20 MJ/Nm3 , an equivalent lower heating value can be obtained by mixing 60 % gas with reduced heating value with 40 % CH4, with respect to the composition de‐ scribed in [21].

#### **2.3. Converting the aviation gas turbines from liquid to gas fuels operation**

Therefore, the validation criteria for replacing a fuel with an equivalent one are given by: autoignition temperature, flame temperature (with higher influence on NOx formation), flame velocity, flashback, efficiency, NOx and CO emissions, flue gas dew point, etc. Autoig‐ nition temperature of gas fuel in mixture with air is the temperature on which the instanta‐ neous and explosive autoignition occurs, without the existence of an incandescent source of ignition. The turbulent flame is generally less stable than the laminar flame, the instability in flame front break-up field being emphasized by the increase in tube diameter. Free swirl tur‐ bulent flames are more prone to flame front break-up than the laminar ones due to the higer periferal jet velocity. For turbulence angles greater than 30°, the stability area is achieved on the contour of the burner only for rich mixtures [18]. In areas with poor mixture, due to the decrease in velocity, the backflow can occur without flame attachment on the burner edge. The velocity distribution in the swirl flow determines the stabilisation of the flame as a cen‐ tral suspended one. Components with rapid burning, such as hydrogen, accelerate the flame velocity with a tendency to backflow or extinguishment. The backflow tendency of the flames is proportional with the ignition velocity of the fuel gas, a high velocity leading to a high effect. It is also dependent of the primary air proportion and the components with re‐ duced burning velocity can lead to flame front break-up. In order to consider these factors, an empiric relation has been established for the flame front break-up index at interchangea‐

0.5

(4)

, and from the gasification

. An alternative for increasing

*ii i*

*k f LHV <sup>I</sup> k f LHV* æ ö <sup>=</sup> ç ÷ è ø

*bb b*

where: k – constant concerning the flame front break-up limit; f – factor concerning pri‐ mary air; LHV – lower heating value; b and i – indexes regarding the control fuel, re‐ spectively the replacement fuel. A particular issue is raised by the fuels with reduced heating value. Therefore, the landfill gas contains over 40 % CO2, requiring a suitable fuel feeding in order to achieve combustion. The fuels with reduced heating value have a small range of flammability requiring, at partial loads or transient operating regimes, the utilization of a supplementary fuel (such as propane). The mass flow rates necessary for gas turbine operation on reduced heating value gas fuels are high (neglecting the wa‐ ter or steam injection in the gas turbine) compared with the operation on natural gas, fact that modifies the compressor's operating characteristic [20]. From biomass gasifica‐

lower heating value is the mixing with natural gas. Therefore, if the landfill gas has a

% gas with reduced heating value with 40 % CH4, with respect to the composition de‐

, an equivalent lower heating value can be obtained by mixing 60

*ret*

tion with air, it is obtained syngas with LHV of 4-6 MJ/Nm3

with steam or oxygen (see table 4) LHV of 9-13MJ/Nm3

bility Iret [19]:

88 Progress in Gas Turbine Performance

LHV of 17-20 MJ/Nm3

scribed in [21].

The complexity of thermo-gas-dynamic processes defining the gas turbine operation in a cogeneration group require theoretical and experimental research activities on gas tur‐ bines in order to accomplish the conversion from liquid to gas fuels operation. For the gas turbines on market, in exploitation, the exploitation and maintenance technical speci‐ fications are generally known, being provided by the producer. When the object of the research is an existing gas turbine lacking the technical documentation which completely define the contructive solution, the issue must be approached through activities of exper‐ imentation, measurements, CAD 3D modelling, numerical simulation in CFD environ‐ ment, constructive modifications and renewed experimentation in order to validate the constructive solutions, permanently aiming the performances correlated with the maxi‐ mum effectiveness (thrust, power), minimum specific fuel consumption, maximum effi‐ ciency, versatility on fuel conversion, maximum availability, minimum operation and maintenance costs.

#### *2.3.1. General criteria – Researches concerning the modifications on a gas turbine for gas fuel operation*

The basic procedure for an aeroderivative gas turbine is to keep the rotor assembly, com‐ pressor – turbine, which is the "heart" of the gas turbine, form the aviation gas turbine and to redesign the combustion chamber in order to operate on a different fuel than the kerosene. Therefore, for the basic gas turbine in the turboshaft category, at least the com‐ bustion chamber must be designed for gas fuels operation. The shaft of the power tur‐ bine is mechanicaly connected to a driven load, mechanical work consumer, depending on the application involving the aero-derivative gas turbine (electric generator, compres‐ sor, pump, etc.). The command and automatic control system of the aero-derivative gas turbine are designed depending on the application. The bearings can be redesigned, ach‐ ieving a conversion from rolling bearings to slide bearings. For the basic turboprop (des‐ tined for propeller aircrafts), at least the combustion chamber and the reducing gear box and/or the gas generator's turbine must be redesigned, depending if the turboprop does or does not include free turbine. Usually, only the gas generator is used, eliminating the gear box. The issues concerning the automatic control system and the bearings are iden‐ tical to those of the turboshaft. For the basic turbojet (simple flow jet predominantely for military aircrafts) the redesigning of the combustion chamber and the designing of a power turbine gas-dynamicaly connected to the gas generator are necessary [2]. The is‐ sues concerning the automatic control system and the bearings are also identical to those of the turboshaft. Regarding the combustion chamber, is desired to constructively alter‐ ate it as little as possible, maybe only in terms of injection system. Due to the fact that the rest of the parameters characterizing the operating process remain unchanged, those regarding zero velocity and ground conditions of the basic gas turbine, the operation of the combustion chamber can be considered as in terms of gas-dynamic similarity. A first problem that must be studied when replacing the fuel is maintaining the combustion ef‐ ficiency. A second one concerns the maintaining of constructive-functional temperature distribution (on the walls of the firing tube, in the outlet area of the combustion cham‐ ber and inlet area of the turbine). On the background of the assembly gas-dyanmic char‐ acteristics, the unevenness of the temperatures field on the outlet of the combustion chamber (temperature map) is determined by the geometric characteristics of the dilution area (diameter, length, number and area of holes, etc.) and the characteristics of fuel feeding in the primary area (atomization, jet angles, fuel specifications, etc.). The global temperature map is defined by equation (5) and the radial unevenness for the rotor blad‐ ed area is given by equation (6) [3]:

$$
\theta\_m = \left(T\_{\text{max}}^\* - T\_3^\*\right) / \left(T\_3^\* - T\_2^\*\right) \tag{5}
$$

**Nozzle no. 10 Ø3 holes at a 2α angle**

no. 2 has been selected (with 10 Ø3 holes at 2α=70<sup>0</sup>

1 900 3

4 700 3 5 800 3

2 700 without central hole 3 800 without central hole

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

6 1000 without central hole

Only nozzles with 10 holes of the same diameter have been experimented in order to ensure velocity, penetration and safety in operation. The central hole afects the stability of the com‐ bustion process, increases the flame radiation and the temperature on the walls of the firing tube. The tie criterion for various injection nozzles for natural gas has been the temperature of the blade on hub. It has been noted that nozzle no. 3 leads to low frequency vibrations in a large range of operating regimes, functionally inadmissible. When operating on natural gas, the combustion efficiency increases with the operating regime, the process being unaf‐ fected by the vaporization, but only by the mixing. Following the experimentation, nozzle

tation regimes, the circumferential temperature map values did not pass 18 %. The same manner of minimum configuration modifications has been applied for the rest of the gas tur‐ bines transformed for operating on natural gas (TURMO, MK 701, etc.). Therefore, the firing tube and the combustion chamber case have been kept unmodified for all gas turbines, only redesigning the injection system. Satisfying results have been obtained for the experimenta‐ tion of TURMO: good stability, but in a more limited range compared with other gas tur‐ bines (due to the dependency on the mixing process); temperature map values of 22 % (for the aimed 20 %). For MK 701, the values on the temperature map have reached max. 20 %. A particular problem is considered when the aim is the integration of the gas turbine, modi‐ fied for operating on natural gas, with an existing boiler. The heat recovery steam generator can be derived form an energy steam boiler, a technological steam boiler or a hot (warm) water boiler. The integration analysis for an aeroderivative gas turbine with a hot water boiler shows that the temperature of the burned gases on the stack must be in the usual val‐ ue range and the pressure loss at the passing through the modified boiler (in the cogenera‐ tion group) must be lower than the pressure loss on the initial boiler [23]. The modifications necessary for operating the gas turbine on gas fuels with reduced lower heating value, com‐ pared with the operation on natural gas, are slightly more complex. Therefore, Mitsubishi Heavy Industries Ltd., with extensive experience in manufacturing gas turbines on BFG (Blast Furnace Gas), considerd the heating value of the gas fuel as the key factor in the modi‐ fications scheduled for the gas turbine [24]. Depending on the actual application, more mod‐

ifications can be operated on the gas turbine, compared with the ones in table 7.

**Table 6.** Configuration of the experimental injection nozzles (see figure 6), for AI 20 GM on natural gas [3]

**Diameter of central hole Ø [mm]**

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, without central hole). For all experimen‐

Radial unevenness for the rotor bladed area express the manner of operation on the turbine blades:

$$\theta\_r = \left(T\_{\text{max} \times r}^\* - T\_3^\*\right) / \left(T\_3^\* - T\_2^\*\right) \tag{6}$$

In equations (5) and (6) the significance of symbols is: *Tmax \** - maximum temperature peak; *T*3 *\** - average temperature in the outlet section of the combustion chamber; *T*<sup>2</sup> \* - average tem‐ perature in the outlet section of the compressor; *Tmaxr* \* - maximum average radial tempera‐ ture, circumferential arithmetic mean on the entire section. Normal values for θm, depending on the gas turbine, are in the 20-25 % range, with reported values of 35 %. In direct connec‐ tion with the temperature map on the walls of the firing tube, the equivalent stress of the material must be considered when replacing a fuel with another. In the case of the AI 24 gas turbine modification for operation on gas fuels in the cogeneration group, a difference of 15 % has been reported in the temperature map, considering the flattening of the temperature peaks when passing through the turbine [22]. The adopted solution has been the generaliza‐ tion of the results obtained by National Research and Development Institute for Gas Tur‐ bines COMOTI Bucharest for the AI 20 GM (figure 1) and MK 701 gas generators. In order to achieve the AI 20 GM gas turbine on natural gas (derived from AI 20 on liquid fuel) the adopted constructive solution has been the modification of the injection system, without al‐ tering the firing tube (figure 6). The researches for this transformation have been based on test bench experiments with liquid fuel (in low pressure similitude conditions). In order to reach the functional optimum on natural gas, several injection nozzles have been designed and experimented, according to table 6 [3].


**Table 6.** Configuration of the experimental injection nozzles (see figure 6), for AI 20 GM on natural gas [3]

ficiency. A second one concerns the maintaining of constructive-functional temperature distribution (on the walls of the firing tube, in the outlet area of the combustion cham‐ ber and inlet area of the turbine). On the background of the assembly gas-dyanmic char‐ acteristics, the unevenness of the temperatures field on the outlet of the combustion chamber (temperature map) is determined by the geometric characteristics of the dilution area (diameter, length, number and area of holes, etc.) and the characteristics of fuel feeding in the primary area (atomization, jet angles, fuel specifications, etc.). The global temperature map is defined by equation (5) and the radial unevenness for the rotor blad‐

> ( ) ( ) \* \* \*\* max 3 3 2 / *<sup>m</sup>*

Radial unevenness for the rotor bladed area express the manner of operation on the turbine

( ) ( ) \* \* \*\* max 3 3 2 / *r r*

ture, circumferential arithmetic mean on the entire section. Normal values for θm, depending on the gas turbine, are in the 20-25 % range, with reported values of 35 %. In direct connec‐ tion with the temperature map on the walls of the firing tube, the equivalent stress of the material must be considered when replacing a fuel with another. In the case of the AI 24 gas turbine modification for operation on gas fuels in the cogeneration group, a difference of 15 % has been reported in the temperature map, considering the flattening of the temperature peaks when passing through the turbine [22]. The adopted solution has been the generaliza‐ tion of the results obtained by National Research and Development Institute for Gas Tur‐ bines COMOTI Bucharest for the AI 20 GM (figure 1) and MK 701 gas generators. In order to achieve the AI 20 GM gas turbine on natural gas (derived from AI 20 on liquid fuel) the adopted constructive solution has been the modification of the injection system, without al‐ tering the firing tube (figure 6). The researches for this transformation have been based on test bench experiments with liquid fuel (in low pressure similitude conditions). In order to reach the functional optimum on natural gas, several injection nozzles have been designed


=- - *T T TT* (5)

=- - *T T TT* (6)

*\** - maximum temperature peak;

\* - maximum average radial tempera‐

\*


ed area is given by equation (6) [3]:

90 Progress in Gas Turbine Performance

blades:

*T*3 *\** q

q

In equations (5) and (6) the significance of symbols is: *Tmax*

perature in the outlet section of the compressor; *Tmaxr*

and experimented, according to table 6 [3].

Only nozzles with 10 holes of the same diameter have been experimented in order to ensure velocity, penetration and safety in operation. The central hole afects the stability of the com‐ bustion process, increases the flame radiation and the temperature on the walls of the firing tube. The tie criterion for various injection nozzles for natural gas has been the temperature of the blade on hub. It has been noted that nozzle no. 3 leads to low frequency vibrations in a large range of operating regimes, functionally inadmissible. When operating on natural gas, the combustion efficiency increases with the operating regime, the process being unaf‐ fected by the vaporization, but only by the mixing. Following the experimentation, nozzle no. 2 has been selected (with 10 Ø3 holes at 2α=70<sup>0</sup> , without central hole). For all experimen‐ tation regimes, the circumferential temperature map values did not pass 18 %. The same manner of minimum configuration modifications has been applied for the rest of the gas tur‐ bines transformed for operating on natural gas (TURMO, MK 701, etc.). Therefore, the firing tube and the combustion chamber case have been kept unmodified for all gas turbines, only redesigning the injection system. Satisfying results have been obtained for the experimenta‐ tion of TURMO: good stability, but in a more limited range compared with other gas tur‐ bines (due to the dependency on the mixing process); temperature map values of 22 % (for the aimed 20 %). For MK 701, the values on the temperature map have reached max. 20 %. A particular problem is considered when the aim is the integration of the gas turbine, modi‐ fied for operating on natural gas, with an existing boiler. The heat recovery steam generator can be derived form an energy steam boiler, a technological steam boiler or a hot (warm) water boiler. The integration analysis for an aeroderivative gas turbine with a hot water boiler shows that the temperature of the burned gases on the stack must be in the usual val‐ ue range and the pressure loss at the passing through the modified boiler (in the cogenera‐ tion group) must be lower than the pressure loss on the initial boiler [23]. The modifications necessary for operating the gas turbine on gas fuels with reduced lower heating value, com‐ pared with the operation on natural gas, are slightly more complex. Therefore, Mitsubishi Heavy Industries Ltd., with extensive experience in manufacturing gas turbines on BFG (Blast Furnace Gas), considerd the heating value of the gas fuel as the key factor in the modi‐ fications scheduled for the gas turbine [24]. Depending on the actual application, more mod‐ ifications can be operated on the gas turbine, compared with the ones in table 7.

*2.4.1. Numerical simulation, experimental activity, methods and equipments*

the injection nozzle have been kept unmodified.

**Figure 7.** TV2-117A gas turbine (left) with detailed combustion chamber area (right)

The numerical simulations for the modified injector (figure 8) have taken into consideration the variation of the injection pressure (7.65 - 8.5 bar), of the injection angle β (70 - 85<sup>0</sup>

the position related to the injector's body L (1 - 5 mm). Following the numerical simulations, the optimum configuration has been selected and the eight injectors have been manufac‐ tured along with the injection ramp (figure 10, right), consisting in a circular pipe connected to each injector. The configurations of the injectors for liquid fuel and landfill gas are given in figure 9. The elements eliminated from the initial configuration are the following: the liq‐ uid fuel feeding system; the liquid fuel automatic control system; the command system for the actuators controlling the guide vanes and the first three statoric stages of the compres‐ sor; the deicing system. The experimentation of TA2 bio has been made in the experimental facility of National Research and Development Institute for Gas Turbines COMOTI Buchar‐ est (figure 10) in the following configuration: TA2 bio gas turbine installed on test bench; test cell lubricating system and fuel feeding system for the gas turbine; exhaust system for

) and

Numerical simulation on the TV2-117A gas turbine (figure 7, left) on kerosene has been made in order to obtain a reference model for the gas turbine conversion on gas fuels, partic‐ ularly landfill gas. An eighth of the geometric model, corresponding to one injection nozzle, has been used in simulations considering the combustion chamber simetry. The boundary conditions have been provided by the producer in the technical specifications for three oper‐ ating regimes: take-off, nominal and cruise (with the corresponding temperatures of 1123, 1063 and 1023 K). For simulating the combustion process in the TA2 bio gas turbine, the used fuel has been a synthetic landfill gas with equal volume proportions of methane (CH4) and carbon dioxid (CO2). The real landfill gas contains other chemical species, in small pro‐ portions, which have been considered impurities and have not been taken into account. The numerical simulations have been made on the TA2 with modified injection system, particu‐ larly on the injection nozzles level (figure 8). The modelling of the injection nozzles has been achieved starting from the geometry of the natural gas nozzles. Only the injector's outer body have been kept from the liquid operating gas turbine, eliminating all elements related to the atomization system of the liquid fuel. Related to the initial configuration of the injec‐ tor, only the diameter of the secondary channel and the configuration of the connection with

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**Figure 6.** Modification of the injection system for AI 20 GM gas turbine [3]


**Table 7.** Necessary modifications for a gas turbine, depending on lower heating value of the fuel [24]

#### **2.4. Converting a gas turbine from liquid to gas fuel operation for landfill gas valorisation**

Converting the gas turbine from liquid fuel to gas fuel operation in order to achieve the val‐ orisation of the landfill gas has known two main steps, respecting the principles in chapter 2.3: converting the TV2-117A gas turbine from operating on liquid fuel (kerosene) to gas fuel (natural gas), resulting the TA2 gas turbine; converting the TA2 from operating on natural gas to operating on landfill gas, resulting TA2 bio. In order to achieve these results, numer‐ ous numerical simulations in CFD environment and tests have been used for validating the adopted solutions.

#### *2.4.1. Numerical simulation, experimental activity, methods and equipments*

Numerical simulation on the TV2-117A gas turbine (figure 7, left) on kerosene has been made in order to obtain a reference model for the gas turbine conversion on gas fuels, partic‐ ularly landfill gas. An eighth of the geometric model, corresponding to one injection nozzle, has been used in simulations considering the combustion chamber simetry. The boundary conditions have been provided by the producer in the technical specifications for three oper‐ ating regimes: take-off, nominal and cruise (with the corresponding temperatures of 1123, 1063 and 1023 K). For simulating the combustion process in the TA2 bio gas turbine, the used fuel has been a synthetic landfill gas with equal volume proportions of methane (CH4) and carbon dioxid (CO2). The real landfill gas contains other chemical species, in small pro‐ portions, which have been considered impurities and have not been taken into account. The numerical simulations have been made on the TA2 with modified injection system, particu‐ larly on the injection nozzles level (figure 8). The modelling of the injection nozzles has been achieved starting from the geometry of the natural gas nozzles. Only the injector's outer body have been kept from the liquid operating gas turbine, eliminating all elements related to the atomization system of the liquid fuel. Related to the initial configuration of the injec‐ tor, only the diameter of the secondary channel and the configuration of the connection with the injection nozzle have been kept unmodified.

**Figure 7.** TV2-117A gas turbine (left) with detailed combustion chamber area (right)

**Figure 6.** Modification of the injection system for AI 20 GM gas turbine [3]

**20.95-41.9 (High)**

Standard (Minor mod.)

Standard (Minor mod.)

**Table 7.** Necessary modifications for a gas turbine, depending on lower heating value of the fuel [24]

**35.61 (Natural gas)**

Standard

Turbine Standard Standard Standard Standard

Standard

**2.4. Converting a gas turbine from liquid to gas fuel operation for landfill gas valorisation**

Converting the gas turbine from liquid fuel to gas fuel operation in order to achieve the val‐ orisation of the landfill gas has known two main steps, respecting the principles in chapter 2.3: converting the TV2-117A gas turbine from operating on liquid fuel (kerosene) to gas fuel (natural gas), resulting the TA2 gas turbine; converting the TA2 from operating on natural gas to operating on landfill gas, resulting TA2 bio. In order to achieve these results, numer‐ ous numerical simulations in CFD environment and tests have been used for validating the

Air compressor Standard Standard Standard Modification

**8.38-29.33 (Medium)**

Standard (Minor mod.)

Standard (Minor mod.) **2.51-8.38 (Low)**

Modification

Modification

**Lower heating value [MJ/Nm3]**

92 Progress in Gas Turbine Performance

Combustor

Fuel system

adopted solutions.

The numerical simulations for the modified injector (figure 8) have taken into consideration the variation of the injection pressure (7.65 - 8.5 bar), of the injection angle β (70 - 85<sup>0</sup> ) and the position related to the injector's body L (1 - 5 mm). Following the numerical simulations, the optimum configuration has been selected and the eight injectors have been manufac‐ tured along with the injection ramp (figure 10, right), consisting in a circular pipe connected to each injector. The configurations of the injectors for liquid fuel and landfill gas are given in figure 9. The elements eliminated from the initial configuration are the following: the liq‐ uid fuel feeding system; the liquid fuel automatic control system; the command system for the actuators controlling the guide vanes and the first three statoric stages of the compres‐ sor; the deicing system. The experimentation of TA2 bio has been made in the experimental facility of National Research and Development Institute for Gas Turbines COMOTI Buchar‐ est (figure 10) in the following configuration: TA2 bio gas turbine installed on test bench; test cell lubricating system and fuel feeding system for the gas turbine; exhaust system for the burned gases; monitoring system for acquiring functional parameters. In figure 10 (right) is a pipe ramp ring, yellow color, for gas fuel supply.

**Figure 10.** TA2 gas turbine (left) and TA2 bio gas turbine in the test cell (centre, right)

**Figure 11.** Boroscoping images of the gas injection nozzles – natural gas (left) and the thermocouples (right)

perimental results for the outlet section of the combustion chamber.

The numeric simulations on kerosene [2, 5] have shown that, for the reduced operating re‐ gimes, the flame reaches in high degree the area between two adjacent injectors. Table 8 presents the numerical results for landfill gas combustion in terms of methane mass fraction, illustrating the jet shape, and burned gases temperature in the oultlet section of the combus‐ tion chamber. Analysis of data in table 8, with respect to temperature maps, aiming to ob‐ tain a compact jet in order to protect the walls of the firing tube, have helped selecting the geometric configuration of the injection nozzle: β = 700 and L= 3 mm, used for designing the functional model experimented on TA2 bio, for a mixture of natural gas and carbon dioxide. The experiments have been developed in several series, figure 12 presenting one of the mod‐ els of variation for the components of the synthetic landfill gas mixture. The experimental results have been synthetized in figures 12 and 13. Figure 13 presents the numerical and ex‐

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*2.4.2. Results and discussion*

**Figure 8.** Injection nozzle configuration for landfill gas [2]

**Figure 9.** Injectors for liquid fuel for TV2-117A (left) and landfill gas for TA2 bio (right) [2]

A series of experimentations have been made, the simulated landfill gas being obtained by mixing natural gas with carbon dioxid (provided from tanks). The measurements have been made with the equipments of the test facilities. A ramp of 17 double thermocouples located at the outlet of the combustion chamber, with measuring points at one third and two thirds of the outer firing tube circumference allow the measurement of the Tex and Tin temperatures on two concentric rings (figure 11 right).

**Figure 10.** TA2 gas turbine (left) and TA2 bio gas turbine in the test cell (centre, right)

**Figure 11.** Boroscoping images of the gas injection nozzles – natural gas (left) and the thermocouples (right)

#### *2.4.2. Results and discussion*

the burned gases; monitoring system for acquiring functional parameters. In figure 10

(right) is a pipe ramp ring, yellow color, for gas fuel supply.

94 Progress in Gas Turbine Performance

**Figure 8.** Injection nozzle configuration for landfill gas [2]

on two concentric rings (figure 11 right).

**Figure 9.** Injectors for liquid fuel for TV2-117A (left) and landfill gas for TA2 bio (right) [2]

A series of experimentations have been made, the simulated landfill gas being obtained by mixing natural gas with carbon dioxid (provided from tanks). The measurements have been made with the equipments of the test facilities. A ramp of 17 double thermocouples located at the outlet of the combustion chamber, with measuring points at one third and two thirds of the outer firing tube circumference allow the measurement of the Tex and Tin temperatures

The numeric simulations on kerosene [2, 5] have shown that, for the reduced operating re‐ gimes, the flame reaches in high degree the area between two adjacent injectors. Table 8 presents the numerical results for landfill gas combustion in terms of methane mass fraction, illustrating the jet shape, and burned gases temperature in the oultlet section of the combus‐ tion chamber. Analysis of data in table 8, with respect to temperature maps, aiming to ob‐ tain a compact jet in order to protect the walls of the firing tube, have helped selecting the geometric configuration of the injection nozzle: β = 700 and L= 3 mm, used for designing the functional model experimented on TA2 bio, for a mixture of natural gas and carbon dioxide. The experiments have been developed in several series, figure 12 presenting one of the mod‐ els of variation for the components of the synthetic landfill gas mixture. The experimental results have been synthetized in figures 12 and 13. Figure 13 presents the numerical and ex‐ perimental results for the outlet section of the combustion chamber.


**Figure 12.** Variation of the mass flow rates of carbon dioxide (CO2) and natural gas (CH4) injected in the combustion

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The experimentations have proved a stable operation of the TA2 bio gas turbine on different operating regimes, mainly defined by the mass flow rate and the ratio between the mass flow rate of the natural gas and carbon dioxid. Figure 13, particularly the central area, shows

**Figure 13.** Comparison between the numerical and experimental temperature

chamber [2]

**Table 8.** Numerical results for landfill gas combustion simulation [2]

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion http://dx.doi.org/10.5772/54404 97

**Parameters Results**

1 70 7.65

96 Progress in Gas Turbine Performance

3 70 7.65

5 70 7.65

3 80 7.65

3 85 8.50

**Table 8.** Numerical results for landfill gas combustion simulation [2]

**L [mm] β [0] p [bar] Fuel injection jet Temperature on combustion chamber**

**outlet**

**Figure 12.** Variation of the mass flow rates of carbon dioxide (CO2) and natural gas (CH4) injected in the combustion chamber [2]

**Figure 13.** Comparison between the numerical and experimental temperature

The experimentations have proved a stable operation of the TA2 bio gas turbine on different operating regimes, mainly defined by the mass flow rate and the ratio between the mass flow rate of the natural gas and carbon dioxid. Figure 13, particularly the central area, shows a concordance of the numerical and experimental data, proving that modification of gas tur‐ bines operating on alternative gas fuels can be made based on numerical simulations in CFD environment. The model of a cogeneration plant for electric and thermal energy is illustrat‐ ed in figure 14.

process in primary area; residence time of the products; "freezing" characteristic of the reac‐ tion near the firing tube, etc. For NOx reduction, the temperature in the area of the combus‐ tion reaction and the areas of maximum temperature and the air jets distribution (stage combustion) need to be reduced. The final configuration of the combustion chamber of a gas turbine is a compromise between the NOx level, performance and flexibility. Global reduc‐ tion of the emissions leads to compromises between the emission levels of different compo‐ nents and the assembly characteristics of the combustion chamber (pressure losses, stability and ignition limits, etc.). New concepts must be promoted in order to solve this issue. The usual methods are represented by the water or steam injection in the combustion chamber of the gas turbine, leading to [12]: reduction of NOx up to 25 ppm (for a 15 % O2 volume partic‐ ipation in dry burned gases); increase in turbine power due to the increase in fluid mass flow rate (which can compensate the effect of increased temperature during summer); in‐ crease of flexibility of the installation in exploitation due to the possibility of load variation through steam flow rate variation. However, the high content of vapours in burned gases can lead to: acid corosion occurence (for fuels containing sulphure); increase in thermal stress on the combustion chamber; reduction of the heat recovery level, etc. Numerical simu‐ lations on TV2-117A (figure 15) for water injection in the combustion chamber (through du‐ plex injectors, on natural gas) have shown that the water injection in truncated cone shape, at 45°, characterized by a 12 l/min mass flow rate, leads to minimum NOx concentration in burned gases of 14 ppm. The analysis of combustion products for TA2 (see chapter 2.4), us‐ ing NASA CEA program [27], has shown a decrease of the average maximum temperature. The composition of the landfill gas has been considered in equal volume proportions of methane and carbon dioxide, while the composition of the syngas has been considered that given by [19]. The calculation algorythm has started from the stoichiometric reaction of each fuel and imposing the operating regime (in terms of average maximum temperature of 1063 K for nominal regime) in order to determine the minimum quantity of air necessary for the reaction. Obtaining the equilibrium reactions has determined the calculation of the air ex‐ cess coefficients for each fuel at the given regime, for dry operation. Starting from these ini‐ tial values, water has been introduced in different proportions, up to 23 %. The supplementary quantity of fuel, necessary to reestablish the operating regime of the gas tur‐ bine, in terms of temperature (considering the pressure as unaffected), has been calculated in relation to the quantity of water. The general combustion reactions for each fuel, for the water injection case, for the nominal operating regime, are given by equation (7) for landfill

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b·(CH4+ CO2) + 2·λ· (O2+ 3.76 N2) + a·2·λ· H2O → w H2O + x CO2+yN2+zO2 (7)

b·(0.25·CO +0.09·CO2+0.12·H2+0.52·N2+0.02· CH4) + 0.225·λ· (O2+ 3.76 N2) <sup>+</sup> *<sup>a</sup>* · 0.225 · *<sup>λ</sup>* · *<sup>H</sup>*2*<sup>O</sup>* <sup>→</sup> *<sup>w</sup> <sup>H</sup>*2*<sup>O</sup>* <sup>+</sup> *<sup>x</sup> CO*<sup>2</sup> <sup>+</sup> *<sup>y</sup> <sup>N</sup>*<sup>2</sup> <sup>+</sup> *<sup>z</sup> <sup>O</sup>*<sup>2</sup> (8)

There have been tracked the thermodynamic of the system and the concentrations of the re‐ action products, focusing on carbon monoxid (CO) and nitrogen oxides (NOx). In these con‐ ditions, for the two regimes, the calculations have been made up to a injected water coefficient (noted "a") in oxidant of maximum 2, equivalent to 23 % water in oxidant. The

gas and equation (8) for syngas:

**Figure 14.** Model of an aeroderivative gas turbine cogeneration plant operating on natural gas and landfill gas

### **3. Flexibility of gas turbine cogeneration groups and emissions reduction – Future researches**

Gas turbine cogeneration groups, alone or in combination with fuel cells, can play an impor‐ tan role in the general assembly of energy production and emissions reduction. The NOx reduction must be regarded considering the ensurance of cogeneration group performances in a flexible manner, optimization being possible for a fuel [25]. A higher efficiency implies the optimization of the entire cogeneration plant (gas turbine, afterburning, heat recovery steam generator, etc.). The efficiency must be maintained for partial loads (even below 50 %) or for environmental conditions modification. Starting from 2002, Siemens has taken into consideration the flexibility, eliminating the high pressure barrel of the heat recovery steam generator which requires a long process to reach a certain temperature (in order to avoid the occurence of thermal tensions). Regarding the flexibility, the efficiency and the emissions re‐ duction in gas turbine cogeneration groups, important steps have been made: reduced NOx burners have been introduced in applications; the lifecycle has been analyzed for efficiency increase; the period between maintenence controls has been extended and the conversion from one fuel to another for multi-fuel engines has improved [7]. The factors determining the formation of pollutant agents exhausted along with the burned gases from the gas tur‐ bines are [26]: temperature and air excess coefficient in primary area; homogenization of the process in primary area; residence time of the products; "freezing" characteristic of the reac‐ tion near the firing tube, etc. For NOx reduction, the temperature in the area of the combus‐ tion reaction and the areas of maximum temperature and the air jets distribution (stage combustion) need to be reduced. The final configuration of the combustion chamber of a gas turbine is a compromise between the NOx level, performance and flexibility. Global reduc‐ tion of the emissions leads to compromises between the emission levels of different compo‐ nents and the assembly characteristics of the combustion chamber (pressure losses, stability and ignition limits, etc.). New concepts must be promoted in order to solve this issue. The usual methods are represented by the water or steam injection in the combustion chamber of the gas turbine, leading to [12]: reduction of NOx up to 25 ppm (for a 15 % O2 volume partic‐ ipation in dry burned gases); increase in turbine power due to the increase in fluid mass flow rate (which can compensate the effect of increased temperature during summer); in‐ crease of flexibility of the installation in exploitation due to the possibility of load variation through steam flow rate variation. However, the high content of vapours in burned gases can lead to: acid corosion occurence (for fuels containing sulphure); increase in thermal stress on the combustion chamber; reduction of the heat recovery level, etc. Numerical simu‐ lations on TV2-117A (figure 15) for water injection in the combustion chamber (through du‐ plex injectors, on natural gas) have shown that the water injection in truncated cone shape, at 45°, characterized by a 12 l/min mass flow rate, leads to minimum NOx concentration in burned gases of 14 ppm. The analysis of combustion products for TA2 (see chapter 2.4), us‐ ing NASA CEA program [27], has shown a decrease of the average maximum temperature. The composition of the landfill gas has been considered in equal volume proportions of methane and carbon dioxide, while the composition of the syngas has been considered that given by [19]. The calculation algorythm has started from the stoichiometric reaction of each fuel and imposing the operating regime (in terms of average maximum temperature of 1063 K for nominal regime) in order to determine the minimum quantity of air necessary for the reaction. Obtaining the equilibrium reactions has determined the calculation of the air ex‐ cess coefficients for each fuel at the given regime, for dry operation. Starting from these ini‐ tial values, water has been introduced in different proportions, up to 23 %. The supplementary quantity of fuel, necessary to reestablish the operating regime of the gas tur‐ bine, in terms of temperature (considering the pressure as unaffected), has been calculated in relation to the quantity of water. The general combustion reactions for each fuel, for the water injection case, for the nominal operating regime, are given by equation (7) for landfill gas and equation (8) for syngas:

a concordance of the numerical and experimental data, proving that modification of gas tur‐ bines operating on alternative gas fuels can be made based on numerical simulations in CFD environment. The model of a cogeneration plant for electric and thermal energy is illustrat‐

**Figure 14.** Model of an aeroderivative gas turbine cogeneration plant operating on natural gas and landfill gas

**3. Flexibility of gas turbine cogeneration groups and emissions reduction**

Gas turbine cogeneration groups, alone or in combination with fuel cells, can play an impor‐ tan role in the general assembly of energy production and emissions reduction. The NOx reduction must be regarded considering the ensurance of cogeneration group performances in a flexible manner, optimization being possible for a fuel [25]. A higher efficiency implies the optimization of the entire cogeneration plant (gas turbine, afterburning, heat recovery steam generator, etc.). The efficiency must be maintained for partial loads (even below 50 %) or for environmental conditions modification. Starting from 2002, Siemens has taken into consideration the flexibility, eliminating the high pressure barrel of the heat recovery steam generator which requires a long process to reach a certain temperature (in order to avoid the occurence of thermal tensions). Regarding the flexibility, the efficiency and the emissions re‐ duction in gas turbine cogeneration groups, important steps have been made: reduced NOx burners have been introduced in applications; the lifecycle has been analyzed for efficiency increase; the period between maintenence controls has been extended and the conversion from one fuel to another for multi-fuel engines has improved [7]. The factors determining the formation of pollutant agents exhausted along with the burned gases from the gas tur‐ bines are [26]: temperature and air excess coefficient in primary area; homogenization of the

ed in figure 14.

98 Progress in Gas Turbine Performance

**– Future researches**

$$\text{b}\cdot\text{(CH}\_4\text{+CO}\_2\text{)} + 2\cdot\text{A}\cdot\text{(O}\_2 + 3.76\text{ N}\_2\text{)} + \text{a}\cdot2\cdot\text{A}\cdot\text{H}\_2\text{O} \rightarrow \text{ wH}\_2\text{O} + \text{xCO}\_2\text{+y N}\_2 + \text{z}\cdot\text{O}\_2\tag{7}$$

$$\text{A} \cdot \text{(0.25-CO + 0.09-CO\_2 + 0.12-H\_2 + 0.52-CH4) + 0.225-h \cdot \text{(O}\_2 + 3.76 \, N\_2) + \text{ a} \cdot 0.225 \cdot \lambda \cdot H\_2O \to \text{w } H\_2O + \text{ x } CO\_2 + \text{ y } N\_2 + z \cdot O\_2 \tag{8}$$

There have been tracked the thermodynamic of the system and the concentrations of the re‐ action products, focusing on carbon monoxid (CO) and nitrogen oxides (NOx). In these con‐ ditions, for the two regimes, the calculations have been made up to a injected water coefficient (noted "a") in oxidant of maximum 2, equivalent to 23 % water in oxidant. The maximum proportion of water in oxidant has been limited by the concentration of oxygen resulted from the combustion, minimum 11 %, necessary for the afterburning process. For the nominal operating regime and approximately 15 % water for landfill gas and 12.5 % for syngas, the gas turbine reaches the minimum limit of oxygen.

**Figure 16.** Variation of NOx concentration for the two fuels, at 1063 K, depending on water proportion in oxidant (a)

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**3.1. Afterburning installation as interface between gas turbine and heat recovery steam**

The burned gases flow when exiting the gas turbine is turbulent and unevenly distributed in transversal section. Therefore, backflow can occur in the transversal section of the recovery boiler. The unevenness of the flow and the variation in burned gases composition affects the operation of the afterburning. Therefore, the afterburning is influenced in terms of efficien‐ cy, emissions, flame stability, as well as corrosion of the elements subjected to the action of burned gases. For a good design of the inlet section in the recovery boiler it must be general‐ ly considered the following factors [30]: geometry and direction of the gas turbine exhaust; size of heat exchange surfaces; location of the afterburning burner; mass flow rate and aver‐ age velocity of burned gases exiting the gas turbine; local velocities near the walls and on the first heat exchange surface. The gas turbine exhaust is generally not directly connected with the recovery boiler. After exiting the gas turbine (the case of 2xST 18 Cogeneration Plant at Suplacu de Barcau), the burned gases pass through a silencer, a by-pass assembly, a transom for the connection with the burner and then the afterburning chamber [8]. The gas‐ es flow must be parallel with the axis of the burner's connector (perpendicular to the burner plane). A uniform distribution of the flow in the transversal section ensures a good opera‐ tion of the heat recovery steam generator, particularly regarding the superheater. Therefore, the necessary premises are created for ensuring low emissions on the cogeneration group. If the burned gases or the air are uneven distributed, significant variation of the temperatures downstream the burner can occur. Velocity variation in the transversal section, upstream the burner, must not exceed, on 90 % of the burner's section, ± 15 % of the average velocity measured on the entire transversal section. In reality, the burned gases temperature down‐ stream the burner will never be perfectly uniform. Even for a perfect flow distribution of the turbine gases, upstream the burner, the temperature in the area of each burner module will

**generator**

**Figure 15.** Numerical simulation of water injection in the combustion chamber of TA2 (left) and atomization tests with the duplex injector (right)

Figure 16 shows the variation of NOx for the two fuels (landfill gas and syngas) for the nom‐ inal regime, depending on the injected water proportion. The results of the calculations il‐ lustrate that the use of afterburning along with the operation of the TA2 gas turbine, with water injection, for the good operation of the system, the NOx produced by the gas turbine at 1063 K can only be reduced to 40 ppm for landfill gas and 38.5 ppm for syngas. The oxy‐ gen injected in the air can lead to nitrogen oxides reduction and combustion enhancement resulting [28]: reduction of ignition temperature; increase in flamability limit; increase in adiabatic temperature of the flame; increase in process stability and control; reduction of low heating value fuels consumption, etc. The adiabatic temperature of the flame increases with approximately 50 °C for 1 % increase in oxygen concentration. The volume of burned gases decreases with 12 % for the combustion of natural gas in 3 % oxygen enriched air [29]. Reduction of pollution through combustion in oxygen enriched environment can be used in afterburning installations (for primary or secondary air). Combustion in oxygen enriched environment can increase the efficiency and the flexibility of the cogeneration plant. When adding hydrogen to a gas fuel, there are affected the stability of the flame, the efficiency of the combustion and the emissions. Flame velocity for hydrogen combustion in air, in stoi‐ chiometric conditions, reaches 200 cm/s compared to the combustion of methane in air, for which the velocity is approximately 40 cm/s [29]. Adding hydrogen to the gas fuel of the gas turbine or afterburning installation can lead to CO and NOx emissions reduction.

maximum proportion of water in oxidant has been limited by the concentration of oxygen resulted from the combustion, minimum 11 %, necessary for the afterburning process. For the nominal operating regime and approximately 15 % water for landfill gas and 12.5 % for

**Figure 15.** Numerical simulation of water injection in the combustion chamber of TA2 (left) and atomization tests

Figure 16 shows the variation of NOx for the two fuels (landfill gas and syngas) for the nom‐ inal regime, depending on the injected water proportion. The results of the calculations il‐ lustrate that the use of afterburning along with the operation of the TA2 gas turbine, with water injection, for the good operation of the system, the NOx produced by the gas turbine at 1063 K can only be reduced to 40 ppm for landfill gas and 38.5 ppm for syngas. The oxy‐ gen injected in the air can lead to nitrogen oxides reduction and combustion enhancement resulting [28]: reduction of ignition temperature; increase in flamability limit; increase in adiabatic temperature of the flame; increase in process stability and control; reduction of low heating value fuels consumption, etc. The adiabatic temperature of the flame increases with approximately 50 °C for 1 % increase in oxygen concentration. The volume of burned gases decreases with 12 % for the combustion of natural gas in 3 % oxygen enriched air [29]. Reduction of pollution through combustion in oxygen enriched environment can be used in afterburning installations (for primary or secondary air). Combustion in oxygen enriched environment can increase the efficiency and the flexibility of the cogeneration plant. When adding hydrogen to a gas fuel, there are affected the stability of the flame, the efficiency of the combustion and the emissions. Flame velocity for hydrogen combustion in air, in stoi‐ chiometric conditions, reaches 200 cm/s compared to the combustion of methane in air, for which the velocity is approximately 40 cm/s [29]. Adding hydrogen to the gas fuel of the gas

turbine or afterburning installation can lead to CO and NOx emissions reduction.

syngas, the gas turbine reaches the minimum limit of oxygen.

with the duplex injector (right)

100 Progress in Gas Turbine Performance

**Figure 16.** Variation of NOx concentration for the two fuels, at 1063 K, depending on water proportion in oxidant (a)

#### **3.1. Afterburning installation as interface between gas turbine and heat recovery steam generator**

The burned gases flow when exiting the gas turbine is turbulent and unevenly distributed in transversal section. Therefore, backflow can occur in the transversal section of the recovery boiler. The unevenness of the flow and the variation in burned gases composition affects the operation of the afterburning. Therefore, the afterburning is influenced in terms of efficien‐ cy, emissions, flame stability, as well as corrosion of the elements subjected to the action of burned gases. For a good design of the inlet section in the recovery boiler it must be general‐ ly considered the following factors [30]: geometry and direction of the gas turbine exhaust; size of heat exchange surfaces; location of the afterburning burner; mass flow rate and aver‐ age velocity of burned gases exiting the gas turbine; local velocities near the walls and on the first heat exchange surface. The gas turbine exhaust is generally not directly connected with the recovery boiler. After exiting the gas turbine (the case of 2xST 18 Cogeneration Plant at Suplacu de Barcau), the burned gases pass through a silencer, a by-pass assembly, a transom for the connection with the burner and then the afterburning chamber [8]. The gas‐ es flow must be parallel with the axis of the burner's connector (perpendicular to the burner plane). A uniform distribution of the flow in the transversal section ensures a good opera‐ tion of the heat recovery steam generator, particularly regarding the superheater. Therefore, the necessary premises are created for ensuring low emissions on the cogeneration group. If the burned gases or the air are uneven distributed, significant variation of the temperatures downstream the burner can occur. Velocity variation in the transversal section, upstream the burner, must not exceed, on 90 % of the burner's section, ± 15 % of the average velocity measured on the entire transversal section. In reality, the burned gases temperature down‐ stream the burner will never be perfectly uniform. Even for a perfect flow distribution of the turbine gases, upstream the burner, the temperature in the area of each burner module will be higher than the temperature between the modules. Therefore, the infrared analysis of the channel connecting the gas turbine and the afterburning installation (silencer – by-pass as‐ sembly – connecting transom), at 2xST 18 Plant, has shown unevenness in temperature dis‐ tribution (figure 17). Considering these phenomena, the afterburning installation can compensate, in good conditions, the mass flow decrease in burned gases produced by the gas turbine at partial loads, keeping a corresponding load on the heat recovery steam gener‐ ator. In case of turbine stopping, the heat recovery steam generator with the fresh air after‐ burning is able to keep the steam production at a certain level.

**Author details**

Jeni Popescu1

**References**

mania

Ene Barbu1\*, Romulus Petcu1

, Cleopatra Cuciumita1

2 Politehnica University, Bucharest, Romania

tehnica Bucuresti; 2010

(accessed June 6, 2012).

technology.html (accessed May 31, 2012).

Asachi Iasi; 1994

June 6, 2012).

\*Address all correspondence to: barbu.ene@comoti.ro

mir-Feidt-Badea.pdf (accessed June 5, 2012).

, Valeriu Vilag1

, Valentin Silivestru1

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

and Sorin Tomescu1

1 National Research and Development Institute for Gas Turbines COMOTI, Bucharest, Ro‐

[1] Cenusa V., Benelmir R., Feidt M., Badea A. On gas turbines and combined cycles. http://www.ati2001.unina.it/newpdf/Sessioni/Macchine/Impianti/03-Cenusa-Benel‐

[2] Petcu R. Contributii teoretice si experimentale la utilizarea gazului de depozit ca sur‐ sa de energie. Teza de doctorat - Decizie Senat nr. 100/12.02.2010. Universitatea Poli‐

[3] Carlanescu C. Contributii la problema selectarii si modificarii motoarelor de aviatie pentru utilizarea in scopuri industriale. Teza de doctorat. Universitatea Gheorghe

[4] Energy and Environmental Analysis. Technology Characterization: Gas Turbines. http://www.epa.gov/chp/documents/catalog\_chptech\_gas\_turbines.pdf (accessed

[5] Barbu E., Vilag V., Popescu J., Ionescu S., Ionescu A., Petcu R., Cuciumita C., Cretu M., Vilcu C., Prisecaru T. Afterburning Installation Integration into a Cogeneration Power Plant with Gas Turbine by Numerical and Experimental Analysis. In: Ernesto Benini (ed.), Advances in Gas Turbine Technology. Rijeka: InTech; 2011. p. 139-164. Available from http://www.intechopen.com/articles/show/title/afterburning-installa‐ tion-integration-into-a-cogeneration-power-plant-with-gas-turbine-by-numerical-

[6] Stationary Sources Branch. Stationary Gas Turbines - 40 CFR Part 60. http://

[7] Breeze P. Efficiency versus flexibility: Advances in gas turbine technology. PEI 01/04/2011. http://www.powerengineeringint.com/articles/print/volume-19/issue-3/ gas-steam-turbine-directory/efficiency-versus-flexibility-advances-in-gas-turbine-

www.cdphe.state.co.us/ap/down/statgas.pdf (accessed June 6, 2012).

, Tudor Prisecaru2

,

103

http://dx.doi.org/10.5772/54404

**Figure 17.** Temperature isotherms, in infrared, in the channel connecting the gas turbine and the afterburning instal‐ lation (silencer – by-pass assembly – connecting transom)

#### **3.2. Future research**

Future research is part of the general context of increasing the flexibility of gas turbine co‐ generation groups, the efficiency and reducing the emissions using numerical simulations in CFD environment and experimentations related to: utilization of alternative fuels in gas tur‐ bines and afterburning installations, injection of fluids in the cogeneration line in order to reduce the emissions, integrating the gas turbine with fuel cells, etc.

#### **4. Conclusions**

Along with the flexibility to alternative fuels feeding, the flexibility of a gas turbine cogener‐ ation plant assumes the accomplishment of several requirements: capability of fast start; ca‐ pability to pass easily from full load to partial loads and back; maintaining the efficiency at full load and partial loads; maintaining the emission to a low level even when operating on partial loads. Using aeroderivative gas turbines in the cogeneration field has allowed the sci‐ entific and technologic knowledge transfer utilization (design concepts, materials, technolo‐ gies, etc.), which ensures a high degree of energy, from aviation to ground applications. The experience of National Research and Development Institute for Gas Turbines COMOTI Bu‐ charest, in the field of aeroderivative gas turbines (AI 20 GM, TURMO, MK 701, etc.) has al‐ lowed the conversion of a gas turbine from liquid fuel to landfill gas, for cogeneration, in stable operating conditions.

### **Author details**

be higher than the temperature between the modules. Therefore, the infrared analysis of the channel connecting the gas turbine and the afterburning installation (silencer – by-pass as‐ sembly – connecting transom), at 2xST 18 Plant, has shown unevenness in temperature dis‐ tribution (figure 17). Considering these phenomena, the afterburning installation can compensate, in good conditions, the mass flow decrease in burned gases produced by the gas turbine at partial loads, keeping a corresponding load on the heat recovery steam gener‐ ator. In case of turbine stopping, the heat recovery steam generator with the fresh air after‐

**Figure 17.** Temperature isotherms, in infrared, in the channel connecting the gas turbine and the afterburning instal‐

Future research is part of the general context of increasing the flexibility of gas turbine co‐ generation groups, the efficiency and reducing the emissions using numerical simulations in CFD environment and experimentations related to: utilization of alternative fuels in gas tur‐ bines and afterburning installations, injection of fluids in the cogeneration line in order to

Along with the flexibility to alternative fuels feeding, the flexibility of a gas turbine cogener‐ ation plant assumes the accomplishment of several requirements: capability of fast start; ca‐ pability to pass easily from full load to partial loads and back; maintaining the efficiency at full load and partial loads; maintaining the emission to a low level even when operating on partial loads. Using aeroderivative gas turbines in the cogeneration field has allowed the sci‐ entific and technologic knowledge transfer utilization (design concepts, materials, technolo‐ gies, etc.), which ensures a high degree of energy, from aviation to ground applications. The experience of National Research and Development Institute for Gas Turbines COMOTI Bu‐ charest, in the field of aeroderivative gas turbines (AI 20 GM, TURMO, MK 701, etc.) has al‐ lowed the conversion of a gas turbine from liquid fuel to landfill gas, for cogeneration, in

reduce the emissions, integrating the gas turbine with fuel cells, etc.

burning is able to keep the steam production at a certain level.

lation (silencer – by-pass assembly – connecting transom)

**3.2. Future research**

102 Progress in Gas Turbine Performance

**4. Conclusions**

stable operating conditions.

Ene Barbu1\*, Romulus Petcu1 , Valeriu Vilag1 , Valentin Silivestru1 , Tudor Prisecaru2 , Jeni Popescu1 , Cleopatra Cuciumita1 and Sorin Tomescu1

\*Address all correspondence to: barbu.ene@comoti.ro

1 National Research and Development Institute for Gas Turbines COMOTI, Bucharest, Ro‐ mania

2 Politehnica University, Bucharest, Romania

#### **References**


[8] Barbu E., Ionescu S., Vilag V., Vilcu C., Popescu J., Ionescu A., Petcu R., Prisecaru T., Pop E., Toma T. Integrated analysis of afterburning in a gas turbine cogenerative power plant on gaseous fuel, WSEAS Transaction on Environment and Develop‐ ment, 2010; 6(6) p. 405-416. http://www.wseas.us/e-library/transactions/environment/ 2010/89-806.pdf (accessed June 5, 2012).

[20] Rainer K. Gas turbine fuel considerations. http://www.scribd.com/doc/76918626/Gas-

Gas Turbine Cogeneration Groups Flexibility to Classical and Alternative Gaseous Fuels Combustion

http://dx.doi.org/10.5772/54404

105

[21] Fossum M., Beyer R. Co-combustion: Biomass fuel gas and natural gas. http:// media.godashboard.com/gti/IEA/ieaCofirNOrep.pdf (accessed June 16, 2012).

[22] Ene M., Ion C., Salcianu R. Cercetari de transformare a unei camere de ardere pentru functionare cu gaze naturale. In: TURBO '98, 13-15 iulie 1998, Bucuresti, Romania

[23] Zubcu V., Zubcu D., Stanciu D., Homulescu V. Instalatie de cogenerare cu compo‐ nente recuperate, conditii de compatibilitate. in: TURBO '98, 13-15 iulie 1998, Bucur‐

[24] Komori T., Yamagami N., Hara H. Design for blast furnace gas firing gas turbine. http://www.mnes-usa.com/power/news/sec1/pdf/2004\_nov\_04b.pdf (accessed June

[25] Richards G., McMillian M., Gemmen R., Rogers W., Cully S. Issues for low-emission, fuel-flexible power systems. Progress in Energy and Combustion Science 2001; 27: p.

[26] Carlanescu C., Manea I., Ion C., Sterie St Turbomotoare – Fenomenologia producerii si controlul noxelor. Bucuresti: Editura Academiei Tehnice Militare; 1998.

[27] Zehe, M.J., Gordon, S. & McBride, B.J. (2002), *CAP: A Computer Code for Generating Tabular Thermodynamic Functions from NASA Lewis Coefficients*, NASA Glenn Research Center, NASA TP—2001-210959-REV1, Cleveland, Ohio, U.S.A., http:// www.grc.nasa.gov/WWW/CEAWeb/TP-2001-210959-REV1.pdf (accessed June 26,

[28] Corna N., Bertulessi G. The use of oxigen in biomass and waste-to-energy plants: A flexible and effective tool for emission and process control, Third International Sym‐ posium on Energy from Biomass and Waste, 8-11 November 2010, Venice, Italy

[29] Drnevich R., Meagher J., Papavassiliou V., Raybold T., Stuttaford P., Switzer L., Rosen L. Low NOx emissions in a fuel flexible gas turbine, Issued August 2004, http:// www.netl.doe.gov/technologies/coalpower/turbines/refshelf/reports/41892%20Prax‐ air%20Final%20Report\_Low%20NOx%20Fuel%20Flexible%20Gas%20Turbine.pdf

[30] Daiber J., Fluid dynamics of the HRSG gas side, Power, March 2005, p. 58-63 http://

www.babcockpower.com/pdf/vpi-45.pdf (accessed June 26, 2012).

Turbine-Fuel-Considerations (accessed June 14, 2012).

esti, Romania

20, 2012).

141–169.

2012).

(accessed June 26, 2012).


[20] Rainer K. Gas turbine fuel considerations. http://www.scribd.com/doc/76918626/Gas-Turbine-Fuel-Considerations (accessed June 14, 2012).

[8] Barbu E., Ionescu S., Vilag V., Vilcu C., Popescu J., Ionescu A., Petcu R., Prisecaru T., Pop E., Toma T. Integrated analysis of afterburning in a gas turbine cogenerative power plant on gaseous fuel, WSEAS Transaction on Environment and Develop‐ ment, 2010; 6(6) p. 405-416. http://www.wseas.us/e-library/transactions/environment/

[9] Jones R., Goldmeer J., Monetti B. Addressing gas turbine fuel flexibility. GE Energy. http://www.ge.com/cn/energy/solutions/s1/GE%20Gas%20Turbine%20Fuel%20Flexi‐

[10] Pimsner V., Vasilescu C., Radulescu G. Energetica turbomotoarelor cu ardere interna.

[11] Marco Antonio Rosa do Nascimento and Eraldo Cruz dos Santos. Biofuel and Gas Turbine Engines, Advances in Gas Turbine Technology. In: Ernesto Benini (ed.), Ad‐ vances in Gas Turbine Technology. Rijeka: InTech; 2011. p. 116-138. InTech, Availa‐ ble from: http://www.intechopen.com/books/advances-in-gas-turbine-technology/

[12] Oprea I. Posibilitati de utilizare a gazelor provenite din biomasa in instalatii de tur‐

[13] Jensen J., Jensen A. Biogas and natural gas, fuel mixture for the future. 1st World Con‐ ference and Exihibition on Biomass and Energy, 2000, Sevilla. Available from http://

[14] Panoiu P., Marinescu C., Panoiu N., Oroianu I., Mihaescu L. Posibilitati de utilizare a dimetileterului in scopuri energetice. http://caz.mecen.pub.ro/panoiu.pdf (accessed

[15] Calin L., Jadaneant M., Romanek A. Gazeificarea biomasei lemnoase. Curierul AGIR,

[16] Chiesa P., Lozza G., Mazzocchi L. Using hydrogen as gas turbine fuel, Journal of Gas Turbine and Power, January 2005, vol. 127 73-80 http://www.netl.doe.gov/technolo‐ gies/coalpower/turbines/refshelf/igcc-h2-sygas/Using%20H2%20as%20a%20GT

[17] Ionel I., Ungureanu C., Popescu F. Analiza nivelului de emisii poluante prin schim‐ barea combustibilui la cuptoarele de tratament termic. http://www.tehnicainstalatii‐

[18] Antonescu N., Polizu R., Muntean V., Popescu M. Valorificarea energetica a deseuri‐

[19] Ionel P., Borcea Fl., Barbu E., Marinescu C., Ciobanu C. Mihaescu L. Utilizarea com‐ bustibililor gazosi regenerabili pentru producerea de energie.Bucuresti. Editura Per‐

lor.ro/articole/images/nr12\_76-82.pdf (accessed June 14, 2012).

bine cu gaze. ETCN-2005, 30 iunie-1 iulie 2005, Bucuresti, p. 135-139

2010/89-806.pdf (accessed June 5, 2012).

Bucuresti, Editura Academiei RSR, 1964

biofuel-and-gas-turbine-engines (accessed June 6, 2012).

www.dgc.eu/pdf/Sevilla2000.pdf (accessed June 11, 2012).

bility.pdf (accessed June 6, 2012).

104 Progress in Gas Turbine Performance

June 11, 2012).

fect; 2008

1-2, ianuarie-iunie 2008, p. 87-90

%20Fuel.pdf (accessed June 12, 2012).

lor. Bucuresti. Editura Tehnica; 1988


**Chapter 5**

**Micro Gas Turbine Engine: A Review**

Eraldo Cruz dos Santos, Eli Eber Batista Gomes,

Additional information is available at the end of the chapter

Microturbines are energy generators whose capacity ranges from 15 to 300 kW. Their basic principle comes from open cycle gas turbines, although they present several typi‐ cal features, such as: variable speed, high speed operation, compact size, simple opera‐ bility, easy installation, low maintenance, air bearings, low NOX emissions and usually

Microturbines came into the automotive market between 1950 and 1970. The first microtur‐ bines were based on gas turbine designed to be used in generators of missile launching sta‐ tions, aircraft and bus engines, among other commercial means of transport. The use of this equipment in the energy market increased between 1980 and 1990, when the demand for

Distributed generation systems may prove more attractive in a competitive market to those seeking to increase reliability and gain independence by self-generating. Manufacturers of gas and liquid-fueled microturbines and advanced turbine systems have bench test results showing that they will either meet or beat current emission goals for nitrogen oxides (NOX) and other pollutants (Hamilton, 2001). Air quality regulation agencies need to account for this technological innovation. Emission control technologies and regulations for distributed generation system are not yet precisely defined. However, control technologies that could

> © 2013 do Nascimento et al.; licensee InTech. This is an open access article distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

© 2013 do Nascimento et al.; licensee InTech. This is a paper distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/3.0), which permits unrestricted use,

distribution, and reproduction in any medium, provided the original work is properly cited.

distributed generating technologies increased as well (LISS, 1999).

Marco Antônio Rosa do Nascimento,

Elkin Iván Gutiérrez Velásques and

Rubén Alexis Miranda Carrillo

Lucilene de Oliveira Rodrigues,

Fagner Luis Goulart Dias,

http://dx.doi.org/10.5772/54444

a recuperator (Hamilton, 2001).

**1. Introduction**

**Chapter 5**

## **Micro Gas Turbine Engine: A Review**

Marco Antônio Rosa do Nascimento, Lucilene de Oliveira Rodrigues, Eraldo Cruz dos Santos, Eli Eber Batista Gomes, Fagner Luis Goulart Dias, Elkin Iván Gutiérrez Velásques and Rubén Alexis Miranda Carrillo

Additional information is available at the end of the chapter

http://dx.doi.org/10.5772/54444

**1. Introduction**

Microturbines are energy generators whose capacity ranges from 15 to 300 kW. Their basic principle comes from open cycle gas turbines, although they present several typi‐ cal features, such as: variable speed, high speed operation, compact size, simple opera‐ bility, easy installation, low maintenance, air bearings, low NOX emissions and usually a recuperator (Hamilton, 2001).

Microturbines came into the automotive market between 1950 and 1970. The first microtur‐ bines were based on gas turbine designed to be used in generators of missile launching sta‐ tions, aircraft and bus engines, among other commercial means of transport. The use of this equipment in the energy market increased between 1980 and 1990, when the demand for distributed generating technologies increased as well (LISS, 1999).

Distributed generation systems may prove more attractive in a competitive market to those seeking to increase reliability and gain independence by self-generating. Manufacturers of gas and liquid-fueled microturbines and advanced turbine systems have bench test results showing that they will either meet or beat current emission goals for nitrogen oxides (NOX) and other pollutants (Hamilton, 2001). Air quality regulation agencies need to account for this technological innovation. Emission control technologies and regulations for distributed generation system are not yet precisely defined. However, control technologies that could

reduce emissions from fossil-fueled components of a distributed generation system to levels similar to other traditional fossil-fueled generation equipment are already available.

eration. The program was abandoned around 1990 by the Gas Research Institute, on the grounds of problems with the final cost of the product (Watts, 1999). Since then, the Gas Research Institute began to support new projects in partnership with several companies, such as the Northern Research & Engineering Energy Systems, also supporting the first efforts of Capstone Turbine Corporation (still under the name of its precursor, NoMac

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 109

Some companies in the United States, England and Sweden have recently introduced in the world market commercial units of microturbines. Among these companies are: Al‐ liedSignal, Elliott Energy Systems, Capstone, Ingersoll-Rand Energy Systems & Power Recuperators WorksTM, Turbec, Browman Power and ABB Distributed Generation &

AlliedSignal microturbine has shaft configuration, works with cycle Regenerative open Brayton, its bearings are pneumatic and it has a drive direct current - alternating current (DC/AC) 50/60 Hz (the frequency is reduced from about 1,200 to 50 Hz or 60 Hz) and the compressor and turbine are the radial single stage. The heat transfer efficiency of this stain‐ less steel regenerator is 80-90%. Besides working with diesel oil and natural gas, this micro‐ turbine can burn naphtha, methane, propane, gasoline, and synthetic gas. Its noise level is estimated at 65 dB. A commercial prototype of 75 kW was designed for a 30% efficiency and

Elliott Energy Systems (a subsidiary of Elliott Turbomachinery Company) has a manufac‐ turing and assembly unit in Stuart, Florida with a production capacity of 4,000 units per year. According to Richard Sanders, executive vice president of sales and marketing, El‐ liott has launched two commercial prototypes: a 45 kW microturbine (TA-45model) and another 80 kW (TA-80), and later, a 200 kW microturbine (TA-200). The TA-45 model is rated at 45 kW (Figure 1) at ISO conditions and its main difference from other manufac‐ turers is that it has oil lubricated bearings and a system starting at 24 volts, which, ac‐ cording to Sanders, is unique to microturbines. The TA-80 and TA-200 microturbines models are similar to the TA-45 model. All three can generate electricity in 120/208/240V and can work with different fuels: natural gas, diesel, kerosene, alcohol, gasoline, pro‐

The development works of the components has taken the Capstone in the 90's, build and tested a prototype of a 24 kW microturbine in 1994. And in 1996, Capstone made a project consisting of 37 prototypes for field testing. According to Biasi, 1998, Paul Craig, the President of Capstone Turbine Corporation, expected the 30-kW business model to have a cost of about \$ 500/kW (installed microturbine) and a generation cost of \$ 45-50/ MWh. Figure 2 shows Capstone microturbine, model C65, which is already commercially

Energy Systems) (Gri, 1999).

Volvo Aero Corporation.

**3. State-of-the-art microturbines**

pane, methanol and ethanol (Biasi, 1998).

available.

its installed cost is estimated from \$ 22,500 to 30,000 (Biasi, 1998).

Combustion processes can result in the formation of significant amounts of nitrogen dioxide (NO2) and carbon monoxide (CO). Some manufacturers of microturbines have developed advanced combustion technologies to minimize the formation of these pollutants. They have assured low emissions levels from microturbines fueled with gaseous and liquid fuels.

#### **2. History**

In fact, the technology of microturbines is not new, as researches on this subject can be found since 1970, when the automotive industry viewed the possibility of using microtur‐ bines to replace traditional reciprocating piston engines. However, for a variety of reasons, microturbines did not achieve great success in the automotive segment. The first generation of microturbines was based on turbines originally designed for commercial applications in generating electricity for airplanes, buses, and other means of commercial transportation.

The interest in the market for stationary power spread in the mid-1980 and accelerated in the 1990s, with its reuse in the automobile market in hybrid vehicles and when demand for distributed generation increased (Liss, 1999). Currently, the operation of hybrid vehicles through a microturbine connected to an electric motor, have received special attention from some of the major car manufacturers such as Ford, and research centers (Barker, 1997).

In 1978, Allison began a project aimed at the development and construction of generating groups for military applications, driven by small gas turbines. The main results obtained during testing of these generators revealed: reduction in fuel consumption of 180 l/h to 60 l/h, compared with previous models, frequency stability of about 1%, noise levels below 90 dB and the possibility of using different fuels (diesel, gasoline, etc.). In 1981, a batch with 200 generators was delivered to the U.S. Army, and since then, more than 2,000 units have been provided to integrate the system of electricity generation for Patriot missile launchers (Patriot Systems) (Scott, 2000).

The deregulation of the electricity market in the United States began in 1978 when the Pow‐ er Utility Regulatory Policy Act (PURPA) revolutionized the energy market in the United States, breaking the monopoly of the electricity generation sector, enabling the beginning of the expansion of distributed generation. Since then there has been a significant increase in the proportion of independent generation in the country and, according to a projection made in 1999 by the Gas Research Institute (GRI), this in-house production should reach 35% in 2015 (Gri, 1999).

With a new market structure, i.e., with the possibility of attracting small consumers of energy, microturbines began to be the target of intense research. Already in 1980, under the support of the Gas Research Institute, a program entitled Advanced Energy System (AES) was initiated with a view to develop a small gas turbine, with typical features of aviation turbine, rated at 50 kW and equipped with a heat recovery for a system cogen‐ eration. The program was abandoned around 1990 by the Gas Research Institute, on the grounds of problems with the final cost of the product (Watts, 1999). Since then, the Gas Research Institute began to support new projects in partnership with several companies, such as the Northern Research & Engineering Energy Systems, also supporting the first efforts of Capstone Turbine Corporation (still under the name of its precursor, NoMac Energy Systems) (Gri, 1999).

Some companies in the United States, England and Sweden have recently introduced in the world market commercial units of microturbines. Among these companies are: Al‐ liedSignal, Elliott Energy Systems, Capstone, Ingersoll-Rand Energy Systems & Power Recuperators WorksTM, Turbec, Browman Power and ABB Distributed Generation & Volvo Aero Corporation.

#### **3. State-of-the-art microturbines**

reduce emissions from fossil-fueled components of a distributed generation system to levels

Combustion processes can result in the formation of significant amounts of nitrogen dioxide (NO2) and carbon monoxide (CO). Some manufacturers of microturbines have developed advanced combustion technologies to minimize the formation of these pollutants. They have assured low emissions levels from microturbines fueled with gaseous and liquid fuels.

In fact, the technology of microturbines is not new, as researches on this subject can be found since 1970, when the automotive industry viewed the possibility of using microtur‐ bines to replace traditional reciprocating piston engines. However, for a variety of reasons, microturbines did not achieve great success in the automotive segment. The first generation of microturbines was based on turbines originally designed for commercial applications in generating electricity for airplanes, buses, and other means of commercial transportation.

The interest in the market for stationary power spread in the mid-1980 and accelerated in the 1990s, with its reuse in the automobile market in hybrid vehicles and when demand for distributed generation increased (Liss, 1999). Currently, the operation of hybrid vehicles through a microturbine connected to an electric motor, have received special attention from some of the major car manufacturers such as Ford, and research centers (Barker, 1997).

In 1978, Allison began a project aimed at the development and construction of generating groups for military applications, driven by small gas turbines. The main results obtained during testing of these generators revealed: reduction in fuel consumption of 180 l/h to 60 l/h, compared with previous models, frequency stability of about 1%, noise levels below 90 dB and the possibility of using different fuels (diesel, gasoline, etc.). In 1981, a batch with 200 generators was delivered to the U.S. Army, and since then, more than 2,000 units have been provided to integrate the system of electricity generation for Patriot missile launchers

The deregulation of the electricity market in the United States began in 1978 when the Pow‐ er Utility Regulatory Policy Act (PURPA) revolutionized the energy market in the United States, breaking the monopoly of the electricity generation sector, enabling the beginning of the expansion of distributed generation. Since then there has been a significant increase in the proportion of independent generation in the country and, according to a projection made in 1999 by the Gas Research Institute (GRI), this in-house production should reach

With a new market structure, i.e., with the possibility of attracting small consumers of energy, microturbines began to be the target of intense research. Already in 1980, under the support of the Gas Research Institute, a program entitled Advanced Energy System (AES) was initiated with a view to develop a small gas turbine, with typical features of aviation turbine, rated at 50 kW and equipped with a heat recovery for a system cogen‐

similar to other traditional fossil-fueled generation equipment are already available.

**2. History**

108 Progress in Gas Turbine Performance

(Patriot Systems) (Scott, 2000).

35% in 2015 (Gri, 1999).

AlliedSignal microturbine has shaft configuration, works with cycle Regenerative open Brayton, its bearings are pneumatic and it has a drive direct current - alternating current (DC/AC) 50/60 Hz (the frequency is reduced from about 1,200 to 50 Hz or 60 Hz) and the compressor and turbine are the radial single stage. The heat transfer efficiency of this stain‐ less steel regenerator is 80-90%. Besides working with diesel oil and natural gas, this micro‐ turbine can burn naphtha, methane, propane, gasoline, and synthetic gas. Its noise level is estimated at 65 dB. A commercial prototype of 75 kW was designed for a 30% efficiency and its installed cost is estimated from \$ 22,500 to 30,000 (Biasi, 1998).

Elliott Energy Systems (a subsidiary of Elliott Turbomachinery Company) has a manufac‐ turing and assembly unit in Stuart, Florida with a production capacity of 4,000 units per year. According to Richard Sanders, executive vice president of sales and marketing, El‐ liott has launched two commercial prototypes: a 45 kW microturbine (TA-45model) and another 80 kW (TA-80), and later, a 200 kW microturbine (TA-200). The TA-45 model is rated at 45 kW (Figure 1) at ISO conditions and its main difference from other manufac‐ turers is that it has oil lubricated bearings and a system starting at 24 volts, which, ac‐ cording to Sanders, is unique to microturbines. The TA-80 and TA-200 microturbines models are similar to the TA-45 model. All three can generate electricity in 120/208/240V and can work with different fuels: natural gas, diesel, kerosene, alcohol, gasoline, pro‐ pane, methanol and ethanol (Biasi, 1998).

The development works of the components has taken the Capstone in the 90's, build and tested a prototype of a 24 kW microturbine in 1994. And in 1996, Capstone made a project consisting of 37 prototypes for field testing. According to Biasi, 1998, Paul Craig, the President of Capstone Turbine Corporation, expected the 30-kW business model to have a cost of about \$ 500/kW (installed microturbine) and a generation cost of \$ 45-50/ MWh. Figure 2 shows Capstone microturbine, model C65, which is already commercially available.

Four Honeywell Power Systems microturbines of 70 kW each were, until 2001, being tested in the Jamacha Landfill in New Hampshire - United States. The gas produced in the landfills was about 37% methane, carbon dioxide and air. The gas was cooled to about 14 °C to re‐ move moisture and impurities and then compressed to about 550 kPa for the microturbine power. For the first 3 minutes of turbine operation, the fuel feed was carried out with pro‐ pane. The system operated in parallel and exported electricity to San Diego Gas & Electric. In September 2001, Honeywell decided to stop manufacturing microturbines and uninstal‐ led the four microturbines from the Jamacha Landfill, Figure 3. Until that time, the microtur‐ bines operated for 2000 hours, without showing degradation in performance. Then, the microturbines from Honeywell Microturbines were replaced by turbines with the same ca‐

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 111

In order to develop a new generation of microturbines, in 1998 ABB Distributed Generation established a 50/50 joint venture with Volvo Aero Corporation. This partnership joined the experience of Volvo gas turbine for hybrid electric vehicles with the experience of ABB in the generation and energy conversion at high frequency. This joint venture resulted in the development of a microturbine for cogeneration. Operating on natural gas, the MT100 mi‐ croturbine generates 100 kW of electricity and 152 kW of thermal energy (hot water). As oth‐ er manufacturers of microturbines, the MT100 has a frequency converter that allows the

Table 1. brings is a summary of the main features of microturbine leading manufacturers.

**Figure 3.** Ingersoll-Rand Power WorksTM installed on the Jamacha Landfill - United States.

pacity from Ingersoll-Rand Power WorksTM, as shown in Figure 4 (Pierce, 2002).

generator to operate at variable speed.

**Figure 1.** Elliott Energy Systems Microturbine, TA-45 model.

**Figure 2.** Capstone microturbine, model C65 (Capstone, 2012).

Four Honeywell Power Systems microturbines of 70 kW each were, until 2001, being tested in the Jamacha Landfill in New Hampshire - United States. The gas produced in the landfills was about 37% methane, carbon dioxide and air. The gas was cooled to about 14 °C to re‐ move moisture and impurities and then compressed to about 550 kPa for the microturbine power. For the first 3 minutes of turbine operation, the fuel feed was carried out with pro‐ pane. The system operated in parallel and exported electricity to San Diego Gas & Electric. In September 2001, Honeywell decided to stop manufacturing microturbines and uninstal‐ led the four microturbines from the Jamacha Landfill, Figure 3. Until that time, the microtur‐ bines operated for 2000 hours, without showing degradation in performance. Then, the microturbines from Honeywell Microturbines were replaced by turbines with the same ca‐ pacity from Ingersoll-Rand Power WorksTM, as shown in Figure 4 (Pierce, 2002).

In order to develop a new generation of microturbines, in 1998 ABB Distributed Generation established a 50/50 joint venture with Volvo Aero Corporation. This partnership joined the experience of Volvo gas turbine for hybrid electric vehicles with the experience of ABB in the generation and energy conversion at high frequency. This joint venture resulted in the development of a microturbine for cogeneration. Operating on natural gas, the MT100 mi‐ croturbine generates 100 kW of electricity and 152 kW of thermal energy (hot water). As oth‐ er manufacturers of microturbines, the MT100 has a frequency converter that allows the generator to operate at variable speed.

Table 1. brings is a summary of the main features of microturbine leading manufacturers.

**Figure 1.** Elliott Energy Systems Microturbine, TA-45 model.

110 Progress in Gas Turbine Performance

**Figure 2.** Capstone microturbine, model C65 (Capstone, 2012).

**Figure 3.** Ingersoll-Rand Power WorksTM installed on the Jamacha Landfill - United States.

Microturbines are lower power machines with different applications than larger gas tur‐

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 113

**•** Variable rotation: the turbine variable speed is between 30,000 and 120,000 rpm depend‐

**•** High frequency electric alternator: the generator operates with a converter for AC/DC. In

**•** Reliability: some microturbines have already reached 25,000 hours of operation (approxi‐

**•** Simplicity: the generator is placed in the same turbine shaft being relatively easy to be manufactured and maintained. Moreover, it presents a great potential for inexpensive

**•** High noise levels: to reduce noise levels during operation, microturbines require a specif‐

**•** Air-cooled bearings: the use of air bearings avoid lubricants contamination by combus‐ tion products, prolongs the equipment useful life and reduces maintenance costs;

**•** Retrieve: microturbine manufacturers generally use heat recovery of exhaust gas to heat the air intake of the combustion chamber, thus achieving a thermal efficiency of 30%.

Microturbines have similar set-up of small, medium and large size gas turbines, as descri‐ bed by Nascimento and Santos (2011), i.e., microturbines are formed by an assembly of a compressor, a combustion chamber and a turbine, as shown in the simplified scheme of Fig‐

State-of-the-art microturbines have markedly improved in the last years. Several microtur‐ bines have been developed by manufacturers with different configurations. Their configura‐ tion depends on the application, although they usually consist of a single-shaft microturbine, annular combustor, single stage radial flow compressor and expander, and a recuperator or not. The optimum microturbine rotational speeds at typical power ratings are

Gas microturbines have the same basic operation principle as open cycle gas turbines (Bray‐ ton open cycle). Figure 5 shows the Brayton open cycle. In this cycle the air is compressed by the compressor, going through the combustion chamber where it receives energy from the fuel and thus raising its temperature. Leaving the combustion chamber, the high tempera‐ ture working fluid is directed to the turbine, where it is expanded by supplying power to

between 60 to 90,000 rpm and pressure ratio of 3 or 4 : 1, in a single stage.

the compressor and for the electric generator or other equipment available.

bines, having typically the following characteristics:

addition, the alternator itself is the engine starter;

mately three years) including shutdown and maintenance;

ing on the manufacturer;

and large scale manufacturing;

ic acoustic system;

**4. Configuration**

ure 5.

**•** Compact: easy installation and maintenance;

**Figure 4.** Prototype Ingersoll-Rand Power WorksTM installed on Jamacha Landfill - United States.


**Table 1.** Technical characteristics of leading microturbine manufacturers.

Microturbines are lower power machines with different applications than larger gas tur‐ bines, having typically the following characteristics:


### **4. Configuration**

**Figure 4.** Prototype Ingersoll-Rand Power WorksTM installed on Jamacha Landfill - United States.

**Set Total Efficiency (LHV)**


TA 45 Elliott Energy System 45 A Shaft 30 - 871 -

TA 80 Elliott Energy System 80 A Shaft 30 - 871 68,000

TA 200 Elliott Energy System 200 A Shaft 30 - 871 43,000

C30 Capstone 30 A Shaft 28 871 96,000

C65 Capstone 65 A Shaft 29 871 85,000

C200 HP Capstone 200 A Shaft 33 870 45,000


MT 100 ABB 100 A Shaft 30 4.5 950 70,000

**Pressure Ratio**

kW % °C Rpm

**TET**

**Nominal Speed**

**Power Output**

**Table 1.** Technical characteristics of leading microturbine manufacturers.

**Model Manufacturers**

112 Progress in Gas Turbine Performance

Microturbines have similar set-up of small, medium and large size gas turbines, as descri‐ bed by Nascimento and Santos (2011), i.e., microturbines are formed by an assembly of a compressor, a combustion chamber and a turbine, as shown in the simplified scheme of Fig‐ ure 5.

State-of-the-art microturbines have markedly improved in the last years. Several microtur‐ bines have been developed by manufacturers with different configurations. Their configura‐ tion depends on the application, although they usually consist of a single-shaft microturbine, annular combustor, single stage radial flow compressor and expander, and a recuperator or not. The optimum microturbine rotational speeds at typical power ratings are between 60 to 90,000 rpm and pressure ratio of 3 or 4 : 1, in a single stage.

Gas microturbines have the same basic operation principle as open cycle gas turbines (Bray‐ ton open cycle). Figure 5 shows the Brayton open cycle. In this cycle the air is compressed by the compressor, going through the combustion chamber where it receives energy from the fuel and thus raising its temperature. Leaving the combustion chamber, the high tempera‐ ture working fluid is directed to the turbine, where it is expanded by supplying power to the compressor and for the electric generator or other equipment available.

Microturbines are a technology based cycle with or without recuperation. To produce an ac‐ ceptable efficiency, the heat in the turbine exhaust system must be partially recovered and used to preheat the turbine air supply before it enters the combustor, using an air-to-air heat exchanger called recuperator or regenerator. This allows the net cycle efficiency to be in‐ creased to as much as 30% while the average net efficiency of unrecovered microturbines is 17 % (Rodgers et. al., 2001a).

The most effective fuel to minimize emissions is clearly natural gas. Natural gas is also the fuel choice for small businesses. Usually the natural gas requires compression to the ambi‐ ent pressure at the compressor inlet of the microturbine. The compressor outlet pressure is

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 115

Capstone microturbine control and power electronic systems allow for different operation modes, such as: grid connect, stand-alone, dual mode and multiple units for potentially en‐ hanced reliability, operating with gas, liquid fuels and biogas. In grid connect, the system follows the voltage and the frequency from the grid. Grid connect applications include base load, peak shaving and load following. One of the key aspects of a grid connect system is that the synchronization and the protective relay functions required to reliably and safely interconnect with the grid can be integrated directly into the microturbine control and pow‐ er electronic systems. This capability eliminates the need for very expensive and cumber‐ some external equipment needed in conventional generation technologies (Rodgers et. al., 2001a). In the stand-alone mode, the system behaves as an independent voltage source and supplies the current demanded by the load. Capstone microturbine when equipped with the stand-alone option includes a large battery used for unassisted black start of the turbine en‐

nominally three to four atmospheres.

gine and for transient electrical load management.

**Figure 6.** Parts of a Capstone microturbine.

**Figure 5.** Gas turbine system scheme of a simple open cycle.

As well as in gas turbines, the maximum net power provided by a microturbine is limited by the temperature the material of the turbine can support, associated with the cooling tech‐ nology and service life required. The two main factors affecting the performance of micro‐ turbines are: components efficiency and gases temperature at the turbine inlet.

Furthermore, microturbines usually employ permanent magnet variable-speed alternators generating very high frequency alternating current which must be first rectified and then converted to AC to match the required supply frequency.

Capstone Microturbines, shown in Figure 6, uses a lean premix combustion system to ach‐ ieve low emissions levels at a full power range. Lean premix operation requires operating at high air-fuel ratio within the primary combustion zone. The large amount of air is thorough‐ ly mixed with fuel before combustion. This premixing of air and fuel enables clean combus‐ tion to occur at a relatively low temperature. Injectors control the air-fuel ratio and the airfuel mixture in the primary zone to ensure that the optimal temperature is achieved for the NOX minimization. The higher air-fuel ratio results in a lower flame temperature, which leads to lower NOX levels. In order to achieve low levels of CO and Hydrocarbons simulta‐ neously with low NOX levels, the air-fuel mixture is retained in the combustion chamber for a relatively long period. This process allows for a more complete combustion of CO and Hy‐ drocarbons (Capstone, 2000).

In addition, the exhaust of microturbines can be used in direct heating or as an air pre-heat‐ er for downstream burners, once it has a high concentration of oxygen. Clean burning com‐ bustion is the key to both low emissions and highly durable recuperator designs.

The most effective fuel to minimize emissions is clearly natural gas. Natural gas is also the fuel choice for small businesses. Usually the natural gas requires compression to the ambi‐ ent pressure at the compressor inlet of the microturbine. The compressor outlet pressure is nominally three to four atmospheres.

Capstone microturbine control and power electronic systems allow for different operation modes, such as: grid connect, stand-alone, dual mode and multiple units for potentially en‐ hanced reliability, operating with gas, liquid fuels and biogas. In grid connect, the system follows the voltage and the frequency from the grid. Grid connect applications include base load, peak shaving and load following. One of the key aspects of a grid connect system is that the synchronization and the protective relay functions required to reliably and safely interconnect with the grid can be integrated directly into the microturbine control and pow‐ er electronic systems. This capability eliminates the need for very expensive and cumber‐ some external equipment needed in conventional generation technologies (Rodgers et. al., 2001a). In the stand-alone mode, the system behaves as an independent voltage source and supplies the current demanded by the load. Capstone microturbine when equipped with the stand-alone option includes a large battery used for unassisted black start of the turbine en‐ gine and for transient electrical load management.

Microturbines are a technology based cycle with or without recuperation. To produce an ac‐ ceptable efficiency, the heat in the turbine exhaust system must be partially recovered and used to preheat the turbine air supply before it enters the combustor, using an air-to-air heat exchanger called recuperator or regenerator. This allows the net cycle efficiency to be in‐ creased to as much as 30% while the average net efficiency of unrecovered microturbines is

As well as in gas turbines, the maximum net power provided by a microturbine is limited by the temperature the material of the turbine can support, associated with the cooling tech‐ nology and service life required. The two main factors affecting the performance of micro‐

Furthermore, microturbines usually employ permanent magnet variable-speed alternators generating very high frequency alternating current which must be first rectified and then

Capstone Microturbines, shown in Figure 6, uses a lean premix combustion system to ach‐ ieve low emissions levels at a full power range. Lean premix operation requires operating at high air-fuel ratio within the primary combustion zone. The large amount of air is thorough‐ ly mixed with fuel before combustion. This premixing of air and fuel enables clean combus‐ tion to occur at a relatively low temperature. Injectors control the air-fuel ratio and the airfuel mixture in the primary zone to ensure that the optimal temperature is achieved for the NOX minimization. The higher air-fuel ratio results in a lower flame temperature, which leads to lower NOX levels. In order to achieve low levels of CO and Hydrocarbons simulta‐ neously with low NOX levels, the air-fuel mixture is retained in the combustion chamber for a relatively long period. This process allows for a more complete combustion of CO and Hy‐

In addition, the exhaust of microturbines can be used in direct heating or as an air pre-heat‐ er for downstream burners, once it has a high concentration of oxygen. Clean burning com‐

bustion is the key to both low emissions and highly durable recuperator designs.

turbines are: components efficiency and gases temperature at the turbine inlet.

17 % (Rodgers et. al., 2001a).

114 Progress in Gas Turbine Performance

**Figure 5.** Gas turbine system scheme of a simple open cycle.

converted to AC to match the required supply frequency.

drocarbons (Capstone, 2000).

In both operational mode, that is, the grid connect and the stand-alone, the microturbine can also be designed to automatically switch between these two modes. This type of functionali‐ ty is extremely useful in a wide variety of applications, and is commonly referred to as dual mode operation. Besides, the microturbines can be configured to operate in parallel with other distributed generation systems in order to obtain a larger power generation system. This capability can be built directly into the system and does not require the use of any ex‐ ternal synchronizing equipment.

Some microturbines can operate with different fuels. The flexibility and the adaptability ena‐ bled by digital control software allow this to happen with no significant changes to the hardware. Power generation systems create large amounts of heat in the process of convert‐ ing fuel into electricity. For the average utility-size power plant, more than two-thirds of the energy content of the input fuel is converted into heat. Conventional power plants discard this waste heat, however, distributed generation technologies, due to their load-appropriate size and sitting, enable this heat to be recovered. Cogeneration systems can produce heat and electricity at or near the load side. Cogeneration plants usually have up to 85% of effi‐ ciency and operation cost lower than other applications. Small cogeneration systems usually use reciprocating engines although microturbines have showed to be a good option for this application. The hot exhaust gas from microturbines is available for cogeneration applica‐ tions. Recovered heat can be used for hot water heating or low-pressure steam applications.

**Figure 7.** Capstone microturbine in the laboratory at UNIFEI.

**Figure 8.** Schematic representation of the test rig and the data acquisition system.

tics at ISO condition.

This microturbine is mainly used for primer power generation or emergency and can work with a variety of liquid fuels. This microturbine uses a recovery cycle to improve its efficiency during operation, due to a relatively low pressure, what facilitates the use of a single shaft radial compression and expansion [Cohen, *et. al.*, (1996), Capstone, (2001), Roger, *et. al.*, (2001b), Bolszo (2009)]. Table 2 shows the engine design characteris‐

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 117

#### **5. Experimental set-up for microturbine**

To perform tests in microturbines, a test bench was built in the Laboratory of Gas Turbines and Gasification of the Institute of Mechanical Engineering, Federal University of Itajubá - IEM/UNIFEI. This bench was composed of a 30 kW regenerative cycle diesel single shaft gas microturbine engine with annular combustion chamber and radial turbomachineries, as shown in Figure 7, and was configured to operate with liquid fuel.

The microturbine engine was tested while in operation with automotive ethanol and pure die‐ sel, respectively. Thermal and electrical parameters, such as mass flows, temperature, compo‐ sition of exhaust gases and generated power were constantly measured during the tests.

Figure 8 shows the scheme of the microturbine with the measuring points. The microturbine engine was tested during operation with ethanol and diesel at steady state condition and at partial, medium and full loads.

As can be seen in Figure 8, all parameters assessed, during laboratory tests, were acquired and post-processed in a supervisory system developed in the laboratory UNIFEI.

In order to establish whether the fuels were able to feed the engine without presenting any problems regarding the fuel injection system, the kinematic viscosity of each fuel was meas‐ ured. The composition of the emission gases and the thermal variables were also measured at medium and full loads for each fuel, and their results are presented below. All tests were performed in the grid connection mode.

**Figure 7.** Capstone microturbine in the laboratory at UNIFEI.

In both operational mode, that is, the grid connect and the stand-alone, the microturbine can also be designed to automatically switch between these two modes. This type of functionali‐ ty is extremely useful in a wide variety of applications, and is commonly referred to as dual mode operation. Besides, the microturbines can be configured to operate in parallel with other distributed generation systems in order to obtain a larger power generation system. This capability can be built directly into the system and does not require the use of any ex‐

Some microturbines can operate with different fuels. The flexibility and the adaptability ena‐ bled by digital control software allow this to happen with no significant changes to the hardware. Power generation systems create large amounts of heat in the process of convert‐ ing fuel into electricity. For the average utility-size power plant, more than two-thirds of the energy content of the input fuel is converted into heat. Conventional power plants discard this waste heat, however, distributed generation technologies, due to their load-appropriate size and sitting, enable this heat to be recovered. Cogeneration systems can produce heat and electricity at or near the load side. Cogeneration plants usually have up to 85% of effi‐ ciency and operation cost lower than other applications. Small cogeneration systems usually use reciprocating engines although microturbines have showed to be a good option for this application. The hot exhaust gas from microturbines is available for cogeneration applica‐ tions. Recovered heat can be used for hot water heating or low-pressure steam applications.

To perform tests in microturbines, a test bench was built in the Laboratory of Gas Turbines and Gasification of the Institute of Mechanical Engineering, Federal University of Itajubá - IEM/UNIFEI. This bench was composed of a 30 kW regenerative cycle diesel single shaft gas microturbine engine with annular combustion chamber and radial turbomachineries, as

The microturbine engine was tested while in operation with automotive ethanol and pure die‐ sel, respectively. Thermal and electrical parameters, such as mass flows, temperature, compo‐ sition of exhaust gases and generated power were constantly measured during the tests.

Figure 8 shows the scheme of the microturbine with the measuring points. The microturbine engine was tested during operation with ethanol and diesel at steady state condition and at

As can be seen in Figure 8, all parameters assessed, during laboratory tests, were acquired

In order to establish whether the fuels were able to feed the engine without presenting any problems regarding the fuel injection system, the kinematic viscosity of each fuel was meas‐ ured. The composition of the emission gases and the thermal variables were also measured at medium and full loads for each fuel, and their results are presented below. All tests were

and post-processed in a supervisory system developed in the laboratory UNIFEI.

ternal synchronizing equipment.

116 Progress in Gas Turbine Performance

**5. Experimental set-up for microturbine**

partial, medium and full loads.

performed in the grid connection mode.

shown in Figure 7, and was configured to operate with liquid fuel.

**Figure 8.** Schematic representation of the test rig and the data acquisition system.

This microturbine is mainly used for primer power generation or emergency and can work with a variety of liquid fuels. This microturbine uses a recovery cycle to improve its efficiency during operation, due to a relatively low pressure, what facilitates the use of a single shaft radial compression and expansion [Cohen, *et. al.*, (1996), Capstone, (2001), Roger, *et. al.*, (2001b), Bolszo (2009)]. Table 2 shows the engine design characteris‐ tics at ISO condition.


**5.1. Adjustments to the microturbine**

LHV must include:

**•** Natural gas line;

**•** Lines pilot;

**•** Compressor;

**•** Fuel line for a LHV;

**•** Steam line to reduce NOX;

**•** Line of nitrogen to purge;

**•** Combustion Chamber.

**5.2. Tests on gas turbine using liquid fuel**

done automatically.

**•** Line blending of fuel for LHV;

was needed, in relation to the fuel low heating value (LHV).

Due to impurities in the gas or fuel, for instance, in the synthesis or biofuel, a redesign of the gas turbine combustor was necessary. For each type of fuel, a different kind of optimization

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 119

To compensate for the lower heating value (LHV) of fuel gases, the fuel injection system must provide a much higher fuel rate than when operating with high heating values. Due to the high rate of mass flow of gas with LHV, the passage of fuel has a much larger cross sec‐ tion than the section corresponding to natural gas. Fuel pipes, control valves and stop valves have larger diameters and shall be designed to include an additional fuel blend, which con‐ sists of the final mixture of the recovered gas with natural gas and steam. The pressure drops and the size of the air spiral entering the flame tube must be adjusted to optimize the combustion process. The system must have high safety standards, so the flanges and the gaskets of the combustor and its connections must be safely welded. The system for low

For safety reasons, the loading of the gas turbine to the rated load is accomplished through the use of the fuel reserve. The procedure for replacing the fuel reserve to the main tank is

The performance of a gas turbine is related to the local conditions of the installation and the

Due to the diesel low solubility at low temperature, tests with ethanol were performed with‐

According to the measuring methodology to be adopted to test gas turbines operating on liquids fuels, the physical-chemical properties of ethanol and diesel are shown in Table 4.

Table 4 also shows the fuel requirements established by the manufacturer of the tested gas turbine along with ASTM D6751 standard specifications for the testing of thermal perform‐

ance. Regarding emissions a standard ISO 11042-1:1996 was used (NWAFOR, 2004).

environment, where pressure and temperature conditions are of great importance.

out premix, and without the use of additives, which increased the cost of fuel.

**Table 2.** Engine Performance data at ISO Condition.

For tracking and measuring the tests parameters a type of supervisory software was used in the test bench (given by the turbine manufacturer) along with the data acquisition and the post processing obtained during the tests.

The composition of the exhaust gases was measured in real time using an Ecoline 6000 gas analyzer, reporting the concentration of O2, CO2 and hydrocarbons (HC) in volume percent‐ age (%v/v) and NO, CO, NO2 and SO2 (ppm) (Sierra, 2008). The fuel high heating value (HHV) was determined by a C-2000 IKA WORKS calorimeter. The accuracy, range and reso‐ lution of each instrument used during the tests are shown in Table 3.


**Table 3.** Accuracy of the measuring instruments.

#### **5.1. Adjustments to the microturbine**

Fuel Pressure 350 kPa

Thermal Efficiency 26% (± 2)

Power Output 29 kW NET (± 1)

Fuel HHV 45,144 kJ/kg

Fuel Flow 12 l/h

Inlet Air Flow 16 Nm³/min

Rotational Speed 96000 rpm

For tracking and measuring the tests parameters a type of supervisory software was used in the test bench (given by the turbine manufacturer) along with the data acquisition and the

The composition of the exhaust gases was measured in real time using an Ecoline 6000 gas analyzer, reporting the concentration of O2, CO2 and hydrocarbons (HC) in volume percent‐ age (%v/v) and NO, CO, NO2 and SO2 (ppm) (Sierra, 2008). The fuel high heating value (HHV) was determined by a C-2000 IKA WORKS calorimeter. The accuracy, range and reso‐

**Instrument Range Resolution Accuracy**

± 10 < 300 ± 4 (%) rdg < 2000 ± 10 (%) rdg "/> 2000

± 5 < 100 ± 4 (%) rdg < 3000

Fuel Flow 0-100 (l/h) 1.0 (ml) ±1.0 (%) scale Temperature 0-350 (°C) 0.31 (°C) ±0.8 (%) scale Pressure 0-10 (bar) 0.01 (bar) ±1.0 (%) scale Power 0-45 (kW) 0.05 (kW) ±0.5 (%) scale Calorimeter -- -- ±0.5 (%)

CO (ppm) 0 - 20000 1

NOx (ppm) 0 - 4000 1

lution of each instrument used during the tests are shown in Table 3.

Pressure Ratio 4

**Table 2.** Engine Performance data at ISO Condition.

118 Progress in Gas Turbine Performance

post processing obtained during the tests.

Gas Analyzer

**Table 3.** Accuracy of the measuring instruments.

Exhaust Temperature 260 °C

Due to impurities in the gas or fuel, for instance, in the synthesis or biofuel, a redesign of the gas turbine combustor was necessary. For each type of fuel, a different kind of optimization was needed, in relation to the fuel low heating value (LHV).

To compensate for the lower heating value (LHV) of fuel gases, the fuel injection system must provide a much higher fuel rate than when operating with high heating values. Due to the high rate of mass flow of gas with LHV, the passage of fuel has a much larger cross sec‐ tion than the section corresponding to natural gas. Fuel pipes, control valves and stop valves have larger diameters and shall be designed to include an additional fuel blend, which con‐ sists of the final mixture of the recovered gas with natural gas and steam. The pressure drops and the size of the air spiral entering the flame tube must be adjusted to optimize the combustion process. The system must have high safety standards, so the flanges and the gaskets of the combustor and its connections must be safely welded. The system for low LHV must include:


For safety reasons, the loading of the gas turbine to the rated load is accomplished through the use of the fuel reserve. The procedure for replacing the fuel reserve to the main tank is done automatically.

#### **5.2. Tests on gas turbine using liquid fuel**

The performance of a gas turbine is related to the local conditions of the installation and the environment, where pressure and temperature conditions are of great importance.

Due to the diesel low solubility at low temperature, tests with ethanol were performed with‐ out premix, and without the use of additives, which increased the cost of fuel.

According to the measuring methodology to be adopted to test gas turbines operating on liquids fuels, the physical-chemical properties of ethanol and diesel are shown in Table 4.

Table 4 also shows the fuel requirements established by the manufacturer of the tested gas turbine along with ASTM D6751 standard specifications for the testing of thermal perform‐ ance. Regarding emissions a standard ISO 11042-1:1996 was used (NWAFOR, 2004).


**6. Performance evaluation**

**6.1. Natural gas**

ior of microturbines operating at partial and full load.

The performance showed in this study was obtained from experimental tests at the Gas Tur‐ bine Laboratory of the Federal University of Itajubá (GOMES, 2002). Both natural gas and liquid fuel Capstone microturbines and their respective fuel supplying and electrical con‐ nection systems were installed and a property measurement was used to obtain the behav‐

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 121

The microturbine tested on natural gas was a Capstone 330 High Pressure. Table 5 gives the technical information of this machine and the features of the natural gas used in the tests. The natural gas microturbine was tested on the stand-alone mode supplying a resistive load. These microturbines can record operational parameters (temperatures, pressures, fuel usage, turbine speed, internal voltages/currents, status, and many others). Such data can be ac‐ cessed with a computer or modem connected to an RS-232 port on the microturbine. To sup‐

**CAPSTONE Microturbine Features**

**Proprieties of Natural Gas (20 ºC and 1 atm)**

**Ambient Conditions**

A large battery started the microturbine when disconnected from the grid, preventing any sudden load increase or decrease in the electrical buffer during the stand-alone operation (Capstone, 2001). The start-up took about 2 minutes and the speed was increased from 0 (zero) to 45,000 rpm, occasion when the microturbine started generating electricity. The ro‐

Low Heat Value 36,145 kJ/m3 High Heat Value 40,025 kJ/m3

Elevation 800 meters Average Temperature 30 ºC

plement these data, additional instrumentation was installed for the tests.

Model 330 (High Pressure)

Specific Mass 0.6165

Full-Load Power (ISO Conditions) 30 kW Fuel Natural Gas Fuel Pressure 358 – 379 kPa Fuel Flow\* 12 m3/h Efficiency (LHV)\* 27%

**Table 5.** General conditions of the analysis

**Table 4.** Ethanol and diesel physical-chemical characteristics.

The experimental determination of the ethanol heating value, kinematic viscosity and density were carried out according to ISO 1928-1976 and ASTM D1989-91 standards (ASME, 1997).

The use of different fuels implies the need of mass flow rate adjustments, according to its LHV and density, as without these adjustments, once established a load, the supply system would feed a quantity of fuel depending on the characteristics of the standard fuel (diesel). If the LHV of the new fuel is lower than standard, the gas turbine power could not reach the required demand.

Initially, the engine operated with conventional diesel fuel for a period of 20 minutes to reach a steady state condition for a load of 10 kW. After 20 minutes, the mass flow rates were changed to the fuel corresponding values. At this stage the fuel started to be replaced in order to increase the content of ethanol, by closing the diesel inlet valve and opening the ethanol valve. In order to ensure that all existing diesel power on the engine internal circui‐ try would be consumed, the engine was left running for 10 minutes with the same load op‐ eration, that is, 10 kW.

In order to check if the fuels were able to supply the engine, without causing problems to the fuel injection system, the kinematic viscosity of each fuel was measured. The composi‐ tion of gas emissions and thermal parameters were also measured in total and average load for each fuel. This whole procedure was performed for the engine operating with loads of 5, 10, 15, 20, 25 and 30 kW in a grid connection mode.

Afterwards the emissions were measured with a gas analyzer, and the load of 5 kW in‐ creased. Ten minutes were necessary until it reached steady state again. Exhaust emissions were measured from the exhaust gases and, as mentioned before, the thermal performance data were stored in a personal computer (PC) unit coupled with a PLC (Programmable Log‐ ic Controller) data acquisition system, which carried out the data reading at every second.

When tests with ethanol were over, the engine was left running, in order to accomplish the purging of the remaining fuel. After that the engine was once again operated with diesel for ten minutes, and then disconnected and stopped.

#### **6. Performance evaluation**

The performance showed in this study was obtained from experimental tests at the Gas Tur‐ bine Laboratory of the Federal University of Itajubá (GOMES, 2002). Both natural gas and liquid fuel Capstone microturbines and their respective fuel supplying and electrical con‐ nection systems were installed and a property measurement was used to obtain the behav‐ ior of microturbines operating at partial and full load.

#### **6.1. Natural gas**

**Properties Ethanol Diesel Fuel Limits ASTM D6751**

Sulfur (% mass) 0 0.20 0.05 < < 0.05

Flash Point (°C) 13 60 38 - 66 "/> 130 Water (% Volume) 0.05 0.05 0.05 0.05

The experimental determination of the ethanol heating value, kinematic viscosity and density were carried out according to ISO 1928-1976 and ASTM D1989-91 standards

The use of different fuels implies the need of mass flow rate adjustments, according to its LHV and density, as without these adjustments, once established a load, the supply system would feed a quantity of fuel depending on the characteristics of the standard fuel (diesel). If the LHV of the new fuel is lower than standard, the gas turbine power could not reach the

Initially, the engine operated with conventional diesel fuel for a period of 20 minutes to reach a steady state condition for a load of 10 kW. After 20 minutes, the mass flow rates were changed to the fuel corresponding values. At this stage the fuel started to be replaced in order to increase the content of ethanol, by closing the diesel inlet valve and opening the ethanol valve. In order to ensure that all existing diesel power on the engine internal circui‐ try would be consumed, the engine was left running for 10 minutes with the same load op‐

In order to check if the fuels were able to supply the engine, without causing problems to the fuel injection system, the kinematic viscosity of each fuel was measured. The composi‐ tion of gas emissions and thermal parameters were also measured in total and average load for each fuel. This whole procedure was performed for the engine operating with loads of 5,

Afterwards the emissions were measured with a gas analyzer, and the load of 5 kW in‐ creased. Ten minutes were necessary until it reached steady state again. Exhaust emissions were measured from the exhaust gases and, as mentioned before, the thermal performance data were stored in a personal computer (PC) unit coupled with a PLC (Programmable Log‐ ic Controller) data acquisition system, which carried out the data reading at every second.

When tests with ethanol were over, the engine was left running, in order to accomplish the purging of the remaining fuel. After that the engine was once again operated with diesel for

Kinematic Viscosity @40 °C (mm²/s) 1.08 1.54 1.9 – 4.1 1.9 – 6 Density @ 25 °C (g/cm³) 0.786 0.838 0.75 – 0.95 -

LHV (kJ/kg) 23,985.00 42,179.27

**Table 4.** Ethanol and diesel physical-chemical characteristics.

10, 15, 20, 25 and 30 kW in a grid connection mode.

ten minutes, and then disconnected and stopped.

(ASME, 1997).

120 Progress in Gas Turbine Performance

required demand.

eration, that is, 10 kW.

The microturbine tested on natural gas was a Capstone 330 High Pressure. Table 5 gives the technical information of this machine and the features of the natural gas used in the tests. The natural gas microturbine was tested on the stand-alone mode supplying a resistive load. These microturbines can record operational parameters (temperatures, pressures, fuel usage, turbine speed, internal voltages/currents, status, and many others). Such data can be ac‐ cessed with a computer or modem connected to an RS-232 port on the microturbine. To sup‐ plement these data, additional instrumentation was installed for the tests.


**Table 5.** General conditions of the analysis

A large battery started the microturbine when disconnected from the grid, preventing any sudden load increase or decrease in the electrical buffer during the stand-alone operation (Capstone, 2001). The start-up took about 2 minutes and the speed was increased from 0 (zero) to 45,000 rpm, occasion when the microturbine started generating electricity. The ro‐ tating components of the microturbine were mounted on a single shaft supported by air bearings and a spin at up 96,000 rpm. Figure 9 shows the speed behavior with the microtur‐ bine power output.

**Figure 11.** Microturbine efficiency at partial loads.

Figure 12 shows CO and NOX emissions behavior of a Capstone natural gas microturbine. Combustion occurs in three different steps. The first step is from start-up to about 5 kW. At

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 123

The second step is between 5 and 20 kW, as shown in Figure 12. In the second step the CO formation decreases continuously while emissions of NOX decrease at first, though increasing but it returns to increase softly slightly up to 113 ppmv. The last step begins at this point. At this step the lean-premix combustion occurs and the NOX formation di‐

Emissions of CO2 depend on the fuel type and the system efficiency. Figure 13 shows CO2

this step CO formation decreases and emissions of NOX increase quickly.

**Figure 12.** CO and NOX emissions of a natural gas microturbine at partial loads.

emissions of a Capstone natural gas microturbine.

minishes to 5 ppmv.

**Figure 9.** Microturbine speed at partial loads.

Capstone microturbine includes a recuperator which allows the microturbine efficiency to be improved. Figure 10 and 11 show respectively, the exhaust temperature and the efficien‐ cy behavior at partial loads. 27 % efficiency is possible at full load.

**Figure 10.** Microturbines exhaust temperature at partial loads.

**Figure 11.** Microturbine efficiency at partial loads.

tating components of the microturbine were mounted on a single shaft supported by air bearings and a spin at up 96,000 rpm. Figure 9 shows the speed behavior with the microtur‐

Capstone microturbine includes a recuperator which allows the microturbine efficiency to be improved. Figure 10 and 11 show respectively, the exhaust temperature and the efficien‐

cy behavior at partial loads. 27 % efficiency is possible at full load.

**Figure 10.** Microturbines exhaust temperature at partial loads.

bine power output.

122 Progress in Gas Turbine Performance

**Figure 9.** Microturbine speed at partial loads.

Figure 12 shows CO and NOX emissions behavior of a Capstone natural gas microturbine. Combustion occurs in three different steps. The first step is from start-up to about 5 kW. At this step CO formation decreases and emissions of NOX increase quickly.

**Figure 12.** CO and NOX emissions of a natural gas microturbine at partial loads.

The second step is between 5 and 20 kW, as shown in Figure 12. In the second step the CO formation decreases continuously while emissions of NOX decrease at first, though increasing but it returns to increase softly slightly up to 113 ppmv. The last step begins at this point. At this step the lean-premix combustion occurs and the NOX formation di‐ minishes to 5 ppmv.

Emissions of CO2 depend on the fuel type and the system efficiency. Figure 13 shows CO2 emissions of a Capstone natural gas microturbine.

The liquid fuel microturbine was tested on the grid connect mode. These data can be accessed with a computer or modem connected to an RS-232 port on the microturbine. To supplement these data, additional instrumentation was installed for the tests. Figure 14 shows the turbine exit temperature and the exhaust temperature at partial loads. These temperatures are before

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 125

Figure 15 shows the liquid fuel microturbine efficiency at partial loads. Up to 24.5 % effi‐ ciency is possible at full load while the microturbine efficiency is at its highest when Cap‐

stone microturbines operate over an output range between 12 kW and full load.

and after the recuperator were used and their difference ranges from 300 to 450 ºC.

**Figure 14.** Microturbine exit and exhaust temperature at partial loads.

**Figure 15.** Microturbine efficiency at partial loads

**Figure 13.** CO2 emissions of a natural gas microturbine at partial loads.

#### **6.2. Liquid fuel**

The microturbine tested on diesel was a Capstone 330 Liquid Fuel. Table 6 gives the techni‐ cal information of this machine and the features of the diesel used in the tests.


**Table 6.** General conditions of the analysis

The liquid fuel microturbine was tested on the grid connect mode. These data can be accessed with a computer or modem connected to an RS-232 port on the microturbine. To supplement these data, additional instrumentation was installed for the tests. Figure 14 shows the turbine exit temperature and the exhaust temperature at partial loads. These temperatures are before and after the recuperator were used and their difference ranges from 300 to 450 ºC.

**Figure 14.** Microturbine exit and exhaust temperature at partial loads.

**Figure 13.** CO2 emissions of a natural gas microturbine at partial loads.

The microturbine tested on diesel was a Capstone 330 Liquid Fuel. Table 6 gives the techni‐

**CAPSTONE Microturbine Features**

**Proprieties of Liquid Fuel (20 ºC and 1 atm)**

**Ambient Conditions**

Low Heat Value 42,923 kJ/kg High Heat Value 45,810 kJ/kg

Elevation 800 meters Average Temperature 30 ºC

cal information of this machine and the features of the diesel used in the tests.

Model 330 (Liquid Fuel)

Fuel Diesel #2 (ASTM D975)

Full-Load Power\* 29 kW

Fuel Pressure 35 – 70 kPa Fuel Flow\* 12.5 l/h Efficiency (LHV)\* 26%

Specific Mass 0.848

**6.2. Liquid fuel**

124 Progress in Gas Turbine Performance

\* ISO Conditions

**Table 6.** General conditions of the analysis

Figure 15 shows the liquid fuel microturbine efficiency at partial loads. Up to 24.5 % effi‐ ciency is possible at full load while the microturbine efficiency is at its highest when Cap‐ stone microturbines operate over an output range between 12 kW and full load.

**Figure 15.** Microturbine efficiency at partial loads

Figure 16 shows the CO and NOX emissions behavior of a Capstone liquid fuel microtur‐ bine. The CO formation decreases, whereas emissions of NOX increase as the power output increases due to a rise in the flame temperature.

**7. Microturbines and Internal combustion engine´s emissions**

which allows for a complete burning.

SCR: Selective Catalytic Reduction.

\* ISO Conditions; \*\* On Site Conditions (See Table 1)

**Table 8.** CO and NOX emissions of Capstone microturbines

**8. Case studies under Brazilian conditions**

**ICE**

**Natural Gas Without Control**

**Table 7.** NOX emissions of internal combustion engines (ICE) (Weston, *et. al*., 2001)

**FUEL Natural Gas Diesel**

Efficiency\* % (LHV) 27 26 Nominal Power\* kW 30 29 CO (@15%O2)\*\* g/MWh 210 80 NOX (@15%O2)\*\* g/MWh 520 280

consumers are considering producing their own electricity (Gomes, 2002).

tion and a cogeneration system supplying buildings in a residential segment

Due to the Brazilian governmental incentive to develop the gas industry, the feasibility of many natural gas applications has been doubted. Consequently, the demand for efficiently and environmentally friendly power generation technologies has increased. Many electricity

This study analyses the possibility of natural gas application with Capstone microturbines in three cases of power generation: peak shaving in a small industry, base load in a gas sta‐

Nowadays it is a trend on microturbines market to reduce investments. This paper analyses the influence of the investment cost of microturbines on the feasibility and cost of the gener‐

Table 7 and 8 compare emissions data from internal combustion engines and microturbines. In the absence of a post combustion device, such as a catalytic converter, reciprocating en‐ gines can have very high emission levels. Emission levels of microturbines are lower than levels of internal combustion engines as microturbines combustion is a continuous process

**ICE**

Efficiency % (HHV) 36% 29% 38% 38% Nominal Power kW 1,000 1,000 1,000 1,000 NOX (@15%O2) g/MWh 998 227 9,888 2,132

**Natural Gas SCR**

**ICE Diesel**

**Without Control**

**ICE Diesel SCR**

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 127

**Figure 16.** CO and NOX emissions from liquid fuel microturbine at partial loads.

Figure 17 shows the CO2 and SO2 emissions of a Capstone liquid fuel microturbine. The emissions depend considerably on the liquid fuel features. While SO2 emissions are an im‐ portant emission category for traditional electric utility companies, they are expected to be negligible for distributed generation technologies.

**Figure 17.** CO2 and SO2 emissions from liquid fuel microturbine at partial loads.

#### **7. Microturbines and Internal combustion engine´s emissions**

Table 7 and 8 compare emissions data from internal combustion engines and microturbines. In the absence of a post combustion device, such as a catalytic converter, reciprocating en‐ gines can have very high emission levels. Emission levels of microturbines are lower than levels of internal combustion engines as microturbines combustion is a continuous process which allows for a complete burning.


SCR: Selective Catalytic Reduction.

Figure 16 shows the CO and NOX emissions behavior of a Capstone liquid fuel microtur‐ bine. The CO formation decreases, whereas emissions of NOX increase as the power output

Figure 17 shows the CO2 and SO2 emissions of a Capstone liquid fuel microturbine. The emissions depend considerably on the liquid fuel features. While SO2 emissions are an im‐ portant emission category for traditional electric utility companies, they are expected to be

increases due to a rise in the flame temperature.

126 Progress in Gas Turbine Performance

**Figure 16.** CO and NOX emissions from liquid fuel microturbine at partial loads.

**Figure 17.** CO2 and SO2 emissions from liquid fuel microturbine at partial loads.

negligible for distributed generation technologies.

**Table 7.** NOX emissions of internal combustion engines (ICE) (Weston, *et. al*., 2001)


\* ISO Conditions; \*\* On Site Conditions (See Table 1)

**Table 8.** CO and NOX emissions of Capstone microturbines

#### **8. Case studies under Brazilian conditions**

Due to the Brazilian governmental incentive to develop the gas industry, the feasibility of many natural gas applications has been doubted. Consequently, the demand for efficiently and environmentally friendly power generation technologies has increased. Many electricity consumers are considering producing their own electricity (Gomes, 2002).

This study analyses the possibility of natural gas application with Capstone microturbines in three cases of power generation: peak shaving in a small industry, base load in a gas sta‐ tion and a cogeneration system supplying buildings in a residential segment

Nowadays it is a trend on microturbines market to reduce investments. This paper analyses the influence of the investment cost of microturbines on the feasibility and cost of the gener‐ ated electricity, being the cost of fuel a significant part of the electricity final price. The feasi‐ bility and the cost of the electricity generated with fuel were also assessed. This study used electric energy and natural gas prices charged by several electric power utility companies and gas distributors in Brazil at the time this study was being carried out (November, 2002). Table 9 shows the general conditions used in the cases studies.

**Model of microturbine Capstone 330**

Life time of microturbines 20 years

Net power (peak load) 28 kW

Microturbine installed cost 1.538 US\$/kW

Natural gas consumption (HHV) 650 m3/month

0.33-1.32 R\$/m3

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 129

Number of microturbines 1

Average price of natural gas (taxes included)

**Figure 18.** Electricity demand supplied by utility companies.

**Table 11.** Conditions of the peak shaving case


**Table 9.** General conditions of the analysis

#### **8.1. Peak shaving case**

Many consumers try to reduce their electricity consumption at peak hours due to its high price. If they can produce their electricity, they will reduce the amount of electricity pur‐ chased from utility companies at peak hours, without having to reduce their electricity con‐ sumption. Besides, power generation systems can improve the quality and reliability of the energy supplied by utility companies.

A study was carried out in four Brazilian regions, classified according to the price of natural gas charged by gas distributors of these regions, as shown in Table 10. Table 11 and Figure 18 show the conditions studied and the electricity demand supplied by utility companies with and without peak shaving.


**Table 10.** Brazilian regions analyzed in the peak shaving case

Commercial microturbines available in the Brazilian market are imported from the USA and investments feasibility depends on the currency rate, as can be seen in Table 9.


**Table 11.** Conditions of the peak shaving case

ated electricity, being the cost of fuel a significant part of the electricity final price. The feasi‐ bility and the cost of the electricity generated with fuel were also assessed. This study used electric energy and natural gas prices charged by several electric power utility companies and gas distributors in Brazil at the time this study was being carried out (November, 2002).

**CAPSTONE Microturbine Features**

**Proprieties of Natural Gas (20 ºC and 1 atm)**

Many consumers try to reduce their electricity consumption at peak hours due to its high price. If they can produce their electricity, they will reduce the amount of electricity pur‐ chased from utility companies at peak hours, without having to reduce their electricity con‐ sumption. Besides, power generation systems can improve the quality and reliability of the

A study was carried out in four Brazilian regions, classified according to the price of natural gas charged by gas distributors of these regions, as shown in Table 10. Table 11 and Figure 18 show the conditions studied and the electricity demand supplied by utility companies

Commercial microturbines available in the Brazilian market are imported from the USA and

1st Region São Paulo (SP) and Rio de Janeiro (RJ) 2nd Region Ceará (CE), Pernambuco (PE) and Paraíba (PB)

investments feasibility depends on the currency rate, as can be seen in Table 9.

3rd Region Rio Grande do Norte (RN)

4th Region Others

**Table 10.** Brazilian regions analyzed in the peak shaving case

**Brazilian States**

High Heat Value 39,304 kJ/m3

Currency rate 2.6 R\$/US\$ Interest rate 10 % per year

Model 330 (High Pressure)

Fuel Natural Gas

Specific Mass 0.602

**Table 9.** General conditions of the analysis

energy supplied by utility companies.

with and without peak shaving.

**8.1. Peak shaving case**

128 Progress in Gas Turbine Performance

Table 9 shows the general conditions used in the cases studies.

#### **Figure 18.** Electricity demand supplied by utility companies.

Table 12 displays the economical analysis of the peak shaving case. The investment is not feasible yet, as the payback period is very long. Rio Grande do Norte is the state where this business would be most interesting as payback is 8 years.

**8.2. Base load case**

Average price of natural gas

**Table 13.** Conditions of the base load case.

**Figure 20.** Electricity demand supplied by utility companies.

(taxes included)

In this case, a microturbine produces electricity to a gas station according to the base load demand, as shows Figure 20. The conditions of this case are in table 13, whereas Table 14

shows the Brazilian regions analyzed in the base load case.

Number of microturbines 1

**Model of microturbine Capstone 330**

Life time of microturbines 10 years

Net power 27,5 kW

Microturbine installed cost 1,538 US\$/kW

Natural gas consumption (HHV) 6,918 m3/month

0.24 - 1.02 R\$/m3

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 131


**Table 12.** Economical analysis of the peak shaving case

Figure 19 shows payback period in relation to microturbine cost. There is a strong fall on the payback period of the states of SP and RJ, due to a decrease in the microturbine cost.

A few manufactures intend to decrease microturbine costs to about 400 US\$/kW until 2005 (Dunn & Flavin, 2000). If the microturbine cost is 400 US\$/kW, the payback period will be between 2.5 and 5 years, as shown in Figure 19.

**Figure 19.** The influence of the microturbine cost on the return on investments.

#### **8.2. Base load case**

Table 12 displays the economical analysis of the peak shaving case. The investment is not feasible yet, as the payback period is very long. Rio Grande do Norte is the state where this

Total Investment \* US\$ 46,827 46,827 46,827 46,827 Annual cost\*\* US\$/year 53,827 44,906 55,718 51,102 Annual cost\* US\$/year 55,323 45,231 53,950 51,479 Annual savings US\$/year -1,497 -325 1,769 -378 Electricity generated US\$/MWh 435 321 301 366 Payback Period years 32 15 8 15

Figure 19 shows payback period in relation to microturbine cost. There is a strong fall on the

A few manufactures intend to decrease microturbine costs to about 400 US\$/kW until 2005 (Dunn & Flavin, 2000). If the microturbine cost is 400 US\$/kW, the payback period will be

payback period of the states of SP and RJ, due to a decrease in the microturbine cost.

**SP - RJ CE-PE-PB RN Other States**

business would be most interesting as payback is 8 years.

\* With peak shaving; \*\* Without peak shaving

130 Progress in Gas Turbine Performance

**Table 12.** Economical analysis of the peak shaving case

between 2.5 and 5 years, as shown in Figure 19.

**Figure 19.** The influence of the microturbine cost on the return on investments.

In this case, a microturbine produces electricity to a gas station according to the base load demand, as shows Figure 20. The conditions of this case are in table 13, whereas Table 14 shows the Brazilian regions analyzed in the base load case.


**Table 13.** Conditions of the base load case.

**Figure 20.** Electricity demand supplied by utility companies.


**Table 14.** Brazilian regions analyzed in the base load case.

Table 15 displays the economical analysis of the base load case for gas stations. Up to the present moment this kind of business is not feasible, except in the state of Rio Grande do Norte (RN) where payback period can be 3.1 years, once local gas distribution companies have encouraged thermoelectric small scale power generation, according to natural gas price lower than others kind of fuels.


**Figure 21.** Cost of the electricity generated for different microturbine costs and natural gas average prices.

natural gas average price of 0.24 R\$/m3

**Figure 22.** The natural gas average price influence on the payback period.

In the base load case, the natural gas average price is the most influential component in the return on investments. Figure 22 shows this conclusion for the microturbine cost at this moment, since

can result in a payback period between 3 and 4 years.

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 133

\* With power generation; \*\* Without power generation

**Table 15.** Economical analysis of the base load case.

Figure 21 shows the behavior of the cost of the electricity generated for different micro‐ turbine costs and natural gas average price. Some natural gas distribution companies in Brazil have encouraged the creation of small thermal power generation units, as the cost of natural gas coming from these companies would be about 0.24 R\$/m3 . Based on this fact and on the perspective of microturbine manufactures, Figure 21 shows the cost of the electricity generated could be 58 US\$/MWh. For each 1 US\$/kW decreased from the microturbine cost, the cost of the electricity generated decreases about 0,021 US\$/MWh, for every natural gas average price range, and for each 1 R\$/m3 decreased from the natu‐ ral gas average price, the cost of the electricity generated decreases about 135 US\$/MWh, for every microturbine cost range.

**Brazilian States**

**SP e RJ RS e PR RN Other States**

. Based on this

1st Region São Paulo (SP) and Rio de Janeiro (RJ)

2nd Region Rio Grande do Sul (RS) and Paraná (PR)

Table 15 displays the economical analysis of the base load case for gas stations. Up to the present moment this kind of business is not feasible, except in the state of Rio Grande do Norte (RN) where payback period can be 3.1 years, once local gas distribution companies have encouraged thermoelectric small scale power generation, according to natural gas price

Total Investment US\$ 46827 46827 46827 46827

Annual cost\*\* US\$/year 28748 28117 27956 24389

Annual cost\* US\$/year 45699 33898 20707 26002

Annual savings US\$/year - 16951 - 5780 7249 - 1614

Electricity generated US\$/MWh 181 131 75 99

Payback Period years Not Feasible Not Feasible 3,1 8,2

Figure 21 shows the behavior of the cost of the electricity generated for different micro‐ turbine costs and natural gas average price. Some natural gas distribution companies in Brazil have encouraged the creation of small thermal power generation units, as the cost

fact and on the perspective of microturbine manufactures, Figure 21 shows the cost of the electricity generated could be 58 US\$/MWh. For each 1 US\$/kW decreased from the microturbine cost, the cost of the electricity generated decreases about 0,021 US\$/MWh, for every natural gas average price range, and for each 1 R\$/m3 decreased from the natu‐ ral gas average price, the cost of the electricity generated decreases about 135 US\$/MWh,

of natural gas coming from these companies would be about 0.24 R\$/m3

3rd Region Rio Grande do Norte (RN)

4th Region Others

**Table 14.** Brazilian regions analyzed in the base load case.

\* With power generation; \*\* Without power generation

**Table 15.** Economical analysis of the base load case.

for every microturbine cost range.

lower than others kind of fuels.

132 Progress in Gas Turbine Performance

**Figure 21.** Cost of the electricity generated for different microturbine costs and natural gas average prices.

In the base load case, the natural gas average price is the most influential component in the return on investments. Figure 22 shows this conclusion for the microturbine cost at this moment, since natural gas average price of 0.24 R\$/m3 can result in a payback period between 3 and 4 years.

**Figure 22.** The natural gas average price influence on the payback period.

#### **8.3. Cogeneration case**

In this case, two microturbines and a heat recovery system produced electricity and hot wa‐ ter to buildings in a residential segment, according to the base load demand, as can be seen in Figure 23.

System cogeneration model MG2-C1

Number of Capstone microturbines 2

Number of heat recovery systems 1

Heat recovery systems (hot water generation)

**Table 16.** Conditions of the cogeneration case.

Life time of microturbines 10 years

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 135

Power output 54 kW

Water pressure 10 bar

Water flow 2.22 t/h

Inlet water temperature 25 ºC

Outlet water temperature 67 ºC

Outlet exit gas temperature 93 ºC

Net power 53 kW

System cogeneration installed cost 1,872 US\$/kW

Natural gas consumption (HHV) 13,653 m3/day

Average price of natural gas (taxes included) 0.24 - 0.90 R\$/m3

Table 18 displays the economical analysis of the cogeneration case. Investments costs are lower in the conventional case than in the cogeneration system, and, although the annual cost is higher, savings can be up to US\$ 24,907 per year. The payback period is between 2.8

1st Region Rio de Janeiro (RJ) 2nd Region Paraná (PR)

3rd Region Rio Grande do Norte (RN)

and 3.8 years and the minimal cost of the electricity generated is 84 US\$/MWh.

4th Region Others

**Table 17.** Brazilian regions analyzed in the cogeneration case.

**Brazilian States**

**Figure 23.** Electric demand supplied by utility companies to consumers with and without cogeneration.

A cogeneration plant can result in substantial savings of energy. However, these systems usually result in greater capital expenditures than non-cogeneration plants. This incremental capital investment for cogeneration must be justified by reduced annual energy costs and re‐ duced payback periods.

A course of action involving minimum capital expenditures can be determined as the con‐ ventional case. In this study a low pressure boiler supplying process heat and the purchase of all electric power from utility system is the conventional case. Although the conventional case has the lowest investment cost, it usually has annual operating costs significantly high‐ er than those available with cogeneration alternatives. Table 16 shows the conditions of this case, while Table 17 shows the Brazilian regions analyzed in the base load case.


**Table 16.** Conditions of the cogeneration case.

**8.3. Cogeneration case**

134 Progress in Gas Turbine Performance

duced payback periods.

in Figure 23.

In this case, two microturbines and a heat recovery system produced electricity and hot wa‐ ter to buildings in a residential segment, according to the base load demand, as can be seen

**Figure 23.** Electric demand supplied by utility companies to consumers with and without cogeneration.

case, while Table 17 shows the Brazilian regions analyzed in the base load case.

A cogeneration plant can result in substantial savings of energy. However, these systems usually result in greater capital expenditures than non-cogeneration plants. This incremental capital investment for cogeneration must be justified by reduced annual energy costs and re‐

A course of action involving minimum capital expenditures can be determined as the con‐ ventional case. In this study a low pressure boiler supplying process heat and the purchase of all electric power from utility system is the conventional case. Although the conventional case has the lowest investment cost, it usually has annual operating costs significantly high‐ er than those available with cogeneration alternatives. Table 16 shows the conditions of this


**Table 17.** Brazilian regions analyzed in the cogeneration case.

Table 18 displays the economical analysis of the cogeneration case. Investments costs are lower in the conventional case than in the cogeneration system, and, although the annual cost is higher, savings can be up to US\$ 24,907 per year. The payback period is between 2.8 and 3.8 years and the minimal cost of the electricity generated is 84 US\$/MWh.


**Table 18.** Economical analysis of the cogeneration case.

In the cogeneration case, the fuel cost is the most influential component on the return on in‐ vestment, similar to the base load case. Figure 24 shows fuel costs can represent up to 71% of the cost of the electricity generated.

**Figure 25.** Combined influence of microturbine cost and average price of natural gas on the payback period in the

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 137

**Figure 26.** Combined influence of microturbine cost and average price of natural gas on the payback period in the

states of Rio de Janeiro and Paraná.

state of Rio Grande do Norte.

**Figure 24.** Components of the cost of the electricity generated.

Figure 25, Figure 26 and Figure 27 show the combined influence of microturbine cost and the average price of natural gas on the return on investment in the states of Rio de Janeiro and Paraná (Figure 25), Rio Grande do Norte (Figure 26) and the other states (Figure 27). Based on the perspective of microturbine manufactures and with natural gas average price of 0.25 R\$/m3 , the payback period can be between 1.5 and 3 years.

**RJ PR RN Other States**

Total Investment\* US\$ 23077 23077 23077 23077

Total Investment\*\* US\$ 136797 136797 136797 136797

Annual cost\* US\$/year 128566 110022 90323 98328

Annual cost\*\* US\$/year 110337 96655 65416 77534

Annual savings US\$/year 18228 13367 24907 20795

Electricity generated US\$/MWh 174 146 84 112

Payback Period years 3,3 3,8 2,8 3,1

In the cogeneration case, the fuel cost is the most influential component on the return on in‐ vestment, similar to the base load case. Figure 24 shows fuel costs can represent up to 71% of

Figure 25, Figure 26 and Figure 27 show the combined influence of microturbine cost and the average price of natural gas on the return on investment in the states of Rio de Janeiro and Paraná (Figure 25), Rio Grande do Norte (Figure 26) and the other states (Figure 27). Based on the perspective of microturbine manufactures and with natural gas average price

, the payback period can be between 1.5 and 3 years.

\* Conventional; \*\* Cogeneration

136 Progress in Gas Turbine Performance

**Table 18.** Economical analysis of the cogeneration case.

**Figure 24.** Components of the cost of the electricity generated.

of 0.25 R\$/m3

the cost of the electricity generated.

**Figure 25.** Combined influence of microturbine cost and average price of natural gas on the payback period in the states of Rio de Janeiro and Paraná.

**Figure 26.** Combined influence of microturbine cost and average price of natural gas on the payback period in the state of Rio Grande do Norte.

Although results show microturbines are not feasible to provide energy at peak demand, in this case the microturbines can supply peak demand and improve the level of reliabil‐ ity of the electricity supplying, because they can provide stand-by capabilities should the

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 139

In the base load case this sort of business is feasible just in states of Brazil where natural gas distributing companies have encouraged small thermal power generation by natural gas with lower prices, since the price is the most influential cost component of the elec‐

The most feasible investment in microturbines is in the cogeneration case. In this case, economical feasibility is certain in all states of Brazil as cogeneration systems can pro‐ vide considerable annual savings. Besides, under the perspective of manufacturers, and with the incentive of natural gas distribution companies together with the rise in electric‐ ity prices of Brazilian utility companies, investments in microturbines for the next years

The authors would like to thank CAPES, FAPEMIG, FAPEPE and CNPq, for their financial

Marco Antônio Rosa do Nascimento, Lucilene de Oliveira Rodrigues, Eraldo Cruz dos Santos, Eli Eber Batista Gomes, Fagner Luis Goulart Dias,

Elkin Iván Gutiérrez Velásques and Rubén Alexis Miranda Carrillo

[1] Asme performance test code PTC-22-1997, Gas turbine power plants, 1997.

[2] Barker, T. Micros, Catalysts and Electronics, Power-Gen International 96, Turboma‐

[3] Biasi, V. de Low cost and high efficiency make 30 to 80 kW microturbines attractive,

[4] Bolszo, C. D., and Mcdonell, V. G., Emissions Optmization of a Biodiesel Fired Gas Turbine, Proceedings of the Combustion Institute, 32, LSEVIER, pp. 2949-2956, 2009.

Federal University of Itajubá – UNIFEI, Brazil

chinery, v. 38, n. 1, p. 19-21, 1997.

Gas Turbine World, Jan.-Feb., Southport, 1998.

electric grid fail.

tricity generated.

will be higher than currently.

**Acknowledgements**

support.

**Author details**

**References**

**Figure 27.** Combined influence of microturbine cost and average price of natural gas on the payback period in the other states.

#### **9. Conclusions**

The variable speed operation and the electric power conditioner increase part-load efficien‐ cy of microturbines as they allow for the improvement of part-load fuel savings, especially increased recuperator effectiveness at lower part-load airflows. The variable speed control improves part-load performance but requires a system able to sense load and optimize speed. According to the results shown in this study, the microturbines efficiency is at its highest when Capstone microturbines are operating over an output range between 12 kW and full load.

Capstone microturbines use clean combustion technology to achieve low emissions. Nitro‐ gen oxides (NOX) and carbon monoxide (CO) emission levels of these machines are lower than 7 ppmv@15%O2 at full load when these microturbines are fueled with natural gas.

Microturbines exhibit low emissions of all classes of pollutants and have environmental ben‐ efits as they release fewer emissions compared to other distributed generation technologies, like internal combustion engines. Besides, these units are clean enough to be placed in a community with residential and commercial buildings.

Microturbine generators have shown good perspectives for electricity distributed generation in small scales, once they have high reliability and simple design (high potential for large scale cheap manufacturing).

Although results show microturbines are not feasible to provide energy at peak demand, in this case the microturbines can supply peak demand and improve the level of reliabil‐ ity of the electricity supplying, because they can provide stand-by capabilities should the electric grid fail.

In the base load case this sort of business is feasible just in states of Brazil where natural gas distributing companies have encouraged small thermal power generation by natural gas with lower prices, since the price is the most influential cost component of the elec‐ tricity generated.

The most feasible investment in microturbines is in the cogeneration case. In this case, economical feasibility is certain in all states of Brazil as cogeneration systems can pro‐ vide considerable annual savings. Besides, under the perspective of manufacturers, and with the incentive of natural gas distribution companies together with the rise in electric‐ ity prices of Brazilian utility companies, investments in microturbines for the next years will be higher than currently.

#### **Acknowledgements**

The authors would like to thank CAPES, FAPEMIG, FAPEPE and CNPq, for their financial support.

#### **Author details**

**Figure 27.** Combined influence of microturbine cost and average price of natural gas on the payback period in the

The variable speed operation and the electric power conditioner increase part-load efficien‐ cy of microturbines as they allow for the improvement of part-load fuel savings, especially increased recuperator effectiveness at lower part-load airflows. The variable speed control improves part-load performance but requires a system able to sense load and optimize speed. According to the results shown in this study, the microturbines efficiency is at its highest when Capstone microturbines are operating over an output range between 12 kW

Capstone microturbines use clean combustion technology to achieve low emissions. Nitro‐ gen oxides (NOX) and carbon monoxide (CO) emission levels of these machines are lower than 7 ppmv@15%O2 at full load when these microturbines are fueled with natural gas.

Microturbines exhibit low emissions of all classes of pollutants and have environmental ben‐ efits as they release fewer emissions compared to other distributed generation technologies, like internal combustion engines. Besides, these units are clean enough to be placed in a

Microturbine generators have shown good perspectives for electricity distributed generation in small scales, once they have high reliability and simple design (high potential for large

community with residential and commercial buildings.

other states.

**9. Conclusions**

138 Progress in Gas Turbine Performance

and full load.

scale cheap manufacturing).

Marco Antônio Rosa do Nascimento, Lucilene de Oliveira Rodrigues, Eraldo Cruz dos Santos, Eli Eber Batista Gomes, Fagner Luis Goulart Dias, Elkin Iván Gutiérrez Velásques and Rubén Alexis Miranda Carrillo

Federal University of Itajubá – UNIFEI, Brazil

#### **References**


[5] Capstone Turbine Corporation, Capstone Low Emissions Microturbine Tecnology, White Paper, USA, 2000.

[21] Watts, J. H, Microturbines: a new class of gas turbine engine, Global Gas turbine

Micro Gas Turbine Engine: A Review http://dx.doi.org/10.5772/54444 141

[22] Weston, F., Seidman, N., L., James, C. Model Regulations for the Output of Specified Air Emissions from Smaller-Scale Electric Generation Resources, The Regulatory As‐

News, ASME-IGTI, v. 39, n. 1, p. 4-8, USA, 1999.

sistance Project, 2001.


[21] Watts, J. H, Microturbines: a new class of gas turbine engine, Global Gas turbine News, ASME-IGTI, v. 39, n. 1, p. 4-8, USA, 1999.

[5] Capstone Turbine Corporation, Capstone Low Emissions Microturbine Tecnology,

[6] Capstone Turbine Corporation, Capstone Microturbine Model 330 System Operation

[7] Capstone Turbine Corporation, Capstone Microturbine Product Catalog, USA, 2012: http://www.capstoneturbine.com/prodsol/products/, accessed at: 20/06/2012.

[8] Cohen, H., Rogers, G. F. C., and Saravanamuttoo, H. I. H., Gas Turbine Theory,

[9] Dunn, S. & Flavin, C., Dimensionando a Microenergia. In: Estado do Mundo 2000.

[10] Gomes, E. E. B. Análise Técnico-econômica e Experimental de Microturbinas a Gás Operando com Gás Natural e Óleo Diesel, Master Degree Thesis, Supervised by Nas‐

[11] Hamilton, S. L., Microturbines, Distributed Generation: a nontechnical guide, edited

[12] GRI - Gas Research Institute, The role of Distributed Generation in competitive ener‐ gy markets, Distributed Generation Forum, Gas Research Institute (GRI), 1999.

[13] Liss, W.E., Natural Gas Power Systems for the Distributed Generation Market. Pow‐ er-Gen International '99 Conference. CD-Rom. New Orleans, Louisiana, USA, 1999.

[14] Nascimento, M. A. R.; Santos, E. C., Biofuel and Gas Turbine Engines, Advances in Gas Turbine Technology, InTech, ISBN 978-953-307-611-9, chaper 6, 2011.

[15] Nwafor, O., Emission characteristics of Diesel engine operating on rapeseed methyl

[16] Pierce, J. L. Microturbine Distributed Generation Using Conventional and Waste Fuel, Cogeneration and On-Site Power Production, James & James Science Publish‐

[17] Rodgers, C.; Watts, J.; Thoren, D.; Nichols, K. & Brent, R. Microturbines, Distributed Generation – The Power Paradigm for the New Millennium, edited by Anne-Marie

[19] Scott, W. G. Micro Gas Turbine Cogeneration Applications, International Power and

[20] Sierra, R. G. A., Teste Experimental e Análise Técnico-Econômica do Uso de Biocom‐ bustíveis em uma Microturbina a Gás de Tipo Regenerativo; Dissertação de Mestra‐

Borbely & Jan F. Kreider, cap. 5, pp. 120 – 148, CRC Press LLC. USA, 2001a.

[18] Rodgers, G., and Saravanamutto, H., Gas Turbine Theory, Prentice Hall, 2001b.

ester. Renewable Energy, 29, pp. 119-29, 2004.

ers, p. 45, v. 3, Issue 1, Jan-Feb, 2002.

Light Co., USA, 2000.

do, UNIFEI, 2008.

cimento, M. A. R. and Lora, E. E. S. Federal University of Itajubá, 2002.

by Ann Chambers, cap. 3, pp. 33 – 72, PennWell Corporation, USA, 2001

White Paper, USA, 2000.

Manual, USA, 2001.

140 Progress in Gas Turbine Performance

Fourth edition, 1996.

Brazil, UMA Ed., 2000.

[22] Weston, F., Seidman, N., L., James, C. Model Regulations for the Output of Specified Air Emissions from Smaller-Scale Electric Generation Resources, The Regulatory As‐ sistance Project, 2001.

**Section 2**

**Gas Turbine Combustion**
