Turbine plant efficiency

On the other hand, second law analysis of a Brayton cycle (Example 5 Section 23.6.5) reveals a high exergy content of the turbine exhaust. For this reason, great improvements in gas turbine plant efficiency can be achieved by recovering exergy in the turbine outlet with either recuperation (Example 7 Section 23.6.7) or through the use of a heat recovery boiler used to supply a secondary steam cycle in a cogeneration plant (Example 6 Sechon 23.6.6). 

Assessments of control, operabiHty and part load performance of MHD—steam plants are discussed elsewhere (rl44 and rl45). Analyses have shown that relatively high plant efficiency can be maintained at part load, by reduction of fuel input, mass flow, and MHD combustor pressure. In order to achieve efficient part load operation the steam temperature to the turbine must be maintained. This is accompHshed by the use of flue gas recirculation in the heat recovery furnace at load conditions less than about 75% of fiiU load. 

The next generation of gas turbine-based, combined-cycle power plants, under constmction in many parts of the world, is to feature net plant efficiencies in the 60% range based on LHV of fuel input. These faciUties, scheduled for start-up in the latter 1990s, are anchored by large gas turbines capable of simple-cycle efficiencies >40% LHV in some cases. To develop these machines, manufacturers have scaled up and improved upon designs that have already proved to be highly rehable. 

At least two manufacturers have developed and installed machines rated to produce more than 210 MW of electricity in the simple-cycle mode. In both cases, the machines were designed and manufactured through cooperative ventures between two or more international gas turbine developers. One 50-Hz unit, first installed as a peaking power faciUty in France, is rated for a gross output of 212 MW and a net simple-cycle efficiency of 34.2% for natural-gas firing. When integrated into an enhanced three-pressure, combined-cycle with reheat, net plant efficiencies in excess of 54% reportedly can be achieved. 

A comparison of the effect of the various cycles on the overall thermal efficiency is shown in Fig. 29-40. The most effective cycle is the Brayton-Ranidne (combined) cycle. This cycle has tremendous potential in power plants and in the process industries where steam turbines are in use in many areas. The initial cost of the combined cycle is between 800- 1200 per kW while that of a simple cycle is about 300- 600 per kW. Repowering of existing steam plants by adding gas turbines can improve tne over plant efficiency of an existing steam turbine plant by as much as 3 to 4 percentage points. 

The Fossil Power Plants of the 1990s and into the early part of the new millennium will be the Combined Cycle Power Plants, with the gas turbine as being the centerpiece of the plant. It is estimated that between 1997-2006 there will be an addition of 147.7 GW of power. These plants have replaced the large Steam Turbine Plants, which were the main fossil power plants through the 1980s. The Combined Cycle Power Plant is not new in concept, since some have been in operation since the midl950s. These plants came into their own with the new high capacity and efficiency gas turbines. 

The new marketplace of energy conversion will have many new and novel concepts in combined cycle power plants. Figure 1-1 shows the heat rates of these plants, present and future, and Figure 1-2 shows the efficiencies of the same plants. The plants referenced are the Simple Cycle Gas Turbine (SCGT) with firing temperatures of 2400 °F (1315 °C), Recuperative Gas Turbine (RGT), the Steam Turbine Plant (ST), the Combined Cycle Power Plant (CCPP), and the Advanced Combined Cycle Power Plants (ACCP) such as combined cycle power plants using Advanced Gas Turbine Cycles, and finally the ITybrid Power Plants (HPP). 

The overall efficiency of the entire gas turbine plant, including the cyclic gas turbine power plant (within T) and the heating device (within Z), is given by... 

For an open circuit (non-cyclic) gas turbine plant (Fig. 1.3) a different criterion of performance is sometimes used—the rational efficiency (tjr). This is defined as the ratio of the actual work output to the maximum (reversible) work output that can be achieved between the reactants, each at pressure (po) and temperature (To) of the environment, and products each at the same po. Tq. Thus... 

In Chapter 1, the gas turbine plant was considered briefly in relation to an ideal plant based on the Carnot cycle. From the simple analysis in Section 1.4, it was explained that the closed cycle gas turbine failed to match the Carnot plant in thermal efficiency because of... 

The objective of the gas turbine designer is to make all the proces.ses in the plant as near to reversible as possible, i.e. to reduce the irreversibilities, both internal and external, and hence to obtain higher thermal efficiency (in a closed cycle gas turbine plant) or higher overall efficiency (in an open gas turbine plant). The concepts of availability and exergy may be used to determine the location and magnitudes of the irreversibilities. 

Fig. 2.9 illustrates this approach of tracing exergy through a plant. The various terms in Eq. (2.49) are shown for an irreversible open gas turbine plant based on the JB cycle. The compressor pressure ratio is 12 1, the ratio of maximum to inlet temperature is 5 1 (T,nax = 1450 K with To = 290 K), the compressor and turbine polytropic efficiencies are... 

Subsequently, in Chapter 4, we deal with cycles in which the turbines are cooled. The basic thermodynamics of turbine cooling, and its effect on plant efficiency, are considered. In Chapter 5, some detailed calculations of the performance of gas turbines with cooling are presented. 

A set of calculations using real gas tables illustrates the performance of the several types of gas turbine plants discussed previously, the [CBT]ig, [CBTX]ig, [CBTBTX]ig, [CICBTXIig and [CICBTBTX]ig plants. Fig. 3.15 shows the overall efficiency of the five plants, plotted against the overall pressure ratio (r) for = 1200°C. These calculations have been made with assumptions similar to those made for Figs. 3.13 and 3.14. In addition (where applicable), equal pressure ratios are assumed in the LP and HP turbomachinery, reheating is set to the maximum temperature and the heat exchanger effectiveness is 0.75. 

Fig. 6.4 then shows a more complete calculation of plant efficiency for varying S. The optimum condition of maximum efficiency is reached at 5 = 0.208. The picture changes for a gas turbine with a higher pressure ratio, for which the increase to maximum efficiency is less, as is the optimum value of S 2]. 

In the search for higher plant thermal efficiency, the simplicity of the two basic STIG and EGT cycles, as described by Frutschi and Plancherel, has to some extent been lost in the substantial modifications described above. But there have been other less complex proposals for water injection into the simple unrecuperated open cycle gas turbine one simply involves water injection at entry to the compressor, and is usually known as inlet fog boosting (IFB) the other involves the front part of an RWI cycle, i.e. water injection in an evaporative intercooler, usually in a high pressure ratio aero-derivative gas turbine plant. 

For a mass flow of air (Mg) to the compressor of the gas turbine plant, a mass flow Mf of fuel (of specific enthalpy hfo) is supplied to the two combustion chambers (Mf = (Mf)ff + (Mf)f). The overall efficiency of the combined plant is then... 

The interaction between the gas turbine plant and the steam cycle is complex, and has been the subject of much detailed work by many authors [5-8]. A detailed account of some of these parametric studies can be found in Ref. [1], and hence they are not discussed here. Instead, we first illustrate how the efficiency of the simplest CCGT plant may be calculated. Subsequently, we summarise the important features of the more complex combined cycles. 

Even for this simplest CCGT plant, iterations on such a calculation are required, with various values of p, in order to meet the requirements set on T, the steam turbine entry temperature, and 7s (the calculated value of 7s has to be such that the dewpoint temperature of the gas (7jp) is below the economiser water entry temperature (7b) and that may not be achievable). But with the ratio /i satisfactorily determined, the work output from the lower cycle Wl can be estimated and the combined plant efficiency obtained from... 

With the fuel energy input known from the calculation of the gas turbine plant performance, F = Aff[CV]o, the combined plant efficiency is determined as... 

Most modem CCGT plants use open air cooling in the front part of the gas turbine. An exception is the GE MS9001H plant which utilises the existence of the lower steam plant to introduce steam cooling of the gas turbine. This reduces the difference between the combustion temperature T ot and the rotor inlet temperature The effect of this on the overall combined plant efficiency is discussed in Ref. [1] where it is suggested that any advantage is small. 

But for power station applications, the thermal efficiency is not the only measure of the performance of a plant. While a new type of plant may involve some reduction in running costs due to improved thermal efficiency, it may also involve additional capital costs. The cost of electricity produced is the crucial criterion within the overall economics, and this depends not only on the thermal efficiency and capital costs, but also on the price of fuel, operational and maintenance costs, and the taxes imposed. Yet another factor, which has recently become important, is the production by gas turbine plants of greenhouse gases (mainly carbon dioxide) which contribute to global warming. Many countries are now considering the imposition of a special tax on the amount of CO2 produced by a power plant, and this may adversely affect the economics. So consideration of a new plant in future will involve not only the factors listed above but also the amount of CO2 produced per unit of electricity together with the extra taxes that may have to be paid. 

We also give calculations of the performance of some of these various gas turbine plants. Comparison between such calculations is often difficult, even spot calculations at a single condition with state points specified in the cycle, because of the thermodynamic assumptions that have to be made (e.g. how closely conditions in a chemical reformer approach equilibrium). Performance calculations by different inventors/authors are also dependent upon assumed levels of component performance such as turbomachinery polytropic efficiency, required turbine cooling air flows and heat exchanger effectiveness if these are not identical in the cases compared then such comparisons of overall performance become invalid. However, we attempt to provide some performance calculations where appropriate in the rest of the chapter. 

After allowing for the performance penalties arising from the CO2 removal, Lozza and Chiesa estimated an efficiency of 46.1%, for a maximum gas turbine temperature of 1641 K and a pressure ratio of 15 (compared with the basic CCGT plant efficiency of 56.1%). They concluded that the plant cannot compete, in terms of electricity price, with a semi-closed combined cycle with CO2 removal (Cycle A2). 

Newby et al. found that increasing the PO turbine pressure resulted in higher steam flow (for a given pinch point temperature difference in the HRSG), increased PO turbine power and overall plant efficiency. However, at the highest pressure of 100 bar attempts to increase the steam flow further resulted in incomplete combustion in the main combustor and the overall thermal efficiency did not increase substantially at this pressure level. 

A gas turbine plant with an overall efficiency t]cq = 0.25 matching a heat load Acc, = 2.25 is again considered as the basic CHP plant also implied is a non-useful heat rejection ratio (Cnu)cg( cg = [1 ( cg)( g + 1)1 =. 3/16. For FESR calculations, we again take the conventional plant efficiency as 0.4 and the conventional boiler efficiency as 0.9. At the fully matched condition the.se assumptions previously led to EUF = 0.8125 and FESR = 0.2. 

Wunsch, A. (1985), Highest efficiencies possible by converting gas turbine plants. Brown Boveri Review 1, 455-4.56. 

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