Steam-power cycles with adiabatic vaporization
Abstract Generated abstract
The paper argues that further improvement of conventional steam power plants through very high and supercritical steam parameters faces technical and economic limits, while low pressure installations remain inefficient despite their operational simplicity. It proposes steam power cycles with adiabatic steam generation, using a high pressure water heating boiler and an evaporator generator to separate high pressure liquid heating from lower pressure steam production and expansion. Calculation formulas and cycle characteristics are presented, showing that high thermal efficiency can be maintained at relatively low steam pressures and depends notably on the internal efficiency of the evaporator generator turbine. The author suggests that this arrangement could combine advantages of low and high pressure systems, support higher steam temperatures, improve condenser and boiler performance, and be applicable to nuclear and combined steam gas power plants.
Full Text
Parosilovye Cycles with Adiabatic Steam Formation
S. V. Polozhii
(Presented by Prof. I. N. Butakov, Doctor of Technical Sciences)
Power engineering, and its leading branch—thermal power engineering—are the basis of the country’s electrification. More than 80% of electric power in our country is generated at thermal power plants. This gives exceptional importance to technical progress in thermal power engineering, requiring technical solutions that ensure the most rational development of thermal power plants: reduction of specific capital investments and operating costs, shortening of the time required for construction and commissioning of new capacity, and achievement of high thermal economy and reliability of the country’s power supply. The current practice of solving this problem within the framework of the existing steam-power cycle by introducing superhigh and supercritical steam pressures cannot be considered successful. This conclusion is confirmed by the great difficulties in introducing superhigh and supercritical steam pressures, associated with the need to solve numerous and complex problems caused by high requirements for metal, water, and steam, the need to apply complex cycles and thermal schemes, the presence of special conditions, etc. The difficulties of mastering such installations have led to the fact that in capitalist countries, especially in the USA, which had set out on the path of introducing superhigh and supercritical steam parameters earlier than our country, there has been a retreat from supercritical steam parameters [1—4]. In these countries, the parameters of superheated steam have been reduced for newly constructed superpowerful thermal power plants: pressure from 350—300 bar to 230 bar, temperature from 665—600°C to 530°C. The principal capacity being introduced in the USA has a superheat temperature of only 535—540°C. Installations with temperatures of 565—658°C have been reduced almost 12-fold. In the FRG, the main capacity being commissioned consists of installations with temperatures of 525°C and lower—65%; about 25% of installations have temperatures of 535—540°C, and only 6%—above 540°C. The introduction of supercritical steam parameters in unit installations of enormous capacity abroad has, in essence, turned out to be a huge industrial experiment, on the basis of which it has practically been established that increasing the thermal economy of thermal power plants through the use of supercritical pressures and superhigh steam temperatures is inexpedient. The expected techno-economic effect from supercritical steam parameters has not been fully achieved in any large installation. The practical fuel savings achieved in operation as a result of using supercritical steam pressure with double
with intermediate superheating after 3–4 years of operation in most cases amounts to only 30–50% of the design value, despite the exceptionally high cost of the equipment of such stations. Industrial thermal power plants and other types of low-capacity steam-power installations are in an even more difficult condition. Such installations, owing to the features of the modern steam-power cycle, have low steam parameters and the low thermal economy associated with them, are at a low technical level of development, and have no prospects for improving their economy. The prolonged existence of this type of steam-power installation and their large number are at present leading to an enormous overexpenditure of fuel, metal, and other material assets [5]. In 1961 such thermal power plants overconsumed about 3 million tons of reference fuel. The annual relative loss to the national economy from the operation of all small power stations, in comparison with the operating costs of large stations, amounts to no less than 1.5–2 billion rubles.
Such a condition and development of modern thermal power engineering is due to the existing steam-power cycle, the use of isobaric steam generation, the physical and thermodynamic properties of steam, and the contradictions that arise between the thermodynamic efficiency of the cycle $\eta_t$ and the degree of utilization of the working fluid in the steam turbine $\eta_{0i}$, as well as the technical-economic contradictions in achieving an expedient level of utilization of fuel energy. At the same time, each level of the initial parameters of the modern steam-power cycle has both positive and negative properties, and it is impossible to combine in one installation the positive aspects of low and ultrahigh steam pressures.
A positive property of steam-power installations with low steam pressure is the high efficiency of operation of low-capacity steam turbines, owing to the low pressure, the considerable volumetric steam flow rates, and the low moisture content at the end of expansion. This also explains the thin-walled design of power equipment, the low requirements for the quality of metal, water, and steam, the relative uniformity and cheapness of the metal in all elements of the steam-power installation, the simplicity of the thermal schemes, the power equipment, and so forth. But, on the other hand, low steam pressures lead to a low thermodynamic efficiency of the cycle $\eta_t$, to low values of the resulting efficiency $t$ of the installation $\eta_t \cdot \eta_{0i}$, to large specific consumptions of fuel, steam, and water, and, because of the low working capacity $\kappa z$ of the steam, cause a small unit capacity of steam turbines and the associated decentralization of electricity production at stations of small capacity. This also explains the high cost of electricity production, the large initial and current capital investments, and the low technical level of development of such steam-power installations. The use of ultrahigh and supercritical steam pressures, on the contrary, makes it possible to have a high thermodynamic efficiency of the steam-power cycle $\eta_t$, a high working capacity $\kappa z$ of the steam, to create turbines of enormous capacity, to carry out concentration and centralization of electric-power production, and in installations of enormous capacity to obtain a high resulting efficiency $\eta_t \cdot \eta_{0i}$. But, on the other hand, the physical and thermodynamic properties of steam deteriorate sharply at ultrahigh and supercritical pressures; the conditions of steam operation in the steam turbine deteriorate, to the point where steam turbines cannot operate at ultrahigh pressure because of excessive steam moisture and large heat losses in the pre-
of the flow part. Only by using intermediate reheating of the steam and turbines of enormous capacity is it possible to achieve reliable and efficient operation of superhigh-pressure turbines. The requirements on the quality of the metal increase sharply. It becomes necessary to design power equipment from high-alloy steels with large wall thicknesses and a high consumption of metal. The thermal scheme becomes extremely complicated and the reliability of operation of the station is reduced. Considerable time is required for mastering and introducing such equipment, for organizing large-scale research work, etc.
A rational steam-power installation and the cycles of such installations must not have the basic shortcomings of modern steam-power installations of both low and superhigh steam pressures; they must combine the advantages of such installations and overcome the difficulties in the development of powerful block-type superhigh-pressure GRES units, as well as the stagnation in the development of steam-power installations of small capacity. Such a steam-power installation must preserve the favorable properties of low- and superhigh-pressure steam, ensure high thermal economy of the cycle and efficient operation of the steam in a turbine of both large and small capacity. The solution of this problem is possible by applying, in steam-power installations, adiabatic steam generation, proposed and developed at the Tomsk Polytechnic Institute [6] (Certificate of registration and priority of the Committee for Inventions and Discoveries of the USSR No. 19405 for 1960 and No. 29529 for 1962).
Fig. 1. Theoretical cycle of a steam-power installation with adiabatic steam generation in an \(i\)—\(s\) diagram, taking into account the work of the circulating and feed pumps of the water-heating boiler. \(h_y\) is the work of the circulating pump per 1 kg of water; \(G\) kg of water participates in the cycle. \(h_n\) is the work of the feed pump per kg of steam in the cycle.
A steam-power installation with adiabatic steam generation consists of a superhigh-pressure water-heating boiler, an evaporator-generator, structurally made in the form of a single-disk turbine, in the evaporating elements of which steam is produced and the kinetic energy of the flow, a steam superheater and a low-pressure steam turbine, the feed and circulating pumps of the water-heating boiler, and regenerative feed-water heaters [6]. The thermodynamic cycle of steam-power installations with adiabatic steam generation in an \(i\)—\(s\) diagram is shown in Fig. 1.
The calculation of cycles with adiabatic steam generation is carried out according to the formulas of Table 1. The calculation of the thermal scheme of the station and the determination of thermal indices, steam consumption, fuel consumption, etc., are carried out in the same way,
Calculation formulas for cycles of steam-power plants
Table 1
| Quantity | Theoretical | With allowance for losses |
|---|---|---|
| Amount of saturated steam per kg of water | $$x_s=\frac{(S'_0-S'_1)\,T_{\mathrm{H}}}{r_1}$$ | $$x=x_s+\frac{(1-\eta^{\mathrm{ng}}_{0i})}{t_s}$$ |
| Amount of high-pressure water per kg of steam | $$G_s=\frac{1}{x_s}=\frac{r_1}{(S'_0-S'_1)\,T_{\mathrm{H}}}$$ | $$G=\frac{1}{x}$$ |
| Work of 1 kg of water in the steam generator | $$l_s=i'_0-i'_1-(S'_0-S'_1)\,T_{\mathrm{H}}$$ | $$l=t_s\cdot\eta^{\mathrm{ng}}_{0i}$$ |
| Work of the high-pressure steam per kg of steam in the cycle | $$L_s=G_s\cdot l_s$$ | $$L_{\mathrm{M}}=G\cdot l$$ |
| Work of 1 kg of steam in the cycle | $$l_{\mathrm{sp}}=i_1-i_2$$ | $$l_{\mathrm{p}}=l_{\mathrm{sp}}\cdot\eta^{\mathrm{pt}}_{0i}$$ |
| Total work of the cycle | $$L_0=L_s+l_{\mathrm{sp}}$$ | $$L=L_{\mathrm{M}}+l_{\mathrm{p}}$$ |
| Heat consumption for heating the condensate steam to \(t_{\mathrm{n}1}\) | $$Q_1=i'_1-i'_2$$ | $$Q_1=(i'_1-i'_2)$$ |
| Heat consumption for high-pressure water up to \(t_{\mathrm{n}0}\) | $$Q_2=G_s(i'_0-i'_1)$$ | $$Q'_2=G\cdot(i'_0-i'_1)$$ |
| Total heat consumption for water heating | $$Q_{\mathrm{B}}=Q_1+Q_2$$ | $$Q'_{\mathrm{B}}=Q_1+Q'_2$$ |
| Heat of superheating of 1 kg of steam | $$Q_{\mathrm{p}}=i_1-i'_1$$ | $$Q_{\mathrm{p}}=(i_1-i'_1)$$ |
| Total heat consumption in the cycle | $$Q=Q_{\mathrm{B}}+Q_{\mathrm{p}}$$ | $$Q'=Q'_{\mathrm{B}}+Q_{\mathrm{p}}$$ |
| Thermal efficiency, “gross” | $$\eta^{\mathrm{br}}_t=\frac{L_0}{Q}$$ | $$\eta^{\mathrm{br}}_t=\frac{L}{Q'}$$ |
| Thermal efficiency, “net,” accounting for pump work | $$\eta^{\mathrm{n}}_t=\frac{L_0-h_{\mathrm{n}}}{Q-h_{\mathrm{n}}}$$ | $$\eta^{\mathrm{n}}_t=\frac{L-h_{\mathrm{in}}}{Q'-h_{\mathrm{in}}}$$ |
| Specific steam consumption of the plant | $$d_0=\frac{1}{L_0}$$ | $$d_{\mathrm{e}}=\frac{1}{L\cdot\eta_{\mathrm{m}}\cdot\eta_{\mathrm{gen}}}$$ |
| Specific heat consumption | $$q_0=d_0\cdot Q$$ | $$q_{\mathrm{e}}=d_{\mathrm{e}}\cdot Q'$$ |
Table 1 (continued)
| Name of quantity | Theoretical | Taking losses into account |
|---|---|---|
| The same with regenerative heating of feedwater | $$d_0^{\mathrm p}=\frac{1}{L_s+l_{sn}\left[1-\sum a\,\dfrac{i_{0r}-i_2}{i_1-i_2}\right]}$$ | $$d_e^{\mathrm p}=\frac{1}{\left\{L^{\mathrm{in}}+l_n\left[1+\sum a\,\dfrac{i_n-i_2}{i_1-i_2}\right]\right\}\eta_m\eta_{\mathrm{gen}}}$$ |
| Share of energy production by the IG | $$\beta=\frac{L_s}{L_0}$$ | $$\beta'=\frac{L^{\mathrm{in}}}{L_0}$$ |
| Work of the circulating pump per kg of steam | $$h_y=\frac{i_{\mathrm{in}}-i'_1}{x_s}$$ | $$h'_y=\frac{i_{\mathrm{in}}-i'_1}{x\cdot\eta_n}$$ |
| Work of the feed pump | $$h_n=i_{2n}-i_2'$$ | $$h'_n=\frac{h_n}{\eta_n}$$ |
| Total energy consumption for pumps | $$h_{\mathrm n}=h_y+h_n$$ | $$h'_{\mathrm n}=h'_y+h'_n$$ |
| Specific steam consumption by the unit | $$d_0=\frac{1}{L_0}$$ | $$d_e=\frac{1}{L\cdot\eta_m\eta_{\mathrm{gen}}}$$ |
| Specific heat consumption | $$q_0=d_0\cdot Q$$ | $$q_e=d_e\cdot Q'$$ |
| Specific steam consumption with regenerative heating of feedwater | $$d_0^{\mathrm p}=\frac{1}{l_s+l_{sn}\left[1-\sum a\,\dfrac{i_{0r}-i_2}{i_1-i_2}\right]}$$ | $$d_e^{\mathrm p}=\frac{1}{\left\{L^{\mathrm{in}}+l_n\left[1-\sum a\,\dfrac{i_{0r}-i_2}{i_1-i_2}\right]\right\}\eta_m\eta_{\mathrm{gen}}}$$ |
| Specific heat consumption with regenerative heating of feedwater | $$q_0^{\mathrm p}=d_0^{\mathrm p}\cdot Q^{\mathrm p}$$ | $$q_e^{\mathrm p}=\frac{d_e^{\mathrm p}\cdot Q^{\mathrm p}}{\eta_{\mathrm{KU}}\eta_{\mathrm{pot}}}$$ |
| Share of energy production in the evaporator-generator turbine | $$\beta=\frac{L_s}{L_0}$$ | $$\beta'=\eta_{\mathrm{KU}}\eta_{\mathrm{pot}}\frac{L^{\mathrm{in}}}{L}$$ |
| Work of the circulating pump per kg of steam in the cycle | $$h_y=\frac{i_{\mathrm{in}}-i'_1}{x_s}$$ | $$h'_y=\frac{i_{\mathrm{in}}-i'_1}{x\cdot\eta_n}$$ |
| Work of the feed pump | $$h_n=i_{2n}-i'_2$$ | $$h'_n=\frac{h_n}{\eta_n}$$ |
| Total energy consumption for pumps | $$h_{\mathrm n}=h_y+h_n$$ | $$h'_{\mathrm n}=h'_y+h'_n$$ |
as for ordinary steam-power plants. Figure 2 gives the thermal and energy characteristics of the theoretical cycle with adiabatic steam generation as a function of the pressure of the steam phase of the cycle. The thermal efficiency of the cycle, with unchanged pressure of the heated water, final pressure, and temperature of the superheated steam, depends little on the initial steam pressure (Fig. 2—the ratio of the data of curves 3 to 6). For example, at a steam pressure in the evaporator-generator of 40 bar the thermal efficiency of such a cycle is 46%, while at a steam pressure in the EG of 160 bar, at the same superheat temperature, the thermal efficiency is 46.5%. The high value of the thermal efficiency of cycles with adiabatic steam generation at low steam pressure is a major advantage of such steam-power plants.
In actual cycles with adiabatic steam generation, the useful work, the amount of steam formed, the thermal efficiency, and other characteristics depend on the internal relative efficiency of the evaporator-generator turbine. The presence of internal losses in the evaporator-generator turbine reduces the useful work and increases the amount of steam obtained from 1 kg of liquid. The energy expenditure for the circulation pump of the hot-water boiler also depends on the relative internal efficiency of the EG turbine (Table 1). As the pressure of the steam phase of the cycle increases, the total energy consumption for the hot-water boiler pumps falls somewhat. Figure 3 gives the value of the “net” efficiency of steam-power plants with adiabatic steam generation as a function of the efficiency of the evaporator-generator turbines. The actual value of the efficiency of evaporator-generator turbines, according to theoretical and design studies, is 0.64–0.79.
Fig. 2. Heat and work of the cycle with adiabatic steam generation as a function of the pressure in the evaporator-generator:
1—work in the evaporator-generator,
2—work of the steam in the cycle,
3—total work of the cycle,
4—heat of liquid heating,
5—heat of steam superheating,
6—total heat input in the cycle.
Initial water pressure \(P_0=220\) bar, steam pressure in the condenser \(P_2=0.039\) bar, steam superheat temperature \(t_1=600^\circ\mathrm{C}\).
A different organization of the technological process for producing steam and energy in steam-power plants with adiabatic steam generation—the presence of two pressures of the working fluid, the production of saturated steam and work in the evaporator-generator, the absence of steam generation in heated heating surfaces, and the ultra-high pressure of the water at the...
low temperature, a low vapor pressure at high temperature, and a high value of the thermal efficiency at low vapor pressure—makes it possible to combine the positive properties of modern low- and superhigh-pressure steam-power installations in a single installation, to eliminate the shortcomings of installations with both low and superhigh steam pressures. In addition, it becomes possible to use a very high steam temperature in the cycle, comparable with the temperature in gas-turbine installations, and also to make economically expedient use of
Fig. 3. Influence of the relative internal efficiency of the evaporator-regenerator turbine on the net efficiency of the station, taking into account the energy expenditure for the circulating and feed pumps of the water-heating boiler, and comparison of them with the efficiencies of modern steam-power installations.
1–5—installations with adiabatic steam generation \((P_0 = 220\ \text{bar},\ P_2 = 0.039\ \text{bar})\).
1—steam pressure 35 bar, temperature 440°C; 2—steam pressure 35 bar, temperature 550°C; 3—steam pressure 35 bar, temperature 600°C; 4—steam pressure 80 bar, temperature 550°C; 5—steam pressure 80 bar, temperature 600°C; 6–8—modern installations.
6—steam pressure 35 bar, temperature 440°C; 7—steam pressure 90 bar, temperature 500°C; 8—steam pressure 180 bar, temperature 550°C.
Assumed: relative internal efficiency of steam turbines—0.85; mechanical and generator efficiency—0.96; boiler efficiency—0.85.
a deep vacuum in the condensers of steam turbines and to build water-heating boilers with a low temperature of the exhaust gases, and on this basis to raise still further the thermal economy of steam-power installations with adiabatic steam generation.
The use of adiabatic steam generation is also progressive for nuclear and steam-gas power installations.
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