**Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC**

Takeharu Hasegawa *Central Research Institute of Electric Power Industry Japan* 

## **1. Introduction**

From the viewpoints of securing a stable supply of energy and protecting our global environment in the future, the integrated gasification combined cycle (IGCC) power generation of various gasifying methods has been introduced in the world. Gasified fuels are chiefly characterized by the gasifying agents and the synthetic gas cleanup methods and can be divided into four types. The calorific value of the gasified fuel varies according to the gasifying agents and feedstocks of various resources, and ammonia originating from nitrogenous compounds in the feedstocks depends on the synthetic gas clean-up methods. In particular, air-blown gasified fuels provide low calorific fuel of 4 MJ/m3 and it is necessary to stabilize combustion. In contrast, the flame temperature of oxygen-blown gasified fuel of medium calorie between approximately 9–13 MJ/m3 is much higher, so control of thermal-NOx emissions is necessary. Moreover, to improve the thermal efficiency of IGCC, hot/dry type synthetic gas clean-up is needed. However, ammonia in the fuel is not removed and is supplied into the gas turbine where fuel-NOx is formed in the combustor. For these reasons, suitable combustion technology for each gasified fuel is important. In this paper, I will review our developments of the gas turbine combustors for the three type gasified fuels produced from the following gasification methods through experiments using a small diffusion burner and the designed combustors' tests of the simulated gasified fuels.


Figure 1 provides an outline of a typical oxygen-blown IGCC system. In this system, raw materials such as coal and crude are fed into the gasifier by slurry feed or dry feed with nitrogen. The synthetic gas is cleaned through a dust removing and desulfurizing process. The cleaned synthetic gas is then fed into the high-efficiency gas turbine topping cycle, and the steam cycle is equipped to recover heat from the gas turbine exhaust. This IGCC system is similar to LNG fired gas turbine combined cycle generation, except for the gasification and the synthetic gas cleanup process, primarily. IGCC requires slightly more station service power than an LNG gas turbine power generation.

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 241

to accept the various IGCC systems, 1773K-class gas turbine combustors of medium-Btu fuels by wet-type or hot/dry-type synthetic gas cleanup methods have undergone study

The energy resources and geographical conditions of each country, along with the diversification of fuels used for the electric power industry (such as biomass, poor quality coal and residual oil), are most significant issues for IGCC gas turbine development, as has been previously described: The development of biomass-fueled gasification received considerable attention in the United States and northern Europe in the early 1980s (Kelleher, 1985), and the prospects for commercialization technology (Consonni, 1997) appear considerably improved at present. Paisley and Anson (1997) performed a comprehensive economical evaluation of the Battele biomass gasification process, which utilizes a hot-gas conditioning catalyst for dry synthetic gas cleanup. In northern Europe, fixed-bed gasification heating plants built in the 1980s had been in commercial operation; the available technical and economical operation data convinced small district heating companies that biomass or peat-fueled gasification heating plants in the size class of 5 MW were the most profitable (Haavisto, 1996). However, during the period of stable global economy and oil prices, non-fossil-fueled gasification received little interest. Then, in the early 2000s when the Third Conference of Parties to the United Nations Framework Convention on Climate Change (COP3) invoked mandatory carbon dioxide emissions reductions on countries, biomass-fueled gasification technology began to receive considerable attention as one alternative. With the exception of Japanese national research and development project, almost all of the systems using the oxygen-blown gasification are in their final stages for

(Hasegawa et al., 2003, 2007).

commencing commercial operations overseas.

**1.2 Progress in gas turbine combustion technologies for IGCCs** 

the development of a gas turbine for each gasification method.

The plant thermal efficiency has been improved by enhancing the turbine inlet temperature, or combustor exhaust temperature. The thermal-NOx emissions from the gas turbines increase, however, along with a rise in exhaust temperature. In addition, gasified fuel containing NH3 emits fuel-NOx when hot/dry gas cleanup equipment is employed. It is therefore viewed as necessary to adopt a suitable combustion technology for each IGCC in

Dixon-Lewis and Williams (1969), expounded on the oxidation characteristics of hydrogen and carbon monoxide in 1969. The body of research into the basic combustion characteristics of gasified fuel includes studies on the flammability limits of mixed gas, consisting of CH4 or H2 diluted with N2, Ar or He (Ishizuka & Tsuji, 1980); a review of the flammability and explosion limits of H2 and H2/CO fuels (Cohen, 1992); the impact of N2 on burning velocity (Morgan & Kane, 1952); the effect of N2 and CO2 on flammability limits (Coward & Jones, 1952; Ishibasi et al, 1978); and the combustion characteristics of low calorific fuel (Folsom, 1980; Drake, 1984); studies by Merryman et al. (1997), on NOx formation in CO flame; studies by Miller et al. (1984), on the conversion characteristics of HCN in H2-O2-HCN-Ar flames; studies by Song et al. (1980), on the effects of fuel-rich combustion on the conversion of the fixed nitrogen to N2; studies by White et al. (1983), on a rich-lean combustor for low-Btu and medium-Btu gaseous fuels; and research of the CRIEPI into fuel-NOx emission characteristics of low-calorific fuel, including NH3 through experiments using a small diffusion burner and analyses based on reaction kinetics (Hasegawa et al, 2001). It is widely

Fig. 1. Schematic diagram of typical IGCC system

#### **1.1 Background of IGCC development in the world**

The development of the gas turbine combustor for IGCC power generation received considerable attention in the 1970s. Brown (1982), summarized the overall progress of IGCC technology worldwide up until 1980. The history and application of gasification was also mentioned by Littlewood (1977). Concerning fixed-bed type gasification processes, Hobbs et al. (1993) extensively reviewed the technical and scientific aspects of the various systems. Other developments concerning the IGCC system and gas turbine combustor using oxygenblown gasified coal fuel include: The Cool Water Coal Gasification Project (Savelli & Touchton, 1985), the flagship demonstration plant of gasification and gasified fueled gas turbine generation; the Shell process (Bush et al., 1991) in Buggenum, the first commercial plant, which started test operation in 1994 and commercial operation in 1998; the Wabash River Coal Gasification Repowering Plant (Roll, 1995) in the United States, in operation since 1995; the Texaco process at the Tampa power station (Jenkins, 1995), in commercial operation since 1996; and an integrated coal gasification fuel cell combined cycle pilot plant, consisting of a gasifier, fuel cell generating unit and gas turbine, in test operation since 2002 by Electric Power Development Co. Ltd. in Japan. Every plant adopted the oxygen-blown gasification method. With regard to fossil-based gasification technology as described above, commercially-based power plants have been developed, and new development challenges toward global carbon capture storage (Isles, 2007; Beer, 2007) are being addressed.

Meanwhile, from 1986 to 1996, the Japanese government and electric power companies undertook an experimental research project for the air-blown gasification combined cycle system using a 200-ton-daily pilot plant. Recently, the government and electric power companies have also been promoting a demonstration IGCC project with a capacity of 1700 tons per day (Nagano, 2009). For the future commercializing stage, the transmission-end thermal efficiency of air-blown IGCC, adopting the 1773 K (1500°C)-class (average combustor exhaust gas temperature at about 1773 K) gas turbine, is expected to exceed 48%(on HHV basis), while the thermal efficiency of the demonstration plant using a 1473 K (1200°C)-class gas turbine is only 40.5%. IGCC technologies would improve thermal efficiency by five points or higher compared to the latest pulverized coal-firing, steam power generation. The Central Research Institute of Electric Power Industry (CRIEPI), developed an air-blown two-stage entrained-flow coal gasifier (Kurimura et al., 1995), a hot/dry synthetic gas cleanup system (Nakayama et al., 1990), and 150MW, 1773K(1500°C) class gas turbine combustor technologies for low-Btu fuel (Hasegawa et al., 1998a). In order

Air

The development of the gas turbine combustor for IGCC power generation received considerable attention in the 1970s. Brown (1982), summarized the overall progress of IGCC technology worldwide up until 1980. The history and application of gasification was also mentioned by Littlewood (1977). Concerning fixed-bed type gasification processes, Hobbs et al. (1993) extensively reviewed the technical and scientific aspects of the various systems. Other developments concerning the IGCC system and gas turbine combustor using oxygenblown gasified coal fuel include: The Cool Water Coal Gasification Project (Savelli & Touchton, 1985), the flagship demonstration plant of gasification and gasified fueled gas turbine generation; the Shell process (Bush et al., 1991) in Buggenum, the first commercial plant, which started test operation in 1994 and commercial operation in 1998; the Wabash River Coal Gasification Repowering Plant (Roll, 1995) in the United States, in operation since 1995; the Texaco process at the Tampa power station (Jenkins, 1995), in commercial operation since 1996; and an integrated coal gasification fuel cell combined cycle pilot plant, consisting of a gasifier, fuel cell generating unit and gas turbine, in test operation since 2002 by Electric Power Development Co. Ltd. in Japan. Every plant adopted the oxygen-blown gasification method. With regard to fossil-based gasification technology as described above, commercially-based power plants have been developed, and new development challenges

toward global carbon capture storage (Isles, 2007; Beer, 2007) are being addressed.

Meanwhile, from 1986 to 1996, the Japanese government and electric power companies undertook an experimental research project for the air-blown gasification combined cycle system using a 200-ton-daily pilot plant. Recently, the government and electric power companies have also been promoting a demonstration IGCC project with a capacity of 1700 tons per day (Nagano, 2009). For the future commercializing stage, the transmission-end thermal efficiency of air-blown IGCC, adopting the 1773 K (1500°C)-class (average combustor exhaust gas temperature at about 1773 K) gas turbine, is expected to exceed 48%(on HHV basis), while the thermal efficiency of the demonstration plant using a 1473 K (1200°C)-class gas turbine is only 40.5%. IGCC technologies would improve thermal efficiency by five points or higher compared to the latest pulverized coal-firing, steam power generation. The Central Research Institute of Electric Power Industry (CRIEPI), developed an air-blown two-stage entrained-flow coal gasifier (Kurimura et al., 1995), a hot/dry synthetic gas cleanup system (Nakayama et al., 1990), and 150MW, 1773K(1500°C) class gas turbine combustor technologies for low-Btu fuel (Hasegawa et al., 1998a). In order

Compressor Cooling water

Gas turbine

Generator

Heat recovery steam generator

Stack

Transformer

Steam turbine

Generator

Coal

Pulverizer

Slag hopper

**1.1 Background of IGCC development in the world** 

Fig. 1. Schematic diagram of typical IGCC system

Gasifier Heat

> Gasification agent Char Char recovery equipment

exchanger (Desulfurizing/ Char collecting)

**COAL GASIFIER HOT GAS CLEANUP GAS TURBINE**

to accept the various IGCC systems, 1773K-class gas turbine combustors of medium-Btu fuels by wet-type or hot/dry-type synthetic gas cleanup methods have undergone study (Hasegawa et al., 2003, 2007).

The energy resources and geographical conditions of each country, along with the diversification of fuels used for the electric power industry (such as biomass, poor quality coal and residual oil), are most significant issues for IGCC gas turbine development, as has been previously described: The development of biomass-fueled gasification received considerable attention in the United States and northern Europe in the early 1980s (Kelleher, 1985), and the prospects for commercialization technology (Consonni, 1997) appear considerably improved at present. Paisley and Anson (1997) performed a comprehensive economical evaluation of the Battele biomass gasification process, which utilizes a hot-gas conditioning catalyst for dry synthetic gas cleanup. In northern Europe, fixed-bed gasification heating plants built in the 1980s had been in commercial operation; the available technical and economical operation data convinced small district heating companies that biomass or peat-fueled gasification heating plants in the size class of 5 MW were the most profitable (Haavisto, 1996). However, during the period of stable global economy and oil prices, non-fossil-fueled gasification received little interest. Then, in the early 2000s when the Third Conference of Parties to the United Nations Framework Convention on Climate Change (COP3) invoked mandatory carbon dioxide emissions reductions on countries, biomass-fueled gasification technology began to receive considerable attention as one alternative. With the exception of Japanese national research and development project, almost all of the systems using the oxygen-blown gasification are in their final stages for commencing commercial operations overseas.

#### **1.2 Progress in gas turbine combustion technologies for IGCCs**

The plant thermal efficiency has been improved by enhancing the turbine inlet temperature, or combustor exhaust temperature. The thermal-NOx emissions from the gas turbines increase, however, along with a rise in exhaust temperature. In addition, gasified fuel containing NH3 emits fuel-NOx when hot/dry gas cleanup equipment is employed. It is therefore viewed as necessary to adopt a suitable combustion technology for each IGCC in the development of a gas turbine for each gasification method.

Dixon-Lewis and Williams (1969), expounded on the oxidation characteristics of hydrogen and carbon monoxide in 1969. The body of research into the basic combustion characteristics of gasified fuel includes studies on the flammability limits of mixed gas, consisting of CH4 or H2 diluted with N2, Ar or He (Ishizuka & Tsuji, 1980); a review of the flammability and explosion limits of H2 and H2/CO fuels (Cohen, 1992); the impact of N2 on burning velocity (Morgan & Kane, 1952); the effect of N2 and CO2 on flammability limits (Coward & Jones, 1952; Ishibasi et al, 1978); and the combustion characteristics of low calorific fuel (Folsom, 1980; Drake, 1984); studies by Merryman et al. (1997), on NOx formation in CO flame; studies by Miller et al. (1984), on the conversion characteristics of HCN in H2-O2-HCN-Ar flames; studies by Song et al. (1980), on the effects of fuel-rich combustion on the conversion of the fixed nitrogen to N2; studies by White et al. (1983), on a rich-lean combustor for low-Btu and medium-Btu gaseous fuels; and research of the CRIEPI into fuel-NOx emission characteristics of low-calorific fuel, including NH3 through experiments using a small diffusion burner and analyses based on reaction kinetics (Hasegawa et al, 2001). It is widely

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 243

BFG:Blast furnace gas, COG:Coke-ovengas, RDF:Refusederived fuel, Waste:Municipal solid waste,

Fig. 2. Relationship between equivalence ratio and adiabatic flame temperature for gasified

Equivalence ratio

On the other hand, in the case of oxygen-blown gasified fuels, flame temperature is so high that thermal-NOx emissions must be reduced. Therefore, in oxygen-blown IGCC, N2 produced by the air separation unit is used to recover power to increase the thermal efficiency of the plant, and to reduce NOx emissions from the gas turbine combustor by reducing the flame temperature. Furthermore, when hot/dry synthetic gas cleanup is employed, ammonia contained in the gasified fuels is not removed, but converted into fuel-NOx in the combustor. It is therefore necessary to reduce the fuel-NOx emissions in each

Because fuel conditions vary depending on the gasification method, many subjects arose in the development of the gasified fueled combustor. Table 2 summarizes the main subjects of

(a):No description, (b):Drybase Table 1. Various gasified fuels

fuels and CH4.

case of air-blown or oxygen-blown gasifiers.

combustor development for each IGCC method.

accepted that two-stage combustion, as typified by rich-lean combustion, is effective in reducing fuel-NOx emissions (Martin & Dederick, 1976; Yamagishi et al, 1974).

On the other hand, with respect to the combustion emission characteristics of oxygen-blown medium calorific fuel, Pillsbury et al. (1976) and Clark et al. (1982) investigated low-NOx combustion technologies using model combustors. In the 1970s, Battista and Farrell (1979) and Beebe et al. (1982) attempted one of the earliest tests using medium-Btu fuel in a gas turbine combustor. Concerning research into low-NOx combustion technology using oxygen-blown medium calorific fuel, other studies include: Hasegawa et al. (1997), investigation of NOx reduction technology using a small burner; and studies by Döbbeling et al. (1994), on low NOx combustion technology (which quickly mixed fuel with air using the double cone burner from Alstom Power, called an EV burner); Cook et al. (1994), on effective methods for returning nitrogen to the cycle, where nitrogen is injected from the head end of the combustor for NOx control; and Zanello and Tasselli (1996), on the effects of steam content in medium-Btu gaseous fuel on combustion characteristics. In almost all systems, surplus nitrogen was produced from the oxygen production unit and premixed with a gasified medium-Btu fuel (Becker & Shetter, 1992), for recovering power used in oxygen production and suppressing NOx emissions. Since the power to premix the surplus nitrogen with the medium-Btu fuel is great, Hasegawa et al. studied low-NOx combustion technologies using surplus nitrogen injected from the burner (Hasegawa et al, 1998b) and with the lean combustion of instantaneous mixing (Hasegawa et al, 2003). Furthermore, Hasegawa and Tamaru(2007) developed a low-NOx combustion technology for reducing both fuel-NOx and thermal-NOx emissions, in the case of employing hot/dry synthetic gas cleanup with an oxygen-blown IGCC.

#### **1.3 Subjects of gas turbine combustors for IGCCs**

The typical compositions of gasified fuels produced in air-blown or oxygen-blown gasifiers, and in blast furnaces, are shown in Tables 1. Each type of gaseous mixture fuel consists of CO and H2 as the main combustible components, and small percentages of CH4. Fuel calorific values vary widely (2–13 MJ/m3), from about 1/20 to 1/3 those of natural gas, depending upon the raw materials of feedstock, the gasification agent and the gasifier type. Figure 2 shows the theoretical adiabatic flame temperature of fuels which were: (1) gasified fuels with fuel calorific values (HHV) of 12.7, 10.5, 8.4, 6.3, 4.2 MJ/m3; and (2) fuels in which methane is the main component of natural gas. Flame temperatures were calculated using a CO and H2 mixture fuel (CO/H2 molar ratio of 2.33:1), which contained no CH4 under any conditions, and the fuel calorific value was adjusted with nitrogen. In the case of gasified fuel, as the fuel calorific value increased, the theoretical adiabatic flame temperature also increased. Fuel calorific values of 4.2 MJ/m3 and 12.7 MJ/m3 produced maximum flame temperatures of 2050 K and 2530 K, respectively. At fuel calorific values of 8.4 MJ/m3 or higher, the maximum flame temperature of the gasified fuel exceeded that of methane, while the fuel calorific value was as low as one-fifth of methane. Furthermore, each quantity of CO and H2 constituent in the gasified fuels differed, chiefly according to the gasification methods of gasifying agents, raw materials of feedstock, and water-gas-shift reaction as an optional extra for carbon capture system. However, it could be said that the theoretical adiabatic flame temperature was only a little bit affected by the CO/H2 molar ratio in the case of each fuel shown in Tables 1. That is to say, in air-blown gasified fuels, fuel calorific values are so low that flame stabilization is a problem confronting development of the combustor.


BFG:Blast furnace gas, COG:Coke-ovengas, RDF:Refusederived fuel, Waste:Municipal solid waste, (a):No description, (b):Drybase

Table 1. Various gasified fuels

242 Advances in Gas Turbine Technology

accepted that two-stage combustion, as typified by rich-lean combustion, is effective in

On the other hand, with respect to the combustion emission characteristics of oxygen-blown medium calorific fuel, Pillsbury et al. (1976) and Clark et al. (1982) investigated low-NOx combustion technologies using model combustors. In the 1970s, Battista and Farrell (1979) and Beebe et al. (1982) attempted one of the earliest tests using medium-Btu fuel in a gas turbine combustor. Concerning research into low-NOx combustion technology using oxygen-blown medium calorific fuel, other studies include: Hasegawa et al. (1997), investigation of NOx reduction technology using a small burner; and studies by Döbbeling et al. (1994), on low NOx combustion technology (which quickly mixed fuel with air using the double cone burner from Alstom Power, called an EV burner); Cook et al. (1994), on effective methods for returning nitrogen to the cycle, where nitrogen is injected from the head end of the combustor for NOx control; and Zanello and Tasselli (1996), on the effects of steam content in medium-Btu gaseous fuel on combustion characteristics. In almost all systems, surplus nitrogen was produced from the oxygen production unit and premixed with a gasified medium-Btu fuel (Becker & Shetter, 1992), for recovering power used in oxygen production and suppressing NOx emissions. Since the power to premix the surplus nitrogen with the medium-Btu fuel is great, Hasegawa et al. studied low-NOx combustion technologies using surplus nitrogen injected from the burner (Hasegawa et al, 1998b) and with the lean combustion of instantaneous mixing (Hasegawa et al, 2003). Furthermore, Hasegawa and Tamaru(2007) developed a low-NOx combustion technology for reducing both fuel-NOx and thermal-NOx emissions, in the case of employing hot/dry synthetic gas

The typical compositions of gasified fuels produced in air-blown or oxygen-blown gasifiers, and in blast furnaces, are shown in Tables 1. Each type of gaseous mixture fuel consists of CO and H2 as the main combustible components, and small percentages of CH4. Fuel calorific values vary widely (2–13 MJ/m3), from about 1/20 to 1/3 those of natural gas, depending upon the raw materials of feedstock, the gasification agent and the gasifier type. Figure 2 shows the theoretical adiabatic flame temperature of fuels which were: (1) gasified fuels with fuel calorific values (HHV) of 12.7, 10.5, 8.4, 6.3, 4.2 MJ/m3; and (2) fuels in which methane is the main component of natural gas. Flame temperatures were calculated using a CO and H2 mixture fuel (CO/H2 molar ratio of 2.33:1), which contained no CH4 under any conditions, and the fuel calorific value was adjusted with nitrogen. In the case of gasified fuel, as the fuel calorific value increased, the theoretical adiabatic flame temperature also increased. Fuel calorific values of 4.2 MJ/m3 and 12.7 MJ/m3 produced maximum flame temperatures of 2050 K and 2530 K, respectively. At fuel calorific values of 8.4 MJ/m3 or higher, the maximum flame temperature of the gasified fuel exceeded that of methane, while the fuel calorific value was as low as one-fifth of methane. Furthermore, each quantity of CO and H2 constituent in the gasified fuels differed, chiefly according to the gasification methods of gasifying agents, raw materials of feedstock, and water-gas-shift reaction as an optional extra for carbon capture system. However, it could be said that the theoretical adiabatic flame temperature was only a little bit affected by the CO/H2 molar ratio in the case of each fuel shown in Tables 1. That is to say, in air-blown gasified fuels, fuel calorific values are so low that flame stabilization is a

reducing fuel-NOx emissions (Martin & Dederick, 1976; Yamagishi et al, 1974).

cleanup with an oxygen-blown IGCC.

**1.3 Subjects of gas turbine combustors for IGCCs** 

problem confronting development of the combustor.

Fig. 2. Relationship between equivalence ratio and adiabatic flame temperature for gasified fuels and CH4.

On the other hand, in the case of oxygen-blown gasified fuels, flame temperature is so high that thermal-NOx emissions must be reduced. Therefore, in oxygen-blown IGCC, N2 produced by the air separation unit is used to recover power to increase the thermal efficiency of the plant, and to reduce NOx emissions from the gas turbine combustor by reducing the flame temperature. Furthermore, when hot/dry synthetic gas cleanup is employed, ammonia contained in the gasified fuels is not removed, but converted into fuel-NOx in the combustor. It is therefore necessary to reduce the fuel-NOx emissions in each case of air-blown or oxygen-blown gasifiers.

Because fuel conditions vary depending on the gasification method, many subjects arose in the development of the gasified fueled combustor. Table 2 summarizes the main subjects of combustor development for each IGCC method.

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 245

through the transition piece, the exhaust gas from the combustor was introduced into the measuring section where gas components and temperatures were measured. An automatic gas analyzer analyzed the components of the combustion gases. After that, the gas temperature

**Liner Swirler**

Exhaust gases were sampled from the exit of the combustor through water-cooled stainless steel probes located on the centerline of a height-wise cross section of the measuring duct. The sample lines of exhaust gases were thermally insulated with heat tape to maintain the sampling system above the dew point of the exhaust gas. The exhaust gases were sampled from at an area averaged points in the tail duct exit face and continuously introduced into an emission console which measured CO, CO2, NO, NOx, O2, and hydrocarbons by the same methods as the test device for basic studies using the small diffusion burner. The medium-Btu simulated fuel were sampled from the fuel gas supply line at the inlet of combustor, and constituents of CO, H2, CH4, H2O, CO2 and N2 were determined by gas chromatography. Heating values of the simulated gaseous fuel were monitored by a calorimeter and calculated from analytical data of gas components obtained from gas

The temperatures of the combustor liner walls were measured by sheathed type-K thermocouples with a diameter of 1mm attached to the liner wall with a stainless foil welding. The temperature distributions of the combustor exit gas were measured with an

This chapter indicates the characteristics of the combustion technologies being applied to the gasified fuels classified into four types in Table 2. Based on the knowledge through experiments using a small diffusion burner and numerical analyses, prototype combustors

array of three pyrometers, each of which consisted of five type-R thermocouples.

**3. Gas turbine combustors for the gasified fuels** 

were constructed, tested and their performances were demonstrated.

**Igniter Transition piece**

> **Combustion gas**

**Measurement position of Exhaust temperature, composition, and pressure**

**Measurement duct**

was lowered through a quenching pot, using a water spray injection system.

**Fuel gas**

**Air**

**Measurement position of Air temperature and pressure**

**Nitrogen**

**Kerosene**

Fig. 4. Combustion test rig

**2.2 Measurement system** 

chromatography.


Table 2. Subjects for combustors of various gasified fuels

## **2. Test facilities and method for gasified fueled combustors**

This chapter indicates a typical example of a test facility and method for a single-can combustion test using simulated gasified fuels.

## **2.1 Test facilities**

The schematic diagram of the test facilities is shown in Figure 3. The raw fuel obtained by mixing CO2 and steam with gaseous propane was decomposed to CO and H2 inside the fuel-reforming device. A hydrogen separation membrane was used to adjust the CO/H2 molar ratio. N2 was added to adjust the fuel calorific value to the prescribed calorie, and then simulated gases derived from gasifiers were produced.

This facility had another nitrogen supply line, by which nitrogen was directly injected into the combustor. Air supplied to the combustor was provided by using a four-stage centrifugal compressor. Both fuel and air were supplied to the gas turbine combustor after being heated separately with a preheater to the prescribed temperature.

Fig. 3. Schematic diagram and specifications of test facility

The combustion test facility had two test rigs, each of which was capable of performing fullscale atmospheric pressure combustion tests of a single-can for a "several"-hundreds MWclass, multican-type combustor as well as half-scale high-pressure combustion tests, or fullscale high-pressure tests for around a 100MW-class, multican-type combustor. Figure 4 shows a cross-sectional view of the combustor test rig under pressurized conditions. After passing through the transition piece, the exhaust gas from the combustor was introduced into the measuring section where gas components and temperatures were measured. An automatic gas analyzer analyzed the components of the combustion gases. After that, the gas temperature was lowered through a quenching pot, using a water spray injection system.

Fig. 4. Combustion test rig

244 Advances in Gas Turbine Technology

This chapter indicates a typical example of a test facility and method for a single-can

The schematic diagram of the test facilities is shown in Figure 3. The raw fuel obtained by mixing CO2 and steam with gaseous propane was decomposed to CO and H2 inside the fuel-reforming device. A hydrogen separation membrane was used to adjust the CO/H2 molar ratio. N2 was added to adjust the fuel calorific value to the prescribed calorie, and

This facility had another nitrogen supply line, by which nitrogen was directly injected into the combustor. Air supplied to the combustor was provided by using a four-stage centrifugal compressor. Both fuel and air were supplied to the gas turbine combustor after

**H2 Separater Max flow rate: 6.0 kg/s**

The combustion test facility had two test rigs, each of which was capable of performing fullscale atmospheric pressure combustion tests of a single-can for a "several"-hundreds MWclass, multican-type combustor as well as half-scale high-pressure combustion tests, or fullscale high-pressure tests for around a 100MW-class, multican-type combustor. Figure 4 shows a cross-sectional view of the combustor test rig under pressurized conditions. After passing

**Compressor**

**Max pressure: 2.0 MPa Temperature: 373~693K**

**Heater Stack**

**High Pressure Combustion Test Rig**

**Atmospheric Pressure Combustion Test Rig**

Air ・Combustion stability of low-calorific fuel

O2 ・Surplus nitrogen supply

**2. Test facilities and method for gasified fueled combustors** 

being heated separately with a preheater to the prescribed temperature.

**Fuel Reformer**

**3**

**Heater**

Table 2. Subjects for combustors of various gasified fuels

then simulated gases derived from gasifiers were produced.

**Max flow rate: 2.0 kg/s Heating value: 2.5~11.0 MJ/m CO/H2 ratio: 1~3 Temperature: 373~773K**

Fig. 3. Schematic diagram and specifications of test facility

**C3H8 CO2 Steam**

> **N2 NH3**

combustion test using simulated gasified fuels.

・Reduction of thermal-NOx

Gasification agent

**2.1 Test facilities**

Synthetic gas cleanup Wet type Hot/Dry type

calorific fuel

NOx emissions

・Combustion stability of low-

・Reduction of fuel-NOx

・Surplus nitrogen supply ・Reduction of thermal- and fuel-

## **2.2 Measurement system**

Exhaust gases were sampled from the exit of the combustor through water-cooled stainless steel probes located on the centerline of a height-wise cross section of the measuring duct. The sample lines of exhaust gases were thermally insulated with heat tape to maintain the sampling system above the dew point of the exhaust gas. The exhaust gases were sampled from at an area averaged points in the tail duct exit face and continuously introduced into an emission console which measured CO, CO2, NO, NOx, O2, and hydrocarbons by the same methods as the test device for basic studies using the small diffusion burner. The medium-Btu simulated fuel were sampled from the fuel gas supply line at the inlet of combustor, and constituents of CO, H2, CH4, H2O, CO2 and N2 were determined by gas chromatography. Heating values of the simulated gaseous fuel were monitored by a calorimeter and calculated from analytical data of gas components obtained from gas chromatography.

The temperatures of the combustor liner walls were measured by sheathed type-K thermocouples with a diameter of 1mm attached to the liner wall with a stainless foil welding. The temperature distributions of the combustor exit gas were measured with an array of three pyrometers, each of which consisted of five type-R thermocouples.

## **3. Gas turbine combustors for the gasified fuels**

This chapter indicates the characteristics of the combustion technologies being applied to the gasified fuels classified into four types in Table 2. Based on the knowledge through experiments using a small diffusion burner and numerical analyses, prototype combustors were constructed, tested and their performances were demonstrated.

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 247

Figure 6 presents characteristics of the designed and tested 1773K-class combustor. Figure 7 illustrates the external view of the burner of the combustor. The main design concept of the combustor was to secure stable combustion of a low-calorific fuel in a wide range of turn-down operation, low NOx emission and enough cooling-air for the combustor liner. The combustor is designed for advanced rich-lean combustion which is effective in

In order to assure flame stability of low-calorific fuel, an auxiliary combustion chamber is installed at the entrance of the combustor. The ratio of the fuel allocated to the auxiliary combustion chamber is 15 percent of the total amount of fuel. The fuel and the combustion air are injected into the chamber through a sub-swirler with a swirling angle of 30 degree. By setting the stoichiometric condition in this chamber under rated load conditions, a stable flame can be maintained. The rest of the fuel is introduced into the main combustion zone

To restrict the production of fuel NOx that is attributable to NH3 contained in the fuel, a two-stage combustion method (rich-lean combustion method) is introduced. The tested combustor has a two chamber structure, which separates the primary combustion zone from the secondary combustion zone. In addition, the combustor has two main design

The equivalence ratio of the primary combustor is determined setting at 1.6 based on the combustion tests previously conducted using a small diffusion burner (Hasegawa et al.,

Figure 8 shows an outline of the experimental device of the small diffusion burner. The combustion apparatus consists of a cylinder-style combustion chamber with an inner diameter, 'D', of 90mm and a length of 1,000mm, and a primary air swirler and fuel injection nozzle. The combustion chamber is lined with heat insulating material and the casing is cooled with water. In order to simulate two-stage combustion, secondary air inlets at a distance from the edge of the fuel injection nozzles of 3×'D' are used. The diameter of the secondary air inlets at the entry to the combustion chamber is 13mm, and six inlets are positioned on the perimeter of one cross-section. The tested burner consists of a fuel injection nozzle and a primary air swirler. There are twelve injection inlets with a diameter

Burner Combustion Liner

decreasing fuel NOx emissions resulting from fuel bound nitrogen.

from the surrounding of the exit of the auxiliary combustion chamber.

characteristics for reducing fuel NOx as indicated below:

*3.1.1.2.1 Air to fuel ratio in primary combustion zone* 

Fig. 7. Tested combustor

**3.1.1.2 Fuel-NOx reduction** 

2001).

**3.1.1.1 Assurance of flame stabilization** 

### **3.1 Combustor for air-blown gasification system with hot/dry type synthetic gas cleanup**

## **3.1.1 Design concept of combustor**

Figure 5 shows the relation between the combustor exhaust temperature and the air distribution in the gas turbine combustor using low-calorific gasified fuel. To calculate air distribution, the overall amount of air is assumed to be 100 percent. The amount of air for combustion is first calculated at 1.2 times of a theoretical air (φ=0.83), 30 percent of the total air is considered as the cooling air for the combustor liner wall, and the remaining air is considered as diluting air. According to this figure, as the gas turbine temperature rises up to 1773K, the ratio of cooling and diluting air decrease significantly, and the flexibility of the combustor design is minimized. To summarize these characteristics, it can be said that the design concept of the gas tur-bine combustor utilizing low-calorific fuel should consider the following issues when the gas turbine temperature rises:


Fig. 5. Air distribution design of a gas turbine combustor that burns low-Btu gasified fuel

Fig. 6. Design concept of 1773K-class low-Btu fueled combustor

Figure 5 shows the relation between the combustor exhaust temperature and the air distribution in the gas turbine combustor using low-calorific gasified fuel. To calculate air distribution, the overall amount of air is assumed to be 100 percent. The amount of air for combustion is first calculated at 1.2 times of a theoretical air (φ=0.83), 30 percent of the total air is considered as the cooling air for the combustor liner wall, and the remaining air is considered as diluting air. According to this figure, as the gas turbine temperature rises up to 1773K, the ratio of cooling and diluting air decrease significantly, and the flexibility of the combustor design is minimized. To summarize these characteristics, it can be said that the design concept of the gas tur-bine combustor utilizing low-calorific fuel should consider the

**3.1 Combustor for air-blown gasification system with hot/dry type synthetic gas** 

Combustion stability; it is necessary to stabilize the flame of low-calorific fuel.

Cooling structure to cool the combustor wall efficiently with less amount of air.

Fig. 5. Air distribution design of a gas turbine combustor that burns low-Btu gasified fuel

・Cooling air is distributed intensively in the first half of fuel-rich

・Using duplicate structure in the transition piece, cooling air for inner transition piece is recycled for liner wall cooling.

**Reduction of Fuel-NOx**

lean combustion zone.

**Fuel-Lean Secondary Combustion zone**

・Residence time in the fuel-rich combustion zone is set 1.5 times of the previous-type combustor. ・Penetration of the secondary air is diluted to lower the oxidation of ammonia intermediate in the fuel-

Low NOx emission technology to restrain the production of fuel NOx.

**cleanup** 

**3.1.1 Design concept of combustor** 

following issues when the gas turbine temperature rises:

・Adoption of auxiliary combustor.

**Combustion Stability**

**Fuel**

**Air**

**High-Efficiency Cooling**

**Fuel-Rich Primary Combustion zone**

combustion zone.

Fig. 6. Design concept of 1773K-class low-Btu fueled combustor

Figure 6 presents characteristics of the designed and tested 1773K-class combustor. Figure 7 illustrates the external view of the burner of the combustor. The main design concept of the combustor was to secure stable combustion of a low-calorific fuel in a wide range of turn-down operation, low NOx emission and enough cooling-air for the combustor liner. The combustor is designed for advanced rich-lean combustion which is effective in decreasing fuel NOx emissions resulting from fuel bound nitrogen.

## **3.1.1.1 Assurance of flame stabilization**

In order to assure flame stability of low-calorific fuel, an auxiliary combustion chamber is installed at the entrance of the combustor. The ratio of the fuel allocated to the auxiliary combustion chamber is 15 percent of the total amount of fuel. The fuel and the combustion air are injected into the chamber through a sub-swirler with a swirling angle of 30 degree. By setting the stoichiometric condition in this chamber under rated load conditions, a stable flame can be maintained. The rest of the fuel is introduced into the main combustion zone from the surrounding of the exit of the auxiliary combustion chamber.

## **3.1.1.2 Fuel-NOx reduction**

To restrict the production of fuel NOx that is attributable to NH3 contained in the fuel, a two-stage combustion method (rich-lean combustion method) is introduced. The tested combustor has a two chamber structure, which separates the primary combustion zone from the secondary combustion zone. In addition, the combustor has two main design characteristics for reducing fuel NOx as indicated below:

## *3.1.1.2.1 Air to fuel ratio in primary combustion zone*

The equivalence ratio of the primary combustor is determined setting at 1.6 based on the combustion tests previously conducted using a small diffusion burner (Hasegawa et al., 2001).

Figure 8 shows an outline of the experimental device of the small diffusion burner. The combustion apparatus consists of a cylinder-style combustion chamber with an inner diameter, 'D', of 90mm and a length of 1,000mm, and a primary air swirler and fuel injection nozzle. The combustion chamber is lined with heat insulating material and the casing is cooled with water. In order to simulate two-stage combustion, secondary air inlets at a distance from the edge of the fuel injection nozzles of 3×'D' are used. The diameter of the secondary air inlets at the entry to the combustion chamber is 13mm, and six inlets are positioned on the perimeter of one cross-section. The tested burner consists of a fuel injection nozzle and a primary air swirler. There are twelve injection inlets with a diameter

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 249

HHV=4.4MJ/m NH3=1000ppm Tair=673K Tfuel=298K φex =0.44 Vfuel=32m/s

CH4 % <sup>3</sup> 2.6 1.0 0.5 0

0.0 0.5 1.0 1.5 2.0 2.5 3.0 φp

Fig. 9. Effect of methane content on conversion rate of ammonia in the fuel to NOx, defining

From the test results, it is known that the conversion rate of NH3 to NOx is affected by both the equivalence ratio in the primary combustion zone using the two-staged combustion method and CH4 concentration. When the fuel contains CH4, HCN produced in the primary combustion zone is easily converted to NOx in the secondary combustion zone along with the decomposition of NH3. Therefore, there is a particular equivalence ratio, which minimizes the NOx conversion rate. Since the low-calorific fuel derived from the IGCC subject to development contained approximately 1.0 percent of CH4, the equivalence ratio in the primary-combustion zone was set at 1.6. The fuel and the primary combustion-air are injected from the burner, which has 30 degree swirl angle and 15 degree introvert angle.

An innovative idea was applied for secondary air introduction. With the decomposition of fuel N, a large portion of the total fixed nitrogen produced in the primary combustion zone, including NO, HCN and NHi, is converted to NOx in the secondary combustion zone. The influence of secondary air mixing conditions on the NOx production was examined from the viewpoint of reaction kinetics with modular model where each combustion zone means a perfect stirred reactor, neither the effect of diffusion nor that of radiant heat transfer of the flame are taken into account. As a result, it was found that the slower mixing of the secondary air made the conversion rate of NH3 to NOx decline further (Hasegawa et al., 1998a). Based on this result, an exterior wall was installed at the secondary-air inlet section in the tested combustor to make an intermediate pressure zone of the dual structure. By providing this dual structure, the flow speed of the secondary air introduced to the combustor decreased to 70m/s, compared to 120m/s without an exterior wall, thus the

In order to compensate for the declined cooling air ratio associated with the higher temperature of the gas turbine, the tested combustor is equipped with a dual-structure transition piece so that the cooling air in the transition piece can be recycled to cool down the combustor liner wall. The cooling air that flowed into the transition piece from the exterior wall cools the interior wall with an impingement method, and moves to the

0

by the experiments using a small diffusion burner (Hasegawa et al., 2001)

20

40

C.R. %

*3.1.1.2.2 Introduction method of secondary air* 

secondary air mixing was weakened. *3.1.1.2.3 Cooling of combustor liner wall* 

combustor liner at the upper streamside.

60

80

100

of 1.5mm on the fuel injection nozzle with an injection angle, *θ*, of 90-degree. The primary air swirler has an inner diameter of 24.0mm, an outer diameter of 36.4mm, and twelve vanes with a swirl angle, *θa*, of 45-degree. Swirl number, *S*, which is calculated from the following equation, is 0.84.

$$S = \frac{2}{3} \times \frac{1 - B^{\circ}}{1 - B^{\circ}} \times \tan \theta\_s \tag{1}$$

Where *B* (boss ratio of swirl vane)=0.66.

D : inner diameter of cylinder-style combustion chamber, 90mm θ: injection angle of fuel nozzle, 90 degrees

Fig. 8. Combustion chamber and diffusion burner of basic experimental device

Figure 9 presents an example of the test results which indicates the influence of the equivalence ratio of the primary combustion zone to the conversion rate of NH3 to NOx, C.R., at the exit of the secondary combustion zone. It also indicates the influence of the CH4 concentration in the fuel.

 *C R* th 3 NOx NOx volume flow rate of exhaust . . NH volume flow rate of fuel (2)

To obtain the conversion rate of NH3 to NOx, the concentration of thermal-NOx, '[NOxth]', was first measured after stopping the supply of NH3, then the concentration of total NOx, '[NOx]', was measured while NH3 was supplied, and finally fuel-NOx was calculated by deducting the concentration of thermal-NOx from that of total NOx. In the tests investigating fuel-NOx emissions, 1000ppm of NH3 is contained in the low-Btu fuel which consists of CO, H2 (CO/H2 molar ratio of 2.33:1), and small amount of CH4. In the case of changing CH4 concentration, fuel calorific value was adjusted by N2 dilution.

From the test results, it is known that the conversion rate of NH3 to NOx is affected by both the equivalence ratio in the primary combustion zone using the two-staged combustion method and CH4 concentration. When the fuel contains CH4, HCN produced in the primary combustion zone is easily converted to NOx in the secondary combustion zone along with the decomposition of NH3. Therefore, there is a particular equivalence ratio, which minimizes the NOx conversion rate. Since the low-calorific fuel derived from the IGCC subject to development contained approximately 1.0 percent of CH4, the equivalence ratio in the primary-combustion zone was set at 1.6. The fuel and the primary combustion-air are injected from the burner, which has 30 degree swirl angle and 15 degree introvert angle.

#### *3.1.1.2.2 Introduction method of secondary air*

248 Advances in Gas Turbine Technology

of 1.5mm on the fuel injection nozzle with an injection angle, *θ*, of 90-degree. The primary air swirler has an inner diameter of 24.0mm, an outer diameter of 36.4mm, and twelve vanes with a swirl angle, *θa*, of 45-degree. Swirl number, *S*, which is calculated from the following

*B*

 3 2

*<sup>B</sup> <sup>S</sup>*

To Exhaust Duct

Sampling Gas to Analyzer

Secondary Air

Primary Air

Fig. 8. Combustion chamber and diffusion burner of basic experimental device

 *C R* th 3

changing CH4 concentration, fuel calorific value was adjusted by N2 dilution.

Figure 9 presents an example of the test results which indicates the influence of the equivalence ratio of the primary combustion zone to the conversion rate of NH3 to NOx, C.R., at the exit of the secondary combustion zone. It also indicates the influence of the CH4

To obtain the conversion rate of NH3 to NOx, the concentration of thermal-NOx, '[NOxth]', was first measured after stopping the supply of NH3, then the concentration of total NOx, '[NOx]', was measured while NH3 was supplied, and finally fuel-NOx was calculated by deducting the concentration of thermal-NOx from that of total NOx. In the tests investigating fuel-NOx emissions, 1000ppm of NH3 is contained in the low-Btu fuel which consists of CO, H2 (CO/H2 molar ratio of 2.33:1), and small amount of CH4. In the case of

NOx NOx volume flow rate of exhaust . . NH volume flow rate of fuel (2)

D Hot-Face Insulation

Fuel

D : inner diameter of cylinder-style combustion chamber, 90mm

2 1

*a*

φ90 φ36.4 θ

> Fuel Air

3 1 (1)

tan

equation, is 0.84.

Where *B* (boss ratio of swirl vane)=0.66.

Fuel Nozzle

θ: injection angle of fuel nozzle, 90 degrees

concentration in the fuel.

3D 5D 8D

10.5D

An innovative idea was applied for secondary air introduction. With the decomposition of fuel N, a large portion of the total fixed nitrogen produced in the primary combustion zone, including NO, HCN and NHi, is converted to NOx in the secondary combustion zone. The influence of secondary air mixing conditions on the NOx production was examined from the viewpoint of reaction kinetics with modular model where each combustion zone means a perfect stirred reactor, neither the effect of diffusion nor that of radiant heat transfer of the flame are taken into account. As a result, it was found that the slower mixing of the secondary air made the conversion rate of NH3 to NOx decline further (Hasegawa et al., 1998a). Based on this result, an exterior wall was installed at the secondary-air inlet section in the tested combustor to make an intermediate pressure zone of the dual structure. By providing this dual structure, the flow speed of the secondary air introduced to the combustor decreased to 70m/s, compared to 120m/s without an exterior wall, thus the secondary air mixing was weakened.

#### *3.1.1.2.3 Cooling of combustor liner wall*

In order to compensate for the declined cooling air ratio associated with the higher temperature of the gas turbine, the tested combustor is equipped with a dual-structure transition piece so that the cooling air in the transition piece can be recycled to cool down the combustor liner wall. The cooling air that flowed into the transition piece from the exterior wall cools the interior wall with an impingement method, and moves to the combustor liner at the upper streamside.

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 251

Figure 11 shows the temperature distribution of the combustor liner wall at the rated load condition. From this figure, it could be said that the overall liner wall temperature almost remained under 1123K (850°C), the allowable heat resistant temperature, while the wall temperature increased to an adequate level and a stable flame was maintained in both the

> 0 100 200 300 400 500 600 700 800 900 Axial distance mm

**Axial distance mm** 

**3.2 Combustor for oxygen-blown gasification system with wet type synthetic gas** 

In the case of oxygen-blown IGCC, which has an air-separation unit to produce oxygen as gasification agent, medium-Btu gasified fuels are produced compared with the case of the air-blown gasified low-Btu fuels. That is, the maximum flame temperature of medium-Btu fuel is higher than that of each low-Btu fuel or high-calorie gas such as natural gas. Thermal-NOx emissions are expected to increase in the case of medium-Btu fueled combustors. Furthermore, in the oxygen-blown IGCC system, large quantity of nitrogen is produced in the air separation unit. In almost all of the systems, a part of nitrogen is used to feed raw material such as coal into the gasifier and so on, gasified fuels are premixed with the rest of the nitrogen and injected into the combustor to increase electric power and to decrease thermal-NOx emissions from the gas turbine. It is necessary to return a large quantity of the surplus nitrogen (as much as the fuel flow rate) to the cycle from the standpoint of recovering power for oxygen production. So, we intend to inject the surplus nitrogen directly into higher temperature regions from the burner and to decrease thermal-NOx emissions produced from these regions effectively. Analyses confirmed that the thermal efficiency of the plant improved by approximately 0.3 percent absolutely by means of nitrogen direct injection into the combustor, compared with a case where nitrogen is

Figure 12 presents characteristics of the designed, medium-Btu fueled 1573K (1300°C)-class combustor based on the above considerations. The main design concept for the tested

Tex=1773K HHV=4.2MJ/m3

**3.1.2.2 Thermal characteristics of combustor liner wall** 

premixed with gasified fuel before injection into the combustor.

**3.2.2 Design concept of combustor** 

Liner wall temp. K

**Liner wall temperature K** 

Fig. 11. Combustor wall temperature distribution

**cleanup** 

**3.2.1 Subjects of combustor** 

auxiliary-combustion chamber and the primary combustion zone.

For the auxiliary combustor and the primary combustion zone in which temperatures are expected to be especially high, the layer-built cooling structure that combined impingement cooling and film cooling was employed. For the secondary combustion zone, the film cooling method was used.

In addition to the above design characteristics, the primary air inlet holes are removed in order to maintain the given fuel-rich conditions in the primary combustion zone. Also, the overall length of the combustor, including the auxiliary chamber, is 1317mm, and the inside diameter is 356mm.

#### **3.1.2 Test results**

Combustion tests are conducted on under atmospheric pressure conditions. Concerning the pressure influence on the performance of the combustor, a half scale combustor, which has been developed by halving in dimension, was tested under pressurized conditions. Supplied fuels into the combustor were adjusted as same components as that of air-blown entrained-flow gasified coal fuel shown in Table 1. The standard rated conditions in the combustion tests are summarized in Table 3. Combustion Intensity at the design point is 2.0×102 W/(m3•Pa).


Table 3. Rated test conditions

#### **3.1.2.1 Combustion emission characteristics**

Figure 10 shows the combustion emission characteristics, under the gas turbine operational conditions. When the gas turbine load was 25 percent or higher, which is the single fuel firing of gasified fuel, the conversion rate of NH3 to NOx was reduced as low as 40 percent (NOx emissions corrected at 16 percent O2 was 60ppm), while the combustion efficiency shows around 100 percent in each gas turbine load.

Fig. 10. Combustion emission characteristics

## **3.1.2.2 Thermal characteristics of combustor liner wall**

250 Advances in Gas Turbine Technology

For the auxiliary combustor and the primary combustion zone in which temperatures are expected to be especially high, the layer-built cooling structure that combined impingement cooling and film cooling was employed. For the secondary combustion zone, the film

In addition to the above design characteristics, the primary air inlet holes are removed in order to maintain the given fuel-rich conditions in the primary combustion zone. Also, the overall length of the combustor, including the auxiliary chamber, is 1317mm, and the inside

Combustion tests are conducted on under atmospheric pressure conditions. Concerning the pressure influence on the performance of the combustor, a half scale combustor, which has been developed by halving in dimension, was tested under pressurized conditions. Supplied fuels into the combustor were adjusted as same components as that of air-blown entrained-flow gasified coal fuel shown in Table 1. The standard rated conditions in the combustion tests are summarized in Table 3. Combustion Intensity at the design point is

Figure 10 shows the combustion emission characteristics, under the gas turbine operational conditions. When the gas turbine load was 25 percent or higher, which is the single fuel firing of gasified fuel, the conversion rate of NH3 to NOx was reduced as low as 40 percent (NOx emissions corrected at 16 percent O2 was 60ppm), while the combustion efficiency

> 0 20 40 60 80 100 Gas turbine load %

700K 633K 1773K 1.4MPa 0.62 2.0×102 W/(m3•Pa)

99.5

99.6

99.7

η

**Combustion efficiency %** 

%

99.8

99.9

100

*T*air *T*fuel *T*ex *P*  ex Combustion Intensity

cooling method was used.

diameter is 356mm.

**3.1.2 Test results** 

2.0×102 W/(m3•Pa).

Table 3. Rated test conditions

**3.1.2.1 Combustion emission characteristics** 

shows around 100 percent in each gas turbine load.

0

20

HHV=4.2MJ/m CH4=1.0% NH3=1000ppm 3

40

C.R. %

Fig. 10. Combustion emission characteristics

60

80

100

Figure 11 shows the temperature distribution of the combustor liner wall at the rated load condition. From this figure, it could be said that the overall liner wall temperature almost remained under 1123K (850°C), the allowable heat resistant temperature, while the wall temperature increased to an adequate level and a stable flame was maintained in both the auxiliary-combustion chamber and the primary combustion zone.

Fig. 11. Combustor wall temperature distribution

## **3.2 Combustor for oxygen-blown gasification system with wet type synthetic gas cleanup**

## **3.2.1 Subjects of combustor**

In the case of oxygen-blown IGCC, which has an air-separation unit to produce oxygen as gasification agent, medium-Btu gasified fuels are produced compared with the case of the air-blown gasified low-Btu fuels. That is, the maximum flame temperature of medium-Btu fuel is higher than that of each low-Btu fuel or high-calorie gas such as natural gas. Thermal-NOx emissions are expected to increase in the case of medium-Btu fueled combustors.

Furthermore, in the oxygen-blown IGCC system, large quantity of nitrogen is produced in the air separation unit. In almost all of the systems, a part of nitrogen is used to feed raw material such as coal into the gasifier and so on, gasified fuels are premixed with the rest of the nitrogen and injected into the combustor to increase electric power and to decrease thermal-NOx emissions from the gas turbine. It is necessary to return a large quantity of the surplus nitrogen (as much as the fuel flow rate) to the cycle from the standpoint of recovering power for oxygen production. So, we intend to inject the surplus nitrogen directly into higher temperature regions from the burner and to decrease thermal-NOx emissions produced from these regions effectively. Analyses confirmed that the thermal efficiency of the plant improved by approximately 0.3 percent absolutely by means of nitrogen direct injection into the combustor, compared with a case where nitrogen is premixed with gasified fuel before injection into the combustor.

## **3.2.2 Design concept of combustor**

Figure 12 presents characteristics of the designed, medium-Btu fueled 1573K (1300°C)-class combustor based on the above considerations. The main design concept for the tested

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 253

 N2 supply not supplied blended with Air blended with Fuel Tex=1300℃ N2/Fuel=2.1kg/kg Tair=370℃ Tfuel=360℃ Tex=1773K HHV=12.7MJ/m N2/Fuel=2.1kg/kg Tair=503K Tfuel=503K

3

**<sup>0</sup> <sup>30</sup> <sup>60</sup> <sup>90</sup> <sup>0</sup>**

**Radial Distance mm**

**50**

**100**

**1473**

**150**

**Axial Distance from Burner mm**

**Fuel: Air : N2 :**

Fig. 14. Effect of nitrogen injection on combustion gas temperature distribution using a

(1)No injection of nitrogen (2)Nitrogen injected

Based on these basic test results, we arranged the nitrogen injection intakes in the burner and adopted the lean primary combustion, as shown in figure 12. The nitrogen injected directly into a combustor has the effect of decreasing power to compress nitrogen, compared with the case where the nitrogen was blended with fuel or air evenly. And it is possible to control the mixing of fuel, air and nitrogen positively by way of nitrogen being injected separately into the combustor. The nitrogen direct injection from the burner dilutes the

**200**

**250**

**300**

012345 φp

φp

Fig. 13. Effect of nitrogen injection on thermal-NOx emission characteristics in two-stage

Figure 14 shows the combustion gas temperature distribution in the both cases of no nitrogen injection and of nitrogen injection of 1.0kg/kg N2/Fuel from the burner under atmospheric pressure condition, using a model combustor. In tests, the combustor outlet gas temperature is set at 1373K. From figure 14, we know that nitrogen injection from the burner has a great influence over decreasing hot regions by around 200K in this test conditions. So, in this way of nitrogen injection, thermal-NOx production was restrained one fifth that of

0

**<sup>0</sup> <sup>30</sup> <sup>60</sup> <sup>90</sup> <sup>0</sup>**

**Radial Distance mm**

**50**

**100**

**150**

**Axial Distance from Burner mm**

**200**

**250**

**1273**

**1073**

**1473 1673 1873K**

**1673**

**300**

20

40

60

NOx ppm

Thermal NOx(16%O2) ppm

combustion, using a small diffusion burner

the case no nitrogen injection.

model combustor

th

80

100

120

combustor was to secure a low-NOx and stable combustion of medium-Btu fuel with nitrogen injection in a wide range of turn-down operations. The overall length of the combustion liner is 650mm and the inside diameter is 230mm.

Fig. 12. Design concepts of medium-Btu fueled combustor for wet-type synthetic gas cleanup

According to the combustor cooling, a convection method was employed in the transition piece, and moves to the combustor liner on the upstream side. For the primary combustion zone where temperatures are expected to be especially high, the dual-cooling structure was employed, in which the cooling air was impinged from the air flow guide sleeve to the combustion liner and used as film cooling air for the combustion liner. For the secondary combustion zone, the film-cooling method was used.

To restrict thermal-NOx production originating from nitrogen fixation and CO emissions, the burner was designed with nitrogen injection function, based on combustion tests previously conducted using a small diffusion burner (Hasegawa et al., 2001) and a small model combustor (Hasegawa et al., 2003).

Figure 13 presents an example of the test results using the small diffusion burner shown in figure 8, which indicate the influence of the primary equivalence ratio on NOx emission characteristics in two-staged combustion for comparing three cases: 1) a fuel calorific value (*HHV*) of 12.7MJ/m3, without nitrogen injection; 2) a fuel calorific value of 12.7MJ/m3, where nitrogen is blended with the combustion air from the burner; 3) a fuel blended with nitrogen of the same quantity as case 2), or low-Btu fuel of 5.1MJ/m3. From figure 13, we know that nitrogen blended with fuel or air injected from the burner has a great influence over decreasing NOx emissions from nitrogen fixation. On the other hand, not shown in here, in the case where nitrogen blended with air was injected into the combustor, CO emissions decreased as low as medium-Btu gasified fuel not blended with nitrogen, while CO emissions significantly increased when fuel was blended with nitrogen. That is, in the medium-Btu fuel combustion with nitrogen injection, all of the surplus nitrogen should be injected into the primary combustion zone to reduce the thermal-NOx emissions and should not be blended with fuel, or the primary zone should be fuel lean condition for a low NOx and stable combustion in a wide range of turn-down operations.

combustor was to secure a low-NOx and stable combustion of medium-Btu fuel with nitrogen injection in a wide range of turn-down operations. The overall length of the

・Adoption of combination of film cooling with impingement

・Using duplicate structure in the transition piece, cooling air for inner transition piece is recycled for liner wall cooling.

According to the combustor cooling, a convection method was employed in the transition piece, and moves to the combustor liner on the upstream side. For the primary combustion zone where temperatures are expected to be especially high, the dual-cooling structure was employed, in which the cooling air was impinged from the air flow guide sleeve to the combustion liner and used as film cooling air for the combustion liner. For the secondary

To restrict thermal-NOx production originating from nitrogen fixation and CO emissions, the burner was designed with nitrogen injection function, based on combustion tests previously conducted using a small diffusion burner (Hasegawa et al., 2001) and a small

Figure 13 presents an example of the test results using the small diffusion burner shown in figure 8, which indicate the influence of the primary equivalence ratio on NOx emission characteristics in two-staged combustion for comparing three cases: 1) a fuel calorific value (*HHV*) of 12.7MJ/m3, without nitrogen injection; 2) a fuel calorific value of 12.7MJ/m3, where nitrogen is blended with the combustion air from the burner; 3) a fuel blended with nitrogen of the same quantity as case 2), or low-Btu fuel of 5.1MJ/m3. From figure 13, we know that nitrogen blended with fuel or air injected from the burner has a great influence over decreasing NOx emissions from nitrogen fixation. On the other hand, not shown in here, in the case where nitrogen blended with air was injected into the combustor, CO emissions decreased as low as medium-Btu gasified fuel not blended with nitrogen, while CO emissions significantly increased when fuel was blended with nitrogen. That is, in the medium-Btu fuel combustion with nitrogen injection, all of the surplus nitrogen should be injected into the primary combustion zone to reduce the thermal-NOx emissions and should not be blended with fuel, or the primary zone should be fuel lean condition for a low NOx

Combustion Stability

near the burner.

・Using nitrogen-bypass function in lowload condition, stable flame is maintained

combustion liner is 650mm and the inside diameter is 230mm.

High-Efficiency Cooling

for high temperature region

Fig. 12. Design concepts of medium-Btu fueled combustor for wet-type synthetic gas

**Fuel N2**

**Air**

:Air :Fuel :N2

cleanup

Reduction of thermal-NOx

・Adoption of nitrogen direct injection burner ・Quick-diluting with air in the primary zone

combustion zone, the film-cooling method was used.

and stable combustion in a wide range of turn-down operations.

model combustor (Hasegawa et al., 2003).

Fig. 13. Effect of nitrogen injection on thermal-NOx emission characteristics in two-stage combustion, using a small diffusion burner

Figure 14 shows the combustion gas temperature distribution in the both cases of no nitrogen injection and of nitrogen injection of 1.0kg/kg N2/Fuel from the burner under atmospheric pressure condition, using a model combustor. In tests, the combustor outlet gas temperature is set at 1373K. From figure 14, we know that nitrogen injection from the burner has a great influence over decreasing hot regions by around 200K in this test conditions. So, in this way of nitrogen injection, thermal-NOx production was restrained one fifth that of the case no nitrogen injection.

Fig. 14. Effect of nitrogen injection on combustion gas temperature distribution using a model combustor

Based on these basic test results, we arranged the nitrogen injection intakes in the burner and adopted the lean primary combustion, as shown in figure 12. The nitrogen injected directly into a combustor has the effect of decreasing power to compress nitrogen, compared with the case where the nitrogen was blended with fuel or air evenly. And it is possible to control the mixing of fuel, air and nitrogen positively by way of nitrogen being injected separately into the combustor. The nitrogen direct injection from the burner dilutes the

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 255

efficiency shows around 100 percent in the case where the gas turbine load was 25 percent or higher, by bypassing nitrogen to premix with the combustion air at low load conditions.

> By-passing N2 in low-load condition, stable flame is maintained.

20 40 60 80 100

**Gas Turbine Load %**

**3.3 Combustor for oxygen-blown gasification system with hot/dry type synthetic gas** 

In order to improve the thermal efficiency of the oxygen-blown IGCC, it is necessary to adopt the hot/dry synthetic gas cleanup. In this case, ammonia contained in the gasified fuels could not be removed and fuel-NOx is emitted from the gas turbine. It is necessary to develop to low NOx combustion technologies that reduce fuel-NOx emissions originating

From the characteristic of medium-Btu, gasified fuel as mentioned above, it could be said that the design of a gas turbine combustor with nitrogen supply, should consider the

1. Low NOx-emission technology: Thermal-NOx production from nitrogen fixation using nitrogen injection, and fuel-NOx emissions originating from ammonia using a two-

2. Higher thermal efficiency: Nitrogen injection must be tailored so as to decrease the power to compress nitrogen, which is returned into the gas turbine in order to recover a

Figure 16 presents the configuration and its function of a designed, medium-Btu fueled 1773K (1500°C)-class combustor based on the above considerations. The main design concepts for the tested combustor were to secure stable combustion of medium-Btu fuel with nitrogen injection in a wide range of turn-down operations, and low NOx combustion for reducing fuel-NOx and thermal-NOx emissions. In order to secure stable combustion, we installed an auxiliary combustion chamber at the entrance of the combustor. To reduce thermal-NOx emissions, the nitrogen injection nozzles was set up in the main-swirler, which is installed at exit of the auxiliary combustion chamber. The overall length of the

following issues for an oxygen-blown IGCC with the hot/dry synthetic gas cleanup:

Fig. 15. Effect of the gas turbine load on combustion emission characteristics

from ammonia in the fuel at the same time as reducing thermal-NOx ones.

stage combustion must be simultaneously restrained.

part of the power used for the air-separation unit.

combustion liner is 445mm and the inside diameter is 175mm.

99.5

99.6

3

HHV=10.1MJ/m Qair=3.5kg/s N2/Fuel=0.3kg/kg 99.7

99.8

**Combustion Efficiency %**

**Combustion efficiency %** 

**%**

η

99.9

100

0

5

10

15

**NOx(16%O ) ppm**

**cleanup** 

**3.3.1 Subjects of combustor** 

**3.3.2 Design concept of combustor** 

**2**

20

25

flame of medium-Btu fuel. Furthermore we intended to quench the flame as soon as possible, both by sticking the combustion air injection tubes out of the liner dome and by arranging the secondary combustion air holes on the upstream side of the combustion liner. Design of the combustor was intended for the medium-Btu fuel, the nitrogen injection function was combined with the lean combustion technique for a low NOx combustion. By setting the primary combustion zone to fuel lean state under the rated load condition, the NOx emissions are expected to decrease, and by bypassing nitrogen to premix with the combustion air under partial load conditions, a stable flame can be maintained in a wide range of turn-down operations.

#### **3.2.3 Test results**

Table 4 and 5 show the typical properties of the supplied fuel and the standard test conditions, respectively. Higher heating value of the supplied fuel, HHV, was set at 10.1MJ/m3, CH4 was contained higher concentration of 6.8 percent. A part of surplus nitrogen produced from the air-separation unit was used to feed coal or char into the gasifier and the flow rate of the rest was about 0.9 times the fuel flow in the actual process. Since the density of the supplied fuel is higher than that of the commercial gasified fuel and temperature of supplied nitrogen is lower in the case of the test conditions than in the actual operations, the combustor performance is investigated in the case of 0.3kg/kg N2/Fuel ratio, in which firing temperature of the burner outlet corresponds to the case of actual operations. The rated temperature of combustor-outlet gas, *T*ex, is around 1700K.


Table 4. Typical conditions of supplied fuel


Table 5. Rated test conditions

Figure 15 shows the relationship between the gas turbine load and the combustion emission characteristics, under the condition where the pressure in the combustor is set to 1.0MPa of slightly less than that of the practical operation at the equivalent, rated load. When the gas turbine load was 25 percent or higher, which is the single fuel firing of gasified fuel, the NOx emission was reduced as low as 11ppm (corrected at 16 percent O2), while the NOx emission tends to increase slightly with the rise in the gas turbine load. Considering the effects of pressure, it could be said that NOx emission was surmised as low as 12ppm (corrected at 16 percent O2) at any gas turbine load. On the other hand, combustion efficiency shows around 100 percent in the case where the gas turbine load was 25 percent or higher, by bypassing nitrogen to premix with the combustion air at low load conditions.

Fig. 15. Effect of the gas turbine load on combustion emission characteristics

#### **3.3 Combustor for oxygen-blown gasification system with hot/dry type synthetic gas cleanup**

In order to improve the thermal efficiency of the oxygen-blown IGCC, it is necessary to adopt the hot/dry synthetic gas cleanup. In this case, ammonia contained in the gasified fuels could not be removed and fuel-NOx is emitted from the gas turbine. It is necessary to develop to low NOx combustion technologies that reduce fuel-NOx emissions originating from ammonia in the fuel at the same time as reducing thermal-NOx ones.

## **3.3.1 Subjects of combustor**

254 Advances in Gas Turbine Technology

flame of medium-Btu fuel. Furthermore we intended to quench the flame as soon as possible, both by sticking the combustion air injection tubes out of the liner dome and by arranging the secondary combustion air holes on the upstream side of the combustion liner. Design of the combustor was intended for the medium-Btu fuel, the nitrogen injection function was combined with the lean combustion technique for a low NOx combustion. By setting the primary combustion zone to fuel lean state under the rated load condition, the NOx emissions are expected to decrease, and by bypassing nitrogen to premix with the combustion air under partial load conditions, a stable flame can be maintained in a wide

Table 4 and 5 show the typical properties of the supplied fuel and the standard test conditions, respectively. Higher heating value of the supplied fuel, HHV, was set at 10.1MJ/m3, CH4 was contained higher concentration of 6.8 percent. A part of surplus nitrogen produced from the air-separation unit was used to feed coal or char into the gasifier and the flow rate of the rest was about 0.9 times the fuel flow in the actual process. Since the density of the supplied fuel is higher than that of the commercial gasified fuel and temperature of supplied nitrogen is lower in the case of the test conditions than in the actual operations, the combustor performance is investigated in the case of 0.3kg/kg N2/Fuel ratio, in which firing temperature of the burner outlet corresponds to the case of actual operations.

HHV 10.1 MJ/m3

Figure 15 shows the relationship between the gas turbine load and the combustion emission characteristics, under the condition where the pressure in the combustor is set to 1.0MPa of slightly less than that of the practical operation at the equivalent, rated load. When the gas turbine load was 25 percent or higher, which is the single fuel firing of gasified fuel, the NOx emission was reduced as low as 11ppm (corrected at 16 percent O2), while the NOx emission tends to increase slightly with the rise in the gas turbine load. Considering the effects of pressure, it could be said that NOx emission was surmised as low as 12ppm (corrected at 16 percent O2) at any gas turbine load. On the other hand, combustion

30.4 vol% 27.5 vol% 6.8 vol% 35.3 vol%

603K 583K 333K 1700K 0.3kg/kg 1.4MPa 2.2×102 W/(m3•Pa)

The rated temperature of combustor-outlet gas, *T*ex, is around 1700K.

Composition CO H2 CH4 CO2

*T*air *T*fuel *T*N2 *T*ex N2/Fuel ratio *P*  Combustion Intensity

Table 4. Typical conditions of supplied fuel

Table 5. Rated test conditions

range of turn-down operations.

**3.2.3 Test results** 

From the characteristic of medium-Btu, gasified fuel as mentioned above, it could be said that the design of a gas turbine combustor with nitrogen supply, should consider the following issues for an oxygen-blown IGCC with the hot/dry synthetic gas cleanup:


## **3.3.2 Design concept of combustor**

Figure 16 presents the configuration and its function of a designed, medium-Btu fueled 1773K (1500°C)-class combustor based on the above considerations. The main design concepts for the tested combustor were to secure stable combustion of medium-Btu fuel with nitrogen injection in a wide range of turn-down operations, and low NOx combustion for reducing fuel-NOx and thermal-NOx emissions. In order to secure stable combustion, we installed an auxiliary combustion chamber at the entrance of the combustor. To reduce thermal-NOx emissions, the nitrogen injection nozzles was set up in the main-swirler, which is installed at exit of the auxiliary combustion chamber. The overall length of the combustion liner is 445mm and the inside diameter is 175mm.

appropriate

conditions with a lower

with nitrogen, while the optimum

p. Those optimum

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 257

With regard to fuel-NOx emissions on the other hand, figure 18 indicates the effects of nitrogen injection conditions on the conversion rate of NH3 in the fuel to NOx, C.R. in the same conditions with figure 13 except for fuel containing NH3. In the tests investigating fuel-NOx emissions, 1000ppm of NH3 is contained in the medium-Btu fuel. In the case of a fuel blended with nitrogen, fuel was diluted, or fuel calorific value decreased to 5.1MJ/m3 and NH3 concentration in the fuel decreased to 400ppm. From figure 18, whether with or without nitrogen supplied, the staged combustion method effectively decreased the fuel-

higher than 1.0, which is a stoichiometric condition, and shows the minimum value at the

with the primary combustion air injected from the burner, and C.R. showed a tendency to become a little higher than in the other two cases. Furthermore, under lean-lean combustion

C.R. becomes higher than in the case of nitrogen premixed with the primary combustion air.

012345 φp

φp

Fig. 18. Effect of nitrogen injection on the conversion rate of NH3 to NOx in two-stage

From the above, it was shown that the supply method of nitrogen premixed with medium-Btu fuel possibly decreases total emissions of thermal-NOx and fuel-NOx, but careful attention must be paid to the homogeneity of mixture of fuel and nitrogen, or thermal-NOx emissions will increase. In the case of nitrogen premixed with the primary combustion air, total NOx emissions grow slightly higher than the case of nitrogen premixed with fuel, and the power to compress nitrogen increases a little or the thermal efficiency of the plant decreases. That is, it is necessary to blend nitrogen with medium-Btu fuel evenly in the combustor, in which the lowest power to compress nitrogen is needed for nitrogen supply into the gas turbine, and not to collide the medium-Btu fuel with combustion air directly. Based on these basic experimental results, we arranged the nitrogen injection intakes between fuel and air intakes in the main-swirler surrounding the primary-flame from the auxiliary combustion chamber for low thermal-NOx emissions. Additionally the fuel, the combustion air, and the nitrogen from the burner are separately injected into the combustor

p become lower when the medium-Btu fuel was blended

p than 1.0, in the case of nitrogen premixed with medium-Btu fuel,

3

 N2 supply not supplied blended with Air blended with Fuel Tex=1300℃ N2/Fuel=2.1kg/kg Tair=370℃ Tfuel=360℃ Tex=1773K HHV=12.7MJ/m NH3=1000ppm N2/Fuel=2.1kg/kg Tair=503K Tfuel=503K

p was in a wide range in the case of nitrogen blended

p at 1.6 or

p, become

p is 1.3 or higher in the case of nitrogen premixed with fuel, and by setting

NOx emissions, or C.R. drastically decreased as the primary equivalence ratio,

higher in the case of nitrogen premixed with primary combustion air.

0

20

40

60

C.R. %

C.R. %

combustion, using a small diffusion burner

80

100

Fig. 16. Design concept of a medium-Btu fueled gas turbine combustor for hot/dry-type synthetic gas cleanup

Figure 17 illustrates the axial distribution of equivalence ratio at the rated load condition. In order to reduce the fuel-NOx emissions, we adopted the two-stage combustion, in which a fuel-rich combustion was carried out in the primary zone maintaining the equivalence ratio of 0.66 at exit of the combustor. And the designed combustor has following characteristics.

Fig. 17. Axial distribution of equivalence ratio at the rated load condition

#### **3.3.2.1 Assurance of flame stabilization**

The ratio of the fuel allocated to the auxiliary combustion chamber is 30 percent of the total amount of fuel. The fuel and air are injected into the chamber through a sub-swirler with a swirling angle of 30-degree. By setting the mean equivalence ratio in the auxiliary chamber at 2.4 under rated load conditions, a stable flame can be maintained in the fuel-rich combustion zone and reduction of NH3 to N2 could proceed in lower load conditions. The rest of the fuel is introduced into the main combustion zone from the surrounding of the exit of the auxiliary combustion chamber.

#### **3.3.2.2 Nitrogen injection**

From figure 13, we just noticed that nitrogen supply, which is blended with fuel or primary air, drastically decreases thermal-NOx emissions, and also NOx emissions decreases with rises in the primary equivalence ratio, p, in the case of using the two-stage combustion. That is, thermal-NOx emissions decrease significantly by setting a fuel-rich condition when

**Reduction of fuel NOx**

**By secondary-air,**

**unburned fuel is combusted.**

**secondary-air holes**

**Reduction of thermal NOx** ・Nitrogen direct injection lowers the flame temp.

**Reducing flame**

:**Air**

synthetic gas cleanup

:**Fuel** :**N2**

**Stable combustion** ・Auxiliary combustor

**decomposes NH3 to N2**

Fig. 16. Design concept of a medium-Btu fueled gas turbine combustor for hot/dry-type

Combustion liner

combustor Transition

piace

Figure 17 illustrates the axial distribution of equivalence ratio at the rated load condition. In order to reduce the fuel-NOx emissions, we adopted the two-stage combustion, in which a fuel-rich combustion was carried out in the primary zone maintaining the equivalence ratio of 0.66 at exit of the combustor. And the designed combustor has following characteristics.

> 0 100 200 300 400 500 600 700 800 Axicial distance mm

The ratio of the fuel allocated to the auxiliary combustion chamber is 30 percent of the total amount of fuel. The fuel and air are injected into the chamber through a sub-swirler with a swirling angle of 30-degree. By setting the mean equivalence ratio in the auxiliary chamber at 2.4 under rated load conditions, a stable flame can be maintained in the fuel-rich combustion zone and reduction of NH3 to N2 could proceed in lower load conditions. The rest of the fuel is introduced into the main combustion zone from the surrounding of the exit

From figure 13, we just noticed that nitrogen supply, which is blended with fuel or primary air, drastically decreases thermal-NOx emissions, and also NOx emissions decreases with

That is, thermal-NOx emissions decrease significantly by setting a fuel-rich condition when

p, in the case of using the two-stage combustion.

0

2

4

6

Equivalence ratio

**3.3.2.1 Assurance of flame stabilization** 

of the auxiliary combustion chamber.

rises in the primary equivalence ratio,

**3.3.2.2 Nitrogen injection** 

8

10

Auxiliary

Fig. 17. Axial distribution of equivalence ratio at the rated load condition

p is 1.3 or higher in the case of nitrogen premixed with fuel, and by setting p at 1.6 or higher in the case of nitrogen premixed with primary combustion air.

With regard to fuel-NOx emissions on the other hand, figure 18 indicates the effects of nitrogen injection conditions on the conversion rate of NH3 in the fuel to NOx, C.R. in the same conditions with figure 13 except for fuel containing NH3. In the tests investigating fuel-NOx emissions, 1000ppm of NH3 is contained in the medium-Btu fuel. In the case of a fuel blended with nitrogen, fuel was diluted, or fuel calorific value decreased to 5.1MJ/m3 and NH3 concentration in the fuel decreased to 400ppm. From figure 18, whether with or without nitrogen supplied, the staged combustion method effectively decreased the fuel-NOx emissions, or C.R. drastically decreased as the primary equivalence ratio, p, become higher than 1.0, which is a stoichiometric condition, and shows the minimum value at the appropriate p. Those optimum p become lower when the medium-Btu fuel was blended with nitrogen, while the optimum p was in a wide range in the case of nitrogen blended with the primary combustion air injected from the burner, and C.R. showed a tendency to become a little higher than in the other two cases. Furthermore, under lean-lean combustion conditions with a lower p than 1.0, in the case of nitrogen premixed with medium-Btu fuel, C.R. becomes higher than in the case of nitrogen premixed with the primary combustion air.

Fig. 18. Effect of nitrogen injection on the conversion rate of NH3 to NOx in two-stage combustion, using a small diffusion burner

From the above, it was shown that the supply method of nitrogen premixed with medium-Btu fuel possibly decreases total emissions of thermal-NOx and fuel-NOx, but careful attention must be paid to the homogeneity of mixture of fuel and nitrogen, or thermal-NOx emissions will increase. In the case of nitrogen premixed with the primary combustion air, total NOx emissions grow slightly higher than the case of nitrogen premixed with fuel, and the power to compress nitrogen increases a little or the thermal efficiency of the plant decreases. That is, it is necessary to blend nitrogen with medium-Btu fuel evenly in the combustor, in which the lowest power to compress nitrogen is needed for nitrogen supply into the gas turbine, and not to collide the medium-Btu fuel with combustion air directly. Based on these basic experimental results, we arranged the nitrogen injection intakes between fuel and air intakes in the main-swirler surrounding the primary-flame from the auxiliary combustion chamber for low thermal-NOx emissions. Additionally the fuel, the

combustion air, and the nitrogen from the burner are separately injected into the combustor

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 259

The nitrogen of NH3 in the fuel has weaker bonding power than N2. In the combustion process, NH3 reacted with the OH, O, and H radicals and then easily decomposed into the

NH3 + OH (O, H) ⇔ NH2 + H2O (OH, H2) (3)

When hydrocarbon is not contained in the fuel, NHi is converted into N2 by reacting with NO in the fuel-rich region. If fuel contains CH4, HCN is produced by reactions 5 and 6 in the fuel-rich region and the HCN is oxidized to NO in the fuel-lean zone (Heap et al, 1976) and

Some HCN is oxidized into NO by reactions 7 and 8, and the rest is decomposed into N radical by the reaction 9. NH radical is decomposed into the NO by reactions 10, 11, and 12. With the rise in CH4 concentration in gasified fuel, the HCN increases, and NOx emissions

 N + OH ⇔ NO + H (12) On the other hand, some NH radical produced by the reactions 3, 4 and 5 are reacted with Zel'dovich NO, Prompt NO and fuel-N oxidized NO, which produced by above reactions,

That is, it is surmised that each of increase in thermal-NOx concentration and fuel-NOx affected the alternative decomposition reaction of intermediate NH radical with NO, so the

These new techniques those adopted the nitrogen direct injection and the two-stage combustion, caused a decrease in flame temperature in the primary combustion zone and the thermal-NOx production near the burner was expected to be controlled. On the contrary, we were afraid that the flame temperature near the burner was declined too low at lower load conditions and so a stable combustion cannot be maintained. The designed combustor was given another nitrogen injection function, in which nitrogen was bypassed to premix with the air derived from the compressor at lower load conditions, and a stable flame can be maintained in a wide range of turn-down operations. Also, because the

each of NOx emissions originated from the nitrogen in the air or fuel-N decreased.

originated from HCN in the fuel-lean secondary combustion zone increase.

and decomposed into N2 by the reaction 13.

NHi (i=1, 2) + OH (H) ⇔ NHi-1 + H2O (H2) (4)

CHi (i=1,2) + N2 ⇔ HCN + NHi-1 (5)

HCN + OH ⇔ CN + H2O (7)

CN + O2 ⇔ CO + NO (8)

CN + O ⇔ CO + N (9)

NH + OH ⇔ N + H2O (10)

N + O2 ⇔ NO + O (11)

NO + NH ⇔ N2 + OH (13)

R-CH + NHi ⇔ HCN + R-Hi, (R- : Alkyl group) (6)

intermediate NHi by the following reactions (Miller et al., 1983).

(Takagi et al, 1979).

through a swirler, (which has a 30-degree swirl angle and a 15-degree introverted angle), to collide medium-Btu fuel with air in an atmosphere where nitrogen is superior in amount to both fuel and air.

#### **3.3.2.3 Fuel-NOx/Thermal-NOx reduction**

In order to decrease fuel-NOx emissions, we adopted fuel-rich combustion in the primary zone and set the equivalence ratio in the primary-combustion zone is determined based on the combustion test results using a small diffusion burner shown in figure 8. Figure 19 presents a relation between the primary equivalence ratio, p, and the conversion rate of NH3 to NOx, C.R., with CH4 concentration as a parameter in two-staged combustion. In test, the average temperature of the exhaust, *T*ex, is set to 1773K and fuel calorific value is 11.4MJ/m3 for fuel containing 1000ppm of NH3, CO and H2 of 2.33 CO/H2 molar ratio. In the same way as low-Btu fuels, the primary equivalence ratio that minimizes the conversion rate of NH3 to NOx is affected by CH4 concentration in the fuel. Because the supplied fuel contains 3 percent of CH4, the equivalence ratio in the primary-combustion zone was set around 1.9 and the equivalence ratio in the auxiliary-combustion chamber was around 2.4 to maintain the flame stabilization and to improve reduction of NH3, simultaneously.

Fig. 19. Effect of the CH4 concentration on conversion rate of NH3 to NOx in two-stage combustion of medium-Btu fuel

The effect of the CH4 concentration on the fuel-NOx produced by NH3 in gasified fuel was studied using the elementary reaction kinetics (Hasegawa et al., 2001). The model of the flow inside the combustor introduced the Pratt model (Pratt et al., 1971) and each stage combustion zone is assumed to be a perfectly stirred reactor. The reaction model employed here was proposed by Miller and Bowman(1989), values for thermodynamic data were taken from the JANAF thermodynamics tables(Chase et al., 1985) or calculated based on the relationship between the Gibbs' standard energy of formation and the chemical equilibrium constant. The values of Gibbs' standard energy of formation were obtained from the CHEMKIN database (Kee et al., 1990). The GEAR method (Hindmarsh, 1974) was used for the numerical analysis. Also, it is assumed that the species are evenly mixed, and diffusion and stirring processes are not taken into consideration in the reaction process. The appropriateness of the model for reaction NH3 with NO in the gasified fuels (Hasegawa, 1998c) has been confirmed by comparison with test results.

through a swirler, (which has a 30-degree swirl angle and a 15-degree introverted angle), to collide medium-Btu fuel with air in an atmosphere where nitrogen is superior in amount to

In order to decrease fuel-NOx emissions, we adopted fuel-rich combustion in the primary zone and set the equivalence ratio in the primary-combustion zone is determined based on the combustion test results using a small diffusion burner shown in figure 8. Figure 19

NH3 to NOx, C.R., with CH4 concentration as a parameter in two-staged combustion. In test, the average temperature of the exhaust, *T*ex, is set to 1773K and fuel calorific value is 11.4MJ/m3 for fuel containing 1000ppm of NH3, CO and H2 of 2.33 CO/H2 molar ratio. In the same way as low-Btu fuels, the primary equivalence ratio that minimizes the conversion rate of NH3 to NOx is affected by CH4 concentration in the fuel. Because the supplied fuel contains 3 percent of CH4, the equivalence ratio in the primary-combustion zone was set around 1.9 and the equivalence ratio in the auxiliary-combustion chamber was around 2.4 to

> 012345 φp

φp

Fig. 19. Effect of the CH4 concentration on conversion rate of NH3 to NOx in two-stage

The effect of the CH4 concentration on the fuel-NOx produced by NH3 in gasified fuel was studied using the elementary reaction kinetics (Hasegawa et al., 2001). The model of the flow inside the combustor introduced the Pratt model (Pratt et al., 1971) and each stage combustion zone is assumed to be a perfectly stirred reactor. The reaction model employed here was proposed by Miller and Bowman(1989), values for thermodynamic data were taken from the JANAF thermodynamics tables(Chase et al., 1985) or calculated based on the relationship between the Gibbs' standard energy of formation and the chemical equilibrium constant. The values of Gibbs' standard energy of formation were obtained from the CHEMKIN database (Kee et al., 1990). The GEAR method (Hindmarsh, 1974) was used for the numerical analysis. Also, it is assumed that the species are evenly mixed, and diffusion and stirring processes are not taken into consideration in the reaction process. The appropriateness of the model for reaction NH3 with NO in the gasified fuels (Hasegawa,

maintain the flame stabilization and to improve reduction of NH3, simultaneously.

3

Tex=1773K HHV=11.4MJ/m NH3=1000ppm

CH4 % 0.0 0.5 1.0 2.0 3.0

p, and the conversion rate of

both fuel and air.

**3.3.2.3 Fuel-NOx/Thermal-NOx reduction** 

presents a relation between the primary equivalence ratio,

0

20

40

60

C.R. %

C.R. %

1998c) has been confirmed by comparison with test results.

combustion of medium-Btu fuel

80

100

The nitrogen of NH3 in the fuel has weaker bonding power than N2. In the combustion process, NH3 reacted with the OH, O, and H radicals and then easily decomposed into the intermediate NHi by the following reactions (Miller et al., 1983).

$$\text{NH}\_3 + \text{OH} \text{ (O,H)} \leftrightarrow \text{NH}\_2 + \text{H}\_2\text{O (OH,H\_2)}\tag{3}$$

$$\text{NHH (i=1,2)} + \text{OH (H)} \Leftrightarrow \text{NHH-1} + \text{H}\_2\text{O (H}\_2\text{)}\tag{4}$$

When hydrocarbon is not contained in the fuel, NHi is converted into N2 by reacting with NO in the fuel-rich region. If fuel contains CH4, HCN is produced by reactions 5 and 6 in the fuel-rich region and the HCN is oxidized to NO in the fuel-lean zone (Heap et al, 1976) and (Takagi et al, 1979).

$$\text{CHi (i=1,2)} + \text{N}\_2 \Leftrightarrow \text{HCN} + \text{NHi-1} \tag{5}$$

$$\text{R-CH} + \text{NHi} \Leftrightarrow \text{HCN} + \text{R-Hi}, \text{(R-} \cdot \text{Alkyl group)}\tag{6}$$

Some HCN is oxidized into NO by reactions 7 and 8, and the rest is decomposed into N radical by the reaction 9. NH radical is decomposed into the NO by reactions 10, 11, and 12. With the rise in CH4 concentration in gasified fuel, the HCN increases, and NOx emissions originated from HCN in the fuel-lean secondary combustion zone increase.

$$\text{CHCN} + \text{OH} \leftrightarrow \text{CN} + \text{H}\_2\text{O} \tag{7}$$

$$\text{CN} + \text{O}\_2 \Leftrightarrow \text{CO} + \text{NO} \tag{8}$$

$$\text{CH} + \text{O} \Leftrightarrow \text{CO} + \text{N} \tag{9}$$

$$\text{NH} + \text{OH} \Leftrightarrow \text{N} + \text{H}\_2\text{O} \tag{10}$$

$$\text{N} + \text{O}\_2 \Leftrightarrow \text{NO} + \text{O} \tag{11}$$

$$\text{N} + \text{OH} \leftrightarrow \text{NO} + \text{H} \tag{12}$$

On the other hand, some NH radical produced by the reactions 3, 4 and 5 are reacted with Zel'dovich NO, Prompt NO and fuel-N oxidized NO, which produced by above reactions, and decomposed into N2 by the reaction 13.

$$\text{NiO} + \text{NH} \leftrightarrow \text{N}\_2 + \text{OH} \tag{13}$$

That is, it is surmised that each of increase in thermal-NOx concentration and fuel-NOx affected the alternative decomposition reaction of intermediate NH radical with NO, so the each of NOx emissions originated from the nitrogen in the air or fuel-N decreased.

These new techniques those adopted the nitrogen direct injection and the two-stage combustion, caused a decrease in flame temperature in the primary combustion zone and the thermal-NOx production near the burner was expected to be controlled. On the contrary, we were afraid that the flame temperature near the burner was declined too low at lower load conditions and so a stable combustion cannot be maintained. The designed combustor was given another nitrogen injection function, in which nitrogen was bypassed to premix with the air derived from the compressor at lower load conditions, and a stable flame can be maintained in a wide range of turn-down operations. Also, because the

Developments of Gas Turbine Combustors for Air-Blown and Oxygen-Blown IGCC 261

combustor's characteristics enough. Furthermore, developed technologies represent a

To keep stable supplies of energy and protect the global environment, it will be important that human beings not only use finite fossil fuel, such as oil and coal, but also reexamine unused resources and reclaim waste, and develop highly effective usage of such resources. The IGCC technologies could have the potential to use highly efficient resources not widely

The author wishes to express their appreciation to the many people who have contributed to

Battista, R.A. & Farrell, R.A. (1979). Development of an Industrial Gas Turbine Combustor

Burning a Variety of Coal-Derived Low Btu Fuels and Distillate, *ASME paper,* 

CO/H2 Molar ratio of carbon monoxide to hydrogen in fuel [mol/mol]

HHV Higher heating value of fuel at 273 K, 0.1 MPa basis [MJ/m3]

NOx(16%O2) NOx emissions corrected at 16% oxygen in exhaust [ppm]

C.R. Conversion rate from ammonia in fuel to NOx [%]

*T*ex Average temperature of combustor exhaust gas [K]

ex Average equivalence ratio at combustor exhaust

p Average equivalence ratio in primary combustion zone

No.79-GT-172, San Diego, California, USA, March 11-15, 1979.

N2/Fuel Nitrogen over fuel supply ratio [kg/kg]

*P* Pressure inside the combustor [MPa] *T*air Temperature of supplied air [K]

*T*fuel Temperature of supplied fuel [K] *T*N2 Temperature of supplied nitrogen [K]

Synthetic gas cleanup Wet type Hot/Dry type

> ・1773K-class combustor ・NOx emissions ≦60ppm\* ・thermal-NOx ≦ 8ppm\* ・P.F.(rated) ≦ 8%

> ・1773K-class combustor ・NOx emissions ≦34ppm\* ・thermal-NOx ≦8ppm\* ・P.F.(rated) ≦ 7%

possible step towards the 1873K-class gas turbine combustor.

・1573K-class gas turbine combustor for BFG ・thermal-NOx ≦20ppm\*

・1573K-class combustor ・thermal-NOx ≦11ppm\* ・P.F.(rated) =10~13%

in use today for power generation.

Air

O2

\* : Concentration corrected at 16% oxygen in exhaust Table 6. Performances of gasified fueled combustors

Gasification agent

**5. Acknowledgment** 

this investigation.

**6. Nomenclature** 

**7. References** 

nitrogen dilution in the fuel-rich region affected the reduction characteristics of NH3, the increase in nitrogen dilution raised the conversion rates of NH3 to NOx. This tendency showed the same as that of the case where nitrogenous compounds in fuel increased, indicated by Sarofim et al.(1975), Kato et al.(1976) and Takagi et al.(1977). That is, it is necessary that the nitrogen bypassing technique is expected to improve fuel-NOx reduction in the cases of higher concentration of NH3.

#### **3.3.3 Test results**

Supplied fuels into the combustor were adjusted as same propertied as that of the slurryfeed coal gasified fuel. In tests, the effects of the CH4 concentrations, etc. in the supplied fuels on the combustion characteristics were investigated and the combustor's performances were predicted in the typical commercial operations. Figure 20 estimates the combustion emission characteristics under the simulated operational conditions of 1773K-class gas turbine for IGCC in the case where gasified fuel contains 0.1 percent CH4 and 500ppm NH3. Total NOx emissions were surmised as low as 34ppm (corrected at 16 percent O2) in the range where the gas turbine load was 25 percent or higher, which is the single fuel firing of gasified fuel, while the NOx emissions tend to increase slightly with the rise in the gas turbine load. In the tests of the simulated fuel that contained no NH3, thermal-NOx emissions were as low as 8ppm (corrected at 16 percent O2). On the other hand, we can expect that combustion efficiency is around 100 percent under operational conditions of the medium-Btu fueled gas turbine.

Fig. 20. Combustion emission characteristics

## **4. Conclusion**

Based on basic combustion test results using small burners and model combustors, Japanese electric industries proposed the correspond combustion technologies for each gasified fuels, designed combustors fitted with a suitable nitrogen injection nozzle, two-stage combustion, or lean combustion for each gasified fuel, and demonstrated those combustors' performances under gas turbine operational conditions. As summarized in Table 6, the developed combustors showed to be completely-satisfied with the performances of 1773Kclass gas turbine combustor in the actual operations. That is, these combustion technologies reduced each type of NOx emissions for each gasified fuel, while maintaining the other combustor's characteristics enough. Furthermore, developed technologies represent a possible step towards the 1873K-class gas turbine combustor.

To keep stable supplies of energy and protect the global environment, it will be important that human beings not only use finite fossil fuel, such as oil and coal, but also reexamine unused resources and reclaim waste, and develop highly effective usage of such resources. The IGCC technologies could have the potential to use highly efficient resources not widely in use today for power generation.


\* : Concentration corrected at 16% oxygen in exhaust

Table 6. Performances of gasified fueled combustors

## **5. Acknowledgment**

260 Advances in Gas Turbine Technology

nitrogen dilution in the fuel-rich region affected the reduction characteristics of NH3, the increase in nitrogen dilution raised the conversion rates of NH3 to NOx. This tendency showed the same as that of the case where nitrogenous compounds in fuel increased, indicated by Sarofim et al.(1975), Kato et al.(1976) and Takagi et al.(1977). That is, it is necessary that the nitrogen bypassing technique is expected to improve fuel-NOx reduction

Supplied fuels into the combustor were adjusted as same propertied as that of the slurryfeed coal gasified fuel. In tests, the effects of the CH4 concentrations, etc. in the supplied fuels on the combustion characteristics were investigated and the combustor's performances were predicted in the typical commercial operations. Figure 20 estimates the combustion emission characteristics under the simulated operational conditions of 1773K-class gas turbine for IGCC in the case where gasified fuel contains 0.1 percent CH4 and 500ppm NH3. Total NOx emissions were surmised as low as 34ppm (corrected at 16 percent O2) in the range where the gas turbine load was 25 percent or higher, which is the single fuel firing of gasified fuel, while the NOx emissions tend to increase slightly with the rise in the gas turbine load. In the tests of the simulated fuel that contained no NH3, thermal-NOx emissions were as low as 8ppm (corrected at 16 percent O2). On the other hand, we can expect that combustion efficiency is around 100 percent under operational conditions of the

> 0 20 40 60 80 100 Gas Turbine Lord %

Based on basic combustion test results using small burners and model combustors, Japanese electric industries proposed the correspond combustion technologies for each gasified fuels, designed combustors fitted with a suitable nitrogen injection nozzle, two-stage combustion, or lean combustion for each gasified fuel, and demonstrated those combustors' performances under gas turbine operational conditions. As summarized in Table 6, the developed combustors showed to be completely-satisfied with the performances of 1773Kclass gas turbine combustor in the actual operations. That is, these combustion technologies reduced each type of NOx emissions for each gasified fuel, while maintaining the other

99.5

99.6

99.7

99.8

η %

Combustion efficiency %

99.9

100

0

20

HHV=8.8MJ/m NH3=500ppm CH4=0.1%

3

40

60

NOx(16%O2) ppm

Fig. 20. Combustion emission characteristics

**4. Conclusion** 

80

100

in the cases of higher concentration of NH3.

**3.3.3 Test results** 

medium-Btu fueled gas turbine.

The author wishes to express their appreciation to the many people who have contributed to this investigation.

## **6. Nomenclature**


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**12** 

*France* 

Georges Descombes

**Characterization of a Spray in the Combustion** 

The use of a turbo-alternator in Lean Premixed Prevaporized combustion (LPP) for hybrid vehicles is beneficial in reducing pollutant emissions at the nominal operating point. The electric thermal hybrid demonstrator studied here consists of a low-emission gas turbine and an alternator which provides the electric power to an electric propulsion motor and a

The combustion chamber of the gas turbine is adapted to the nominal operating point so as to function in pre-vaporized combustion, premixed and lean mixtures. A problematic point, however, is the emission of smoke and unburnt hydrocarbons during start-up because the

In the transitional stages of start, an air-assisted pilot injector vaporizes the fuel in the combustion chamber. The jet is ignited by a spark, the alternator being used as an electric starter. This starting phase causes, however, the formation of a fuel film on the walls which

1 2 3 4 5 6 Exchanger Fuel Ignition Turbine Compressor Alternator

geometry of the combustion chamber is not adapted to moderate air flows.

**1. Introduction** 

storage battery.

can be observed as locally rich pockets.

Fig. 1. Diagram of the turbo alternator

**Chamber of a Low Emission Gas Turbine** 

*Laboratoire de génie des procédés pour l'environnement, l'énergie et la santé* 

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