B URNING V ELOCITY
3.4. PROCESSES AND APPLICATIONS OF MILD COMBUSTION IN GAS TURBINES
3.4.1. MILD C OMBUSTION WITH E XTERNAL C ONTROL
The dilution of the oxidant with inert gases can be achieved in open and closed gas turbine cycles by means of species that are not related to the gas turbine itself. In principle, they can be any kind of gas, but for economic reasons, a feasible choice for open cycles is flue gas from a combustion source. This makes the gas turbine outlet flow a possible candidate, as in sequential combustion and in systems with external recircu- lation. An analysis of a system with an independent source of inert flows demonstrates the theoretical limitations of such a system. To use flue gas as this source of diluent, it must be obtained from a relatively clean combustion process so that it is free from
particulates and corrosive gases that could degrade the operation and life of the compressor and turbine. Under these constraints, the flue gas and the oxidant, say air, can be mixed, compressed, and distributed to the combustion chamber inlet with the oxygen content consistent with ignition and maximum allowable temperatures for MILD combustion.
The simplest reactor configuration pertains to the most advanced type of gas turbine.
High temperature and pressure, which favor autoignition, are created at the combustion chamber inlet to achieve very high compression ratios. These elevated ratios are convenient not only for very high compressor/turbine efficiencies, but also for high turbine inlet temperatures. This is shown in Figure 3.18 (and Chapter 5) where the efficiencies of a Brayton cycle, which is based upon thermodynamic conditions sketched in the inset of the figure, are reported for two different final combustion temperatures and two values of the compressor/turbine efficiency. The inset shows on a temperature-entropy diagram, the isentropic compressor (1–2s) and turbine (3–4s) paths, which shift to the 1–2 and 3–4 paths in order to take into account the real non- isentropic transformation. Combustion takes place at nearly constant pressure along the path 2–3.
Efficiencies depend on the gas turbine compression ratiop2=p1, and are reported in Figure 3.18 for two temperaturesT3and for constant compressor and turbine efficiency, candt, which are assumed to be constant and equal to the isentropic efficiencyiso. The efficiencies are given by Equation (3.9), obtained by simple relations (Hernandez et al., 1995) wheret¼T1=T3 is the temperature ratio, and¼p3=p4¼p2=p1 is the compression ratio.
¼ ttcp(g 1)=g
1p(g11)=g
1tt
c½p(g 1)=g1 n o
(1r)þrt 1p(g11)=g
: (3:9)
All of the curves in Figure 3.18 reveal that the efficiency maxima are situated at compression ratios ranging between 3 and 25. For higher values of compression ratios, the efficiencies decrease continuously. Higher final temperature values of the combustion process,T3, shift the efficiency curve toward higher values and make the compression
16001500 2s 2 4s
T 3
4
S 1
x
T3, K h
ηiso=0.8
ηiso = 0.7 0.5
0.4 0.3 0.2 0.1
00 10 20
p 30 40 50
p=const.
p=const.
Figure 3.18 Efficiency of Brayton cycle as a function of compression ratio for two final combustion temperatures (T3) and two compressor/turbine efficiencies (iso).
ratio go up to 35–40 which is reasonable for practical applications. Limitations on the net power output suggest limiting compression ratios range between 30 and 35.
Figure 3.19 illustrates the relevance of high compression ratios and the consequent need for high values ofT3 andiso for suitable use in MILD combustion. The figure reveals the dependence of both the inlet temperature and the oxygen concentration on compression ratio at inlet combustion chamber pressure, which is the same as that of the compressor outlet. The first quantity, sketched with a short dashed line, is obtained by means of the isentropic compression relation p(g 1)=g, and the oxygen concentration, represented by a continuous line, is obtained by means of an enthalpy balance between the inlet and the outlet of the combustion chamber, as is displayed schematically in the inset of the figure. The example refers to decane fuel mixing with air and flue gas. This mixture burns between the inlet temperature values and an outlet temperature value fixed at 1700 K. The autoignition delay time is calculated according to the procedures reported in the paper by de Joannonet al. (2002) and it is reported in the figure a dotted line by referring to the axis on the right side of the figure.
The comparison of these autoignition delays with those in non-diluted cases con- firms that the beneficial effect of increases in pressure and temperature is partially counter-balanced by the dilution itself. This has the effect of decreasing the oxygen content in the oxidant stream. It has been shown (de Joannon et al., 2002) that this parameter affects the autoignition delay following a power law with an exponent of 0.3 for temperatures lower than 700 K, of around unity for temperatures higher than 1000 K, and of around 2 in a narrow temperature range centered around 900 K.
Another possible application of MILD combustion is in regenerative gas turbines.
This configuration allows for increasing inlet temperature in the combustion chamber for low values of compression ratio. It is well known that the Brayton cycle follows the path drawn in the inset of Figure 3.18. The isobaric heating of the compressed flow occurs between theT2andTxtemperatures by means of a counter-flow heater fed by the turbine flue gas with an efficiency given by Equation (3.10):
r¼TxT2
T4T2; (3:10)
wherer is 0 for the non-regenerative case analyzed before and is unity in the upper limit of regeneration when Tx is equal to T4. This is the maximum temperature the
Tin, K
Tin, XO2 tign
Tout =1700K tign, s XO2
T
Combustion chamber
XO2
.14
.12
.1
.08
Compression ratio
10
1 0.1
0.01
0.001
0 10 20 30 40
900
700
500
300
Figure 3.19 Temperature and oxygen concentrations at combustion chamber inlet versus compression ratio for a fixed inlet turbine temperature on the left axes and the related ignition delay time on the right axis.
compressed flow can reach because it is heated by the flue gas, which is discharged at this temperature. The gas turbine efficiencies for this condition can be calculated by means of Equation (3.9) for regenerator efficiencyr ¼0:7, 0:8, or 0.9. Their values are plotted in Figure 3.20 for fixed turbine inlet temperaturesT3¼1500 and 1600 K and for a fixed compressor-turbine efficiency of 0.8. The six curves in Figure 3.20 for the regenerative gas turbine are similar. They increase steeply at very low compression ratios and reach a maximum at a pressure range between 3 and 7 bar, then they decrease slightly, crossing a common value at a compression ratio around 20.
In an internal recirculation system, the diluent comes from within the same unit, and the system must be designed to minimize energy losses related to the use of recircula- tion. In particular, the high temperature of the flue gases means that higher compression work is needed compared to that for cold gases in traditional systems without recircula- tion. In order to overcome this problem, the two streams need to be at the same pressure and temperature. Therefore, the oxidant, for instance air, and the flue gases must be compressed in different units and heated in separate recuperator and heat recovery systems. Figure 3.21 provides a schematic of a plant layout that applies the concepts for a semi-closed combined cycle gas turbine proposed by Camporealeet al. (2004). Air is compressed in the low pressure compressor C1 and heated in a recuperator R2. Then it is mixed with recirculated flue gases which are compressed with the high pressure compressor C2 and heated in the recuperator R1. The mass flow, the compressor, and the heaters are adjusted in such a way that the oxygen concentration is 10% and the temperature and pressure are 1000 K and 20 bar, respectively. The combustion gases at 1700 K are expanded beforehand in the cooled-blade turbine T1, which first feeds the recirculation loop and the heat recovery steam generator HRSG1 with recirculation ratio 0.6. It then feeds the low pressure turbine T2 and the heat recovery steam generator HRSG2. The thermodynamic cycle of this plant has been assessed to be around 0.6 for a power corresponding to 100 kg/s under realistic efficiency assumptions of the components.
Another potential application of MILD combustion is in closed cycle turbine systems. Here, the fluid which operates the cycle is a gas which is usually externally heated by means of a heat exchanger fed by a fluid at high temperature. It is possible to
16001500 T3, K
hr= 0.7 hr= 0.8
hiso= 0.8
h
hr= 0.9 0.5
0.4 0.3 0.2 0.1 0
0 10 20 p 30 40 50
Figure 3.20 Efficiency of Brayton cycle as function of the compression ratio for different values of regeneration efficiency.
substitute the heater with an internal heating system based on MILD combustion, when the pressure and temperature in the heater and the fluid are suitable. The temperature and pressure would have to be sufficiently high for autoignition to occur in a time consistent with the allowable residence time in the engine, so that the combustion products can be separated by the working fluid. This approach permits the reduction of pressure drop along the cycle gas flow and shifts the maximum temperature in the heater from the external to the internal side for an increase of the total efficiency of the system. A possible application is steam reheating in a Hirn cycle, which is illustrated by the temperature-entropy plot of Figure 3.22 (Milani and Saponaro, 2001).
According to the figure, the first and second heating of the fluid up to the same temperature (around 832 K or 560°C) is performed at high (point 6 at 210 bar) and low (point 8 at 20 bar) pressure by means of the same heat exchangers used in Rankine cycles and fed with flue gas generated by traditional combustion systems. In contrast,
Figure 3.21 Plant layout of a semi-closed combined cycle gas turbine plant where the MILD combustion concept is used (adapted from Camporealeet al., 2004).
the subsequent heating of the steam is performed by hydrogen oxidized with pure oxygen in the steam stream. The flow rate of the reactants is adjusted in the figure in order to reach a temperature of 1503 K (1230°C) (point 9). Finally, expansion pushes the pressure down to 0.06 bar and the temperature down to 513 K (240°C), and is followed by a partial cooling in order to fix the condenser conditions to a temperature and pressure of 320 K (47°C) and 0.05 bar, respectively. Excess steam produced by combustion is discharged after the condenser and the remainder of the steam is recircu- lated through the cycle.
This type of cycle is particularly appropriate for repowering applications, doubling the original power and simultaneously increasing efficiency. In the example presented in Figure 3.22 and analyzed by the authors (Milani and Saponaro, 2001), the efficiency is 0.6, which should be compared to that of a plant without repowering, which is 0.4.
Another interesting application of this type of cycle is when hydrogen is produced in a coupled plant of coal hydro-gasification with separation and sequestration of carbon dioxide. Research activity is intense in this field particularly in Italy (Calabret al., 2005) in order to exploit hydrogen economically and to reduce greenhouse effects. Of course, critical steps must be faced in order to develop feasible plants based on this type of cycle. In some cases, the pressure and temperature can be adjusted to relatively high levels. This is the case for the supercritical steam turbine described in Figure 3.23 by a dashed line, where the pressure and temperature can reach values on the order of 10 bar and 800 K, respectively.
With these conditions, injecting fuel and oxidant into the steam flow should allow the mixture to reach its ignition temperature. The ignition delay time depends on the type of fuel, as is shown in Figure 3.23, where the temporal evolution of temperature is reported for two fuels, hydrogen and methane, and for two concentrations of the fuels at stoichiometric conditions. For a fixed fuel type, the autoignition delay is nearly the same for the two concentrations, whereas the maximum temperature increases with concen- tration. This shows how the power input can be adjusted according to the needed power output. Since the plots represent only one set of specific conditions, they need to be further validated because the kinetics of ter-molecular reactions with the chaperon effect based on water have not yet been completely assessed, and the numerical predictions are
State State
1
T P M T P M
Steam Hydrogen-Steam
2 3 4 5 6 7 8 9 10 11
1400 1200 1000 800 600 400 200 0
Temperature (⬚C)
1 3
4 5
6 8
7
9
10 11 S (KJ/Kg/⬚C)
0 2 4 6 8 10
Figure 3.22 Hirn cycle in \raster "fig22"temperature-entropy plot (Milani and Saponaro, 2001).
based on models developed in relatively low concentrations of water. Preliminary studies in shock tubes and plug flows show that the expected enhancement of recombin- ation in the reaction, as shown in Equation (3.11), plays an effective role and tends to depress the reactivity of the system. Consequently, further studies are needed and are under consideration (Sabiaet al., 2006) in this field.
HþO2þH2O!HO2þH2O; (3:11) For the situation where an internal heating system is based upon the use of hydrogen, the final combustion product is water, so part of the water must be removed and may eventually be used for other purposes. In contrast, the methane system produces water and carbon dioxide, the latter of which must also be removed by degasification after steam condensation. The employment of other fuels is also feasible, but the fuel has to be particularly clean. Otherwise, a cleaning system must be added in order to reuse the steam in the closed cycle. Nonetheless, it is of interest to explore this possibility because the separation of solid as well as solute species in the final condensed polluted water is generally easier than in gaseous flows. As a result, the use of this more complex system becomes more attractive from economic and environmental standpoints.
Applications of water-diluted combustion have not yet appeared for gas turbines even though a limited amount of steam injection has been utilized, and some furnaces at atmospheric pressure have exploited“synthetic air”formed by the addition of oxygen in steam in the same percentage as the air. The combustion process is of some interest because it occurs without the presence of nitrogen, and it is therefore impossible to form any kind of nitrogen oxide or other species in which nitrogen may also be present.
Furthermore, water may be beneficial in incineration processes because it may be a source of OH radicals, which are very reactive in the oxidation processes of any gaseous or condensed organic species.
Although there is a lack of facilities, and consequently combustors, that are actually built, there are some interesting studies that have shown the practical feasibility of such cycles. One of these studies describes the analysis of a 400 MW (electric) zero-atmos- pheric emission power plant, in which the working fluid is water and the carbon dioxide is removed and injected into a well for sequestration (Martinez-Frias et al., 2003).
A schematic of the core of this plant is shown in Figure 3.24, in which the numbers p = 10 bar
H2/O2/N2
To CH4/O2/N2 Fuel/Oxygen = stoichiometric
Temperature (K)
3500 3000 2500 2000 1500 1000 500
Time (s)
0 2 4 6 8 10
Figure 3.23 Temporal evolution of temperature for hydrogen and methane and for two oxygen concentrations at pressurep 10 bar and initial temperatureT 800 K.
reported along the connecting branches are respectively the pressure (bar) and the temperature (K) and the mass flow (Kg/s). The first unit is kept at 124 bar and is fed by O2=CH4at a nearly stoichiometric ratio as well as H2O at 600 K with mass flow rates so high that the outlet temperature is less than 1100 K (not reported in figure). Maximum temperature is quite mild due to the large mass flow ratio between the diluent and the fuel, and autoignition can take place because the steam is partially heated by the reheater placed just downstream of the first high-pressure turbine. A second combustion stage, fed by oxygen and methane, brings the working fluid from 400 K up to 1500 K at a chamber pressure of 11 bar. The inlet temperature and pressure are slightly lower than for autoignition, but neither slight adjustments in operating conditions nor exchanging in place the reheater with the combustion chamber seem to significantly change the overall plant efficiency. At the end, the working fluid consists of a 0.8 mass fraction of steam and 0.2 mass fraction of carbon dioxide, such as results from mass balance on combustion chamber. It enters intermediate and low-pressure stages of the turbine system and converts its thermal power into mechanical power. The exhaust flows through a pre- heater and reaches the condenser where a great part of the water and carbon dioxide is separated, but still contains some moisture. The subsequent seven-stage compression and intercooling systems allow a final separation and compression of the carbon dioxide up to 150 bar for enhanced recovery of oil or coal-bed methane, or sequestration.
A similar concept is applied in the so-called Graz cycle (Heitmeir and Jerticha, 2003), developed via cooperation with Japanese public and private research centers. This cycle is illustrated in Figure 3.25 by means of a temperature-entropy plot. Oxy-fuel combustion takes place along the isobaric branch at 40 bar inside a mixture of steam and carbon dioxide, with inlet and outlet temperatures of 600 and 1700 K, respectively. Expansion, cooling, and re-expansion of the working fluid takes place in the 2–3, 3–4, and 4–5 branches. The two
Figure 3.24 Zero-atmospheric emission power plant with water as working fluid (adapted from Martinez- Friaset al., 2003).
species then separate in the condenser where the water undergoes compression of the working fluid in the liquid phase as well as heating and expansion similar to other H2O cycles, whereas the CO2is compressed for partial recycling and partial sequestration.
Inert gases other than water can be used in closed gas turbine cycles. For instance, many working fluids have been studied for heat exploitation applications pertaining to nuclear reactors due to security restrictions or to relatively low temperature ratios (Dostalet al., 2001), and to solar-assisted cycles (Popelet al., 1998). Specifically, a supercritical carbon dioxide Brayton cycle has been proposed (Popelet al., 1998; Mathieu and Nihart, 1999) in order to make use of the large heat capacity and intense heat transfer capabilities of this fluid with respect to other fluids at supercritical conditions, which are reached for carbon dioxide at much lower temperatures than for water. Other examples, that include the MATIANT cycles (Mathieu and Nihart, 1999), are used not only for nuclear applications, but also for removing carbon dioxide from combustion processes in the liquid state for reuse or final storage. They are based on several cycles like Rankine, Ericsson, and Brayton with or without regeneration, but all of them employ CO2 as the main working fluid (Mathieuet al., 2000). One of these cycles, as can be seen in Figure 3.26, is a regenerative Ericsson-type CO2 cycle (Sundkvistet al., 2001). The fluid undergoes multi-step inter- cooled compression in the 1–2 branch, and is then heated in a regenerator up to 900 K–the same temperature as combustion chamber inlet conditions. This elevated temperature, coupled with a very high pressure of 110 bar, makes the mixture autoignitable, while oxygen and fuel are injected in very dilute conditions so that the final temperature is only 1600 K. After the expansion in the turbine (4–5), there is a second combustion process in the reheater branch (5–6), which lies in the MILD combustion regime since the temperature is between 1400 and 1600 K at 40 bar. Finally, the expansion in the 6–7 branch is followed by cooling of the working fluid in an internal regenerative exchanger (7–8) and water- cooled system that is also used to extract water produced in the combustion process.
1600 1400 1200 1000 800 600 400 200 0
0 2
6 7
8 1
4 5
10 9 3 2
4 Condenser
6 Entropy (kJ/kg K)
LP tubine
H2O saturation line HP turbine
HT turbine Combustor
Temperature (⬚C)
Combustor
O2 CO
2 HTT
1400⬚(C) 1 bar 642⬚(C)
180 bar 567⬚(C)
160⬚(C) 0.25 bar Fuel
40 bar Steam
HPT
HRSG Feed pump
C3
Dearetor CO2 Condenser Water Cond. P
CO2 C2
C1
CO2 compressor
8 10
Figure 3.25 Graz cycle on temperature-entropy plot (adapted from Heitmeir and Jerticha, 2003).