An Encyclopedia of the History of Technology part 32 pps

PART TWO: POWER AND ENGINEERING 292 Figure 5.6: Schematic arrangement of the basic steam turbine types. (a) Simple impulse turbine: de Laval (c.1883). Key: 1: stationary nozzles; 2: rotating blades; 3: wheel; 4: rotating shaft. The lower diagram shows the change in pressure (P) and velocity (V) of the steam as it passes through the turbine. (b) Reaction Turbine: Parsons (c.1884). Key: 1: Nozzles and guide vanes; 2: rotating blades; 3: rotating drum; 4: rotating shaft. The conical drum counteracts the decrease in density of the steam as it passes through the turbine. This arrangement was not used by Parsons who employed a limited number of step changes in drum diameter. The lower diagram shows the change in pressure (P) and velocity (V) of the steam as it passes through the turbine. STEAM AND INTERNAL COMBUSTION ENGINES 293 (c) Pressure-compounded impulse turbine: Rateau (c.1900). Key: 1: nozzles; 2: rotating blades; 3: nozzles in diaphragm; 4: rotating wheels; 5: rotating shaft. The lower diagram shows the change in pressure (P) and velocity (V) of the steam as it passes through the turbine. (d) Velocity-compounded impulse turbine: Curtis (1896). Key: 1: nozzles; 2: rotating blades; 3: stationary guides; 4: rotating wheels; 5: rotating shaft. The lower diagram shows the change in pressure (P) and velocity (V) of the steam as it passes through the turbine. Reproduced from E.F.C.Somerscales, ‘Two Historic Prime Movers’, ASME Paper number 84-WA/HH-2 (1984). impulse turbine that is now commonly known by his name. The principle of this turbine is illustrated in Figure 5.6 (c). As in the Parsons turbine, the pressure drop is divided between a number of stages, thereby limiting the steam speed. This type of turbine avoids the leakage problem of the Parsons turbine by only using each stage pressure drop in the stationary nozzles. The Rateau turbine is classed as a pressure-compounded impulse turbine. The Rateau turbine differed from the Parsons turbine in another important respect. In the latter the moving blades are attached to the periphery of a rotating drum, with the seals between the stages at its outer surface and at the outer ends of the rotating blades. In the Rateau turbine the blades are mounted on the periphery of a disc carried on the turbine shaft, so there is only one interstage seal and it is where the rotating shaft passes through the diaphragm separating the stages (see also Figure 5.6 (c)). This type of construction has two advantages over the system used in the Parsons turbine: the leakage area, and, hence the leakage flow, is smaller because the seal is at the shaft, which has a smaller diameter than the Parsons drum; and a shaft seal involves parts with heavier dimensions than a ring of stationary blades, so it can be made more effective. The disc construction of the Rateau turbine would, because of the pressure drop across the moving blades, result in an unacceptably large axial thrust on the shaft bearings if applied to the Parsons turbine. The axial thrust that does exist in the latter is smaller because it only acts over the blade area, and is accommodated by thrust balancing cylinders, connected externally to points of higher steam pressure, at the high pressure end of the rotor. Because the solid construction of the drum is better suited to the high temperatures encountered in the high pressure stages of modern turbines, the disc- on-shaft construction has been effectively abandoned in these sections and the blades are mounted on vestigial discs machined out of a forged cylindrical rotor. Although the Rateau turbine mitigated the leakage problem of the Parsons turbine, it still required a long shaft to accommodate the large number of pressure compounded stages. In 1896 the American C.G.Curtis patented two concepts that resulted in a substantial reduction in the overall length of the turbine. The first of these was the recovery of the velocity PART TWO: POWER AND ENGINEERING 294 Figure 5.7: Cross section of the first Parsons steam turbine. This was supplied with saturated steam at 5.5bar (80psig), ran at 18,000rpm and produced between 4 kW and 7.5 kW at the generator terminals. Steam was admitted through the inlet (A) and flowed axially to right and left through alternating rows of moving and fixed blades (shown schematically). The steam exhausted into the passages C and D which were cast in the turbine casing. The exhaust pressure (about atmospheric) was maintained by a steam ejector (B). The diagram also shows the governor (designed because available rotary governors could not withstand the high shaft speed), which consisted of a small centrifugal fan, located just to the left of the item lettered C in the diagram, rotating at the turbine shaft speed, and drawing air from one side of a diaphragm (located just above the item lettered C). The motion of the diaphragm, as the vacuum varied with turbine speed, was conveyed by the link (E) to the governor valve controlling the steam admission. The sensitivity of this governor was increased by sensing the generator voltage (actually the magnetic field strength, which is proportional to the voltage) and bleeding more or less air into the vacuum system. In the figure, F is the soft-iron field sensor mounted on the dynamo pole pieces, and coupled to this STEAM AND INTERNAL COMBUSTION ENGINES 295 is the spring loaded brass arm (G) that controls the flow of the bleed air. Forced lubrication using a screw pump (H) was used. The arrows show the direction of oil flow. The small diagram at the foot of the figure shows one of the bearings that were designed to allow some transverse movement of the shaft, due to small residual imbalances. The shaft ran in a long thin sleeve, surrounded by a large number of washers alternately fitting the sleeve and the casing. The washers were held in contact with each other by the pressure of a short helical spring surrounding the sleeve near its end and tightened up by a nut threaded on the sleeve. The bearing oil was supplied by the screw pump. Reproduced with permission from W.G.S.Scaife, in First Parsons International Turbine Conference (Parsons Press, Dublin and London, 1984). energy in the steam jet leaving a converging-diverging nozzle in several rows of moving blades with stationary turning vanes between each row (see Figure 5.6 (d)). This multi-row de Laval turbine is known as a velocity-compounded impulse turbine. The second arrangement patented by Curtis, which was his special contribution to turbine engineering, was to combine velocity- and pressure-compounding (Figure 9). To obtain the necessary funds to support the development of his turbine, Curtis sold most of his patent rights to the General Electric Company of Schenectady, New York, in 1897. However, a satisfactory working machine was not produced until 1902. The Curtis turbine, like the Rateau turbine, was a disc turbine. Eventually, as discussed below, the Curtis and Rateau turbines were combined, so that later forms of impulse turbines that have derived from these two types generally tend to be of the disc construction, or at least have vestigial discs machined from a drum. By 1900 the four basic types of steam turbines had been developed into practical machines. Although there was no general and immediate tendency to adopt this prime mover, in spite of its obvious advantages, by the end of the following decade the steam turbine had established itself in the electric power generation industry, and had also provided a number of convincing demonstrations of its usefulness as a marine power plant. 1900–1910: turbines take over During the first ten years of this century the application of steam turbines to both electric power generation and marine propulsion spread very rapidly in consequence of a number of influential demonstrations of its capabilities. The first of these was in 1891 when a measured steam consumption of 12.2kg/ kWhr (27lb/kWhr) was reported for one of the 100kW condensing (the first such) machines installed at Cambridge. This was a record for steam turbines and equalled the performance of the best simple expansion reciprocating steam engines. PART TWO: POWER AND ENGINEERING 296 Figure 5.8: Vertical section through the 5000kW vertical Curtis steam turbine supplied by the General Electric Co. to the Commonwealth Electric Co. for installation in the Fisk Street Station, Chicago in 1903. Steam inlet conditions: 11.9bar (175psig), 271°C (520°F); exhaust: 0.068bar (28in Hg vacuum); minimum steam consumption 6.81kg/(kW/hr) (15.0lb mass/(kW/hr)). The turbine had two pressure compounded stages consisting of a nozzle, three rows of stationary guide and four rows of moving blades attached to one wheel. The vertical arrangement was discontinued by the General Electric Co. built after 1913. Reproduced with permission from C.Feldman ‘Amerikanische Dampfturbinen’, Zeitschrift des vereines deutscher Ingenieure, vol. 48 (1904), p. 1484. A second significant installation was the two 1000kW turbines, the largest built to date, that were delivered in 1900 by C.A.Parsons & Co. to the city of Elberfeld in Germany. These had a steam consumption of 9.12kg/kWhr (20.1lb/kWhr). These turbines had many of the features of the modern unit: they were two cylinder tandem compound (see below), and the condenser was placed below the level of the operating floor, immediately under the exhaust from the low pressure cylinder. Another very influential steam turbine, which was placed in service in 1903 at Chicago’s Fisk Street station (see Figure 5.8), was the third Curtis steam turbine sold by the General Electric (GE) Company. It had a power output of 5MW and was the most powerful steam turbine built up to that time. It was unusual in being arranged with its axis vertical (a type of construction abandoned by GE in 1913) and with the alternator above the turbine. Pressure-andvelocity-compounding were used, and each of the two pressure- compounded stages consisted of a row of nozzles, three rows of stationary turning vanes, and four rows of moving blades attached to one wheel. STEAM AND INTERNAL COMBUSTION ENGINES 297 1910–1920: blending of types As the patent protection on various types of turbines expired the manufacturers devised hybrid machines. One of the most significant of these combinations was the addition to the Parsons turbine at the inlet of a single velocity-compounded stage, sometimes called a control stage, in place of a number of the high pressure reaction stages. The rotor of this combined Curtis-Parsons turbine is much shorter than the rotor of the pure Parsons type of the same power output. This arrangement confines high temperatures and pressures to a shorter portion of the turbine, minimizing expansion effects due to temperature gradients, which can lead to eventual failure of the machine. Although the Curtis stage is not as efficient as the reaction stage it replaces, the minimization of leakage and improvement in reliability result in a net gain to the user of the turbine. Another technique for minimizing the turbine shaft length, which was probably first introduced into regular steam turbine design practice with a 25MW turbine built in 1913 by C.A.Parsons & Co. for the Fisk Street station in Chicago of the Commonwealth Edison Company, was the tandem compound. In this machine two rotors are arranged in two separate cylinders, with bearings in each cylinder. The steam passes in succession through the two cylinders. The rotors are coupled between the cylinders and the load is connected to one end of the shaft, usually at the low pressure cylinder. 1920–1930: increasing size The period between 1920 and 1930 was characterized by a very rapid growth in steam turbine power output; preliminary attempts to use very high steam pressures; and a large number of serious mechanical failures. Increasing power output implies increasing mass rate of flow of steam, and increasing dimensions. The largest dimensions are encountered at the exhaust from the turbine, where the steam has the largest volume, and the critical dimension in this region is the length of the last row blades, which are the longest in the turbine and, therefore, are subjected to the highest centrifugal stress. Consequently, very careful consideration must be given to the mechanical design of the rotor and the last row blades, their manufacture and the materials used, and the means of attaching the blades to the rotor. Since sound forgings for the rotor can be assured with greater certainty the smaller their size, there was a tendency during the period from 1920 to 1960 to use the disc construction in the low pressure sections of the turbines. Where the length of the last row blades could not be increased a number of alternative techniques were developed in the 1920s including Baumann multiple exhaust, due to K.Baumann of the Metropolitan-Vickers Co., multiple PART TWO: POWER AND ENGINEERING 298 exhaust flows, and multiple low pressure turbines (not used until the 1930s). Extreme forms of this arrangement, involving three or four low pressure turbines, each with a double exhaust flow, have been used on the very high output machines constructed from 1960 onward. In order to accommodate the largest outputs a combination of decreasing the turbine speed and multiple exhaust flows was used in the 1920s. This was in the form of a cross-compound turbine in which the high pressure cylinder operated at 3000 or 3600rpm and the low pressure cylinder ran at 1500 or 1800rpm, the two cylinders being coupled to separate alternators. This arrangement appears to have been first used for three 30MW turbines installed by Westinghouse in 1914 at the 74th Street station in Manhattan of the Interborough Rapid Transit Co. The cross-compound tends to be an expensive solution to the problem of building turbines with large outputs, and European machines, with their lower speeds (1500 and 3000rpm), did not employ it as extensively as turbines built in the United States. The two-speed cross-compound has not been used since the 1970s (except to accommodate very large turbines with outputs of 1000MW or more), because improved materials of construction removed the incentive for its use. The steady growth in turbine power output in the 1920s culminated in a 208MW turbine that was constructed by GE for installation in the State Line station near Chicago in 1929. This was a three-cylinder cross-compound turbine, and it remained the turbine with the world’s largest output until 1955. The application of very high inlet pressures to steam turbines was first attempted in 1925, in order to increase cycle efficiency. Three such turbines were installed at the Edgar Station of the Edison Electric Illuminating Company of Boston (now the Boston Edison Company). The first (3.15MW) had an inlet pressure of 83bar (1200psig) and the other two (10MW) operated at 97bar (1400psig). They exhausted at 24bar (350psig) into the steam line connecting the low pressure boilers to two 32MW units. In many instances in the 1920s and 1930s new high pressure turbines were employed in connection with existing low pressure systems; a method known as superposition. (In a few cases the high pressure turbine was actually mounted on the low pressure turbine casing.) It allows existing, serviceable plant to be increased in efficiency and power output for minimum cost, and for this reason it was popular in the post-depression years of the 1930s. The rapid advances in steam turbine technology in the 1920s were not achieved without some cost. Early in this period the manufacturers of disc turbines experienced an exceptional number of failures. It was discovered that the discs were subject in service to vibrational oscillations leading to fatigue failure. This was overcome by employing better design theories and also by using heavier discs, including discs forged integral with the shaft. Important contributions to the solution of this problem were made both in the United STEAM AND INTERNAL COMBUSTION ENGINES 299 States and in Europe; particularly noteworthy was the work of Wilfred Campbell of the General Electric Company, after whom one of the important design tools developed at that time, the Campbell Diagram, is named. 1930–1940: refinement Between 1930 and 1940 no really striking advances of the type seen in the 1920s occurred. The depression following the stock market crash in 1929 did not provide a business climate in which large orders for steam turbines could be expected. Consequently, turbine manufacturers turned to improving the detailed design of steam turbines and a number of features that are now common practice in steam turbine construction were introduced. As a result of the desire to increase the inlet pressure the double shell construction shown in Figure 5.9 was introduced. This decreases the load on the turbine casing and fastenings by dividing the pressure difference from the turbine inlet pressure to the ambient pressure between two casings, one inside the other. In the 1930s creep, the slow ‘plastic movement’ of steel subjected to high temperatures and pressures, which was first identified in the 1920s, began to be considered in steam turbine design. This phenomenon can significantly alter stress distributions during the time of exposure of the turbine parts to operating conditions. Its effects can be minimized by adding suitable alloying materials to the steel that stabilize the material, and by developing extensive empirical data on the material’s properties for use in design. Warped, or twisted, low pressure blades, were introduced in the 1930s. These compensate for the effect of variations in the blade tangential velocity with radius so as to ensure that the steam impinging on the blade enters the blade smoothly and with the minimum flow disturbance. 1940–1950: increasing speed Steam turbine progress in the 1940s was constrained by the outbreak of the Second World War. After the war ended turbine design showed a definite trend away from the standard speed of 1500/1800rpm and the establishment of the high speed turbine operating at 3000/3600rpm. Higher speed turbines are smaller and lighter than comparable low speed machines. Smaller turbines expand less when heated, so distortion is decreased, which improves the turbine’s long-term reliability. Consequently, the high speed machine is better adapted to increasing inlet pressures and temperatures and to the application of reheat, which were features of steam turbine design from the late 1940s onwards. PART TWO: POWER AND ENGINEERING 300 In about 1948 there was a revival of interest in the reheat cycle (first introduced in Britain at North Tees in 1920, and in the United States at Philo, Ohio, in 1924) which was an added incentive for the introduction of the high speed turbine. In this cycle the temperature of the partially expanded steam is raised to about the original inlet temperature by passing it through a special heat exchanger, the reheater, which is heated either by the boiler combustion gases or by high temperature steam (the latter arrangement was only used in the 1920s), and then returned to the turbine for further expansion to the condenser pressure. The motivation for the earlier application in the 1920s had been to decrease the moisture, which causes blade erosion, in the low pressure sections of the turbine. Because the required additional valves and piping could not be economically justified at that time, the construction of new reheat turbines ceased in the early 1930s. The revival of interest in the late 1940s was stimulated by a need to increase plant efficiency in order to counteract rising fuel costs. Reheat produces a 4–5 per cent improvement in cycle efficiency, and has a number of other advantages compared to non-reheat operation. Thus, there is a reduction in the mass rate of flow of the steam, which, in turn, leads to a decrease in the size of the boiler feed pump, the boiler, the condenser, and of the feed water heating equipment. This, together with the ability to reduce the wetness of the steam at the exhaust makes reheat an attractive feature, and is widely used in modern steam cycles. 1950–1960: very high pressures and temperatures The first turbines handling steam at supercritical pressures (pressures in excess of 221bar/3200psig) were built in the 1950s with the aim of improving cycle efficiency, while avoiding the problems associated with increasing the turbine inlet temperature, which requires the development of new materials, an expensive and time-consuming process. The first supercritical turbine was installed in 1957 at the Philo station of the Ohio Power Company. It had an output of 125MW and used steam at 310bar (4500psig) and 621°C (1150°F). Double shell construction was used, and, because of the exceptionally high pressure, the outer casing is almost spherical in form. A significant feature of these very high pressure turbines is the small blade lengths in the high pressure stages (0.95cm in the Philo turbine), which results from the high density of the steam. The leakage space is, in consequence, a large fraction of the blade length, so turbines operating at very high pressures should, to offset the leakage loss, be designed for large outputs, e.g Philo (1957), 125MW; Philip Sporn (1959), 450MW; Bull Run (1965) 900MW. In about 1955 the ‘average’ steam temperature reached its present plateau of 566°C (1050°F). The attainment of this temperature was made possible by the STEAM AND INTERNAL COMBUSTION ENGINES 301 introduction of the ferritic stainless steels (11–13 Per cent chromium), and required the adoption of special design features. In the high pressure and intermediate pressure sections all parts were designed to allow free expansion and contraction. For example, the steam chests and valves were mounted separately from the turbine casing and connected to the inlet nozzles by flexible piping. To reduce temperature gradients in the turbine, the partially expanded steam was arranged to flow through the outer space of the double shell construction. 1960–1980: increasing size During the 1960s a number of extremely large output machines were placed in service. The mass rate of flow steam is so great (for a typical 660MW turbine about 2.1×10 6 kg/hr or 4.7×10 6 lb mass/hr) that multiple low pressure sections have to be provided. The 1960s was a period when many nuclear power stations commenced operation, and often these used either the boiling water or pressurized water cycles in which steam is supplied at pressures ranging from 31 bar (450psig) to 69bar (1000psig), with the steam dry and saturated at about 260°C (500°F). To compensate for the relatively low energy content of steam, the turbines have large outputs (1300MW), with correspondingly large steam mass flow rates (e.g., a 1300MW ‘nuclear turbine’ handles about 7.3×10 6 kg/hr (16.0×106lb mass/hr)). The resulting long (1.143m, 3.75ft) last row blades in the low pressure sections have forced American practice to adopt the 1800rpm tandem-compound design. Because of the lower speed, European designs of ‘nuclear turbines’ have sometimes been able to employ 3000rpm machines, but most examples of this type of turbine have been tandem-compounds operating at 1500rpm. The saturated inlet conditions result in a high moisture content in the turbine low pressure sections, leading to blade erosion unless reheating is used. Because the water cooled reactor provides only a low temperature heat source, reheating is not as effective as it is in fossil fuel-fired plants, so mechanical moisture separation must be used in addition to reheating. In the 1970s some even larger steam turbines, with power outputs in excess of 1000MW, came into service. To handle the large quantities of steam required by these machines, multiple low pressure stages were arranged in parallel. Figure 5.9 shows the section of one of the two 1300MW turbines completed by Brown Boveri in 1974 for the Gavin station of the American Electric Power Company. Machines of this type are the largest ever to be used on a fossil fuel-fired cycle. The historical development of the steam turbine can be summarized in a number of ways. Figure 5.10 shows the progress in power output, inlet temperature and inlet pressure from 1884 to 1984. . the mechanical design of the rotor and the last row blades, their manufacture and the materials used, and the means of attaching the blades to the rotor. Since sound forgings for the rotor can. of 9.12kg/kWhr (20.1lb/kWhr). These turbines had many of the features of the modern unit: they were two cylinder tandem compound (see below), and the condenser was placed below the level of the. in a substantial reduction in the overall length of the turbine. The first of these was the recovery of the velocity PART TWO: POWER AND ENGINEERING 294 Figure 5.7: Cross section of the first