407
11.3 Performance of Simple-Cycle Power Gas Turbines
(results not shown). This is mainly due to the higher fuel-air ratio as a consequence
of the lower heating value 43 MJ/kg.
The efficiency results of Fig. 11.14 are obtained with cooling air extracted from
the compressor, as in most machines. Further, it is assumed that there is no precooling of this air by an external heat exchanger before supplying it to the turbine.
In some machines such pre-cooling is implemented. Cooling may also be realised
with an externally provided fluid. In particular, machines exist with steam cooling.
The choice for this cooling fluid comes from applications in combined-cycle systems (see Sect. 11.4.4). The heat capacity of steam is higher than that of air, which,
principally, lowers the necessary cooling flow rates. Further, the heat absorbed by
the steam may be used in the steam cycle. However, there are some drawbacks
with steam too. Steam has to be led into the machine, but also led out of it. This is
possible, but is somewhat complicated for rotor parts. Further, film cooling cannot
be used. This increases the heat transfer between the hot gas in the turbine and the
blades and the vanes ( κ = 0.20 J/kJK instead of 0.15 J/kJK) and decreases the heat
exchanger effectiveness of the blades and the vanes ( ε = 0.7 instead of 1). Simulations with cooling with an external fluid have as a result that the global efficiency
of the simple cycle is slightly lower than with cooling by air from the compressor
(results not shown). But these are results assuming that the heat transferred to the
cooling fluid is useless. Taking into account that the heat may contribute to a second
cycle, one may conclude that there is a small advantage with steam cooling in a
combined-cycle application. We will come back to closed loop steam cooling with
the discussion of combined cycles in Sect. 11.4.4.
Figures 11.15, 11.16, 11.17 and 11.18 show specific power, turbine outlet temperature (TOT), fuel flow rate and cooling flow rate obtained by simulation with
pure CH 4 as fuel. The results with kerosene as fuel differ slightly. For instance,
specific power is about 4–5 % lower with kerosene. The specific power is expressed
as the power per unit of mass flow rate of the exhaust gas. The increase of the specific power with the combustion temperature is obvious. The specific power has
a weak maximum as a function of pressure ratio, around pressure ratio 15–20. An
important observation for practise is that with respect to specific power, it is not
useful to choose the pressure ratio too high. Heavy duty gas turbines have typically
a pressure ratio in the order of 20. The simple-cycle efficiency (Fig. 11.14) is then
not at maximum, but these machines are typically used in combined-cycle power
stations and the optimum efficiency with combined cycles occurs at lower pressure
ratio (see Sect. 11.4.4). It is difficult to verify precisely the realism of the simulation results for specific power, because specific power is strongly dependent on
TIT and the value of TIT is never precisely known for real machines. Further, the
results are somewhat dependent on fuel composition. Examples of heavy duty gas
turbines, with TIT around 1350 °C, are: Siemens SGT5-4000F producing 292 MW
at r = 18 with 685 kg/s exhaust mass flow and TOT = 577 °C (the specific power is
426.3 kJ/kg); Mitsubishi M701F, producing 312 MW at r = 17, with 720 kg/s exhaust
mass flow and TOT = 597 °C (the specific power is 433.3 kJ/kg). An example of an
aero-derivative turbine is General Electric LM6000 producing 42.9 MW at r = 29.5,
11.3 Performance of Simple-Cycle Power Gas Turbines
(results not shown). This is mainly due to the higher fuel-air ratio as a consequence
of the lower heating value 43 MJ/kg.
The efficiency results of Fig. 11.14 are obtained with cooling air extracted from
the compressor, as in most machines. Further, it is assumed that there is no precooling of this air by an external heat exchanger before supplying it to the turbine.
In some machines such pre-cooling is implemented. Cooling may also be realised
with an externally provided fluid. In particular, machines exist with steam cooling.
The choice for this cooling fluid comes from applications in combined-cycle systems (see Sect. 11.4.4). The heat capacity of steam is higher than that of air, which,
principally, lowers the necessary cooling flow rates. Further, the heat absorbed by
the steam may be used in the steam cycle. However, there are some drawbacks
with steam too. Steam has to be led into the machine, but also led out of it. This is
possible, but is somewhat complicated for rotor parts. Further, film cooling cannot
be used. This increases the heat transfer between the hot gas in the turbine and the
blades and the vanes ( κ = 0.20 J/kJK instead of 0.15 J/kJK) and decreases the heat
exchanger effectiveness of the blades and the vanes ( ε = 0.7 instead of 1). Simulations with cooling with an external fluid have as a result that the global efficiency
of the simple cycle is slightly lower than with cooling by air from the compressor
(results not shown). But these are results assuming that the heat transferred to the
cooling fluid is useless. Taking into account that the heat may contribute to a second
cycle, one may conclude that there is a small advantage with steam cooling in a
combined-cycle application. We will come back to closed loop steam cooling with
the discussion of combined cycles in Sect. 11.4.4.
Figures 11.15, 11.16, 11.17 and 11.18 show specific power, turbine outlet temperature (TOT), fuel flow rate and cooling flow rate obtained by simulation with
pure CH 4 as fuel. The results with kerosene as fuel differ slightly. For instance,
specific power is about 4–5 % lower with kerosene. The specific power is expressed
as the power per unit of mass flow rate of the exhaust gas. The increase of the specific power with the combustion temperature is obvious. The specific power has
a weak maximum as a function of pressure ratio, around pressure ratio 15–20. An
important observation for practise is that with respect to specific power, it is not
useful to choose the pressure ratio too high. Heavy duty gas turbines have typically
a pressure ratio in the order of 20. The simple-cycle efficiency (Fig. 11.14) is then
not at maximum, but these machines are typically used in combined-cycle power
stations and the optimum efficiency with combined cycles occurs at lower pressure
ratio (see Sect. 11.4.4). It is difficult to verify precisely the realism of the simulation results for specific power, because specific power is strongly dependent on
TIT and the value of TIT is never precisely known for real machines. Further, the
results are somewhat dependent on fuel composition. Examples of heavy duty gas
turbines, with TIT around 1350 °C, are: Siemens SGT5-4000F producing 292 MW
at r = 18 with 685 kg/s exhaust mass flow and TOT = 577 °C (the specific power is
426.3 kJ/kg); Mitsubishi M701F, producing 312 MW at r = 17, with 720 kg/s exhaust
mass flow and TOT = 597 °C (the specific power is 433.3 kJ/kg). An example of an
aero-derivative turbine is General Electric LM6000 producing 42.9 MW at r = 29.5,