Process Engineering Equipment Handbook Episode 3 Part 8 pptx

Shaft gland seal ring Spring-loaded labyrinth-type gland seals are used to minimize air leakage into and steam leakage from the steam turbine. An automatic gland steam control system and condenser can be supplied. Bearing housings The bearing housings are of horizontally split design and are arranged for pressure lubrication of the bearings. Bearings can be inspected and serviced without removing the coupling hub or breaking the pressure-containing casing joint. Journal and thrust bearings Tilting-pad radial bearings provide optimum rotor stability at all operating loads by forming an “oil wedge” at each shoe. This design is very effective in damping vibrations and is far superior to sleeve-type radial bearings. Double-acting, self- leveling thrust bearings are used, providing maximum protection against process upsets. Optional turning gear prior to startup and shutdown for large bearing span rotors can be supplied for automatic or manual operation. Auxiliary Systems* Microprocessor-based steam turbine governor control system The electrohydraulic control system automatically and continuously monitors and sets the steam turbine speed to satisfy customers’ specific requirements. Low maintenance operation typifies this type of system, which also features solid-state electronics, redundant speed pickups, and signal processing channels to reduce problems ordinarily associated with mechanical controls. The result is improved reliability, accuracy, and overall enhanced steam turbine performance. Valve gear assembly The valve gear assembly controls the steam flow to the turbine. It utilizes a simple but rugged bar lift design for increased reliability. The individual valve spindles are designed with spherically seated nuts to compensate for any minor misalignment while reducing the possibility of valve spin. The contact surfaces of both the control valves and seats are stellited for increased life. Another major feature is the large springs that are designed to enable fast closing times for the valves. Overspeed protection system Steam turbines are equipped with an emergency overspeed protection system that is separate from the main governing system. In the event of excessive rotor speed, this system shuts down the steam flow to the steam turbine by closing both the stop and control valves. This is an electronic failure detection system that is redundant and can be tested during operation. Electronic control system The steam turbine can be supplied with a programmable logic controller (PLC) system for digital control of steam turbine auxiliary systems and diagnostic monitoring of the steam turbine unit. This PLC system monitors operation and performance and safeguards against excess pressure and steam flow. T-78 Turbines, Steam * Source: Demag Delaval, USA. Oil supply systems The oil supply system is comprised of an integrated, freestanding console that provides oil for control valve and trip system hydraulics and bearings. These systems are furnished complete with coolers, filters, oil reservoirs, and integral piping. Either positive displacement or centrifugal oil pumps (sized to provide an adequate supply of oil to all bearings) are used. A separate hydraulic system can be supplied for larger capacity steam turbines and steam turbines with controlled extractions. See Fig. T-62. The turning gear system This system incorporates a modern “overrunning automatic clutch” that engages automatically when the rotor reaches turning gear speed, and it automatically disengages when that speed is exceeded. See Fig. T-63. The steam seal system This system features labyrinth-type packing to minimize steam leakage outward and air leakage inward. It vents high-pressure steam leakage to seal the low- pressure packing while the steam turbine is in operation. See Fig. T-64. Turbines, Steam T-79 FIG. T-62 Steam turbine lube oil and control oil consoles maximize the use of space while providing ample access for maintenance. (Source: Demag Delaval.) Vibration monitoring Shaft vibration detecting equipment continuously monitors the actual dynamic conditions of the machine. Temperature sensors Indicated thrust and journal bearing temperatures are systematically monitored. Steam Turbine Theory* General The steam turbine is the most widely used prime mover on the market. In large capacities, it rules without competition; for smaller sizes, the gas turbine and the internal combustion engine are its only competitors; but for the smallest sizes both T-80 Turbines, Steam FIG. T-63 Section through a steam turbine. (Source: Demag Delaval.) * Source: Demag Delaval, USA. the reciprocating steam engine and the internal combustion engine compete with the steam turbine for the market. Steam turbines have been designed and built for an output ranging from a few horsepower to more than 1,300,000 kW, with speeds ranging from less than 1000 rpm to more than 30,000 rpm, for inlet pressures from subatmospheric to above the critical pressure of steam, with inlet temperatures from those corresponding to saturated steam up to 1050°F, and for exhaust vacuums up to 29 1 / 2 inHg. The turbine requires much less space than an internal combustion engine or a reciprocating steam engine and much lighter foundations since reciprocating forces on the foundations are eliminated. Another major advantage is the turbine’s ability to extract power from the steam and then exhaust all the steam or part of it into a heating system or to a manufacturing process, entirely free from oil. Simplicity, reliability, low maintenance cost, and ability to supply both power and heat are the main justifications for the industrial turbine. A small factory or a building complex cannot produce electric power as cheaply as a large central-station power plant, but if steam is needed for industrial purposes or for heating, the production of power can be combined with the utilization of extracted or exhaust steam and the power becomes a cheap by-product. Turbines, Steam T-81 FIG. T-64 Interstage steam sealing. (Source: Demag Delaval.) The small noncondensing turbine also occupies a large and important field in power plants and marine installations because it is particularly well adapted to drive variable-speed auxiliaries and because its exhaust steam can be used to supply heat to the feedwater. A further advantage of the auxiliary turbine is its availability and convenience as a standby unit in case of interruptions to the power supply of motor-driven auxiliaries. Steam cycles The Rankine cycle. Potential energy of steam is transformed into mechanical energy in a turbine. The number of Btu required to perform work at the rate of one horsepower for one hour is 2544; for one kilowatt for one hour it is 3413 Btu. The enthalpy, or heat content, is expressed as Btu per pound of steam. This is, in effect, the potential energy contained in the steam measured above the conventionally accepted zero point (that of condensed steam at 32°F). Practically, it is not possible to release all the energy so that the end point of heat extraction in a condensing turbine is given by the temperature attainable in the condenser. The considerable amount of energy still contained in the steam at this point cannot be recovered and must be rejected to the cooling water. The portion of the potential energy that can be used to produce power is called the available energy and is represented by the isentropic enthalpy difference between the initial steam condition h 1 and the final condition corresponding to the exhaust pressure h 2 . If the condensate enthalpy is h w , the ideal Rankine cycle efficiency, or the thermal efficiency, is The available energy can be converted into mechanical (kinetic) energy only with certain losses because of steam friction and throttling, which increase the entropy of the steam. The end pressure is therefore attained at a higher steam enthalpy h¢ 2 than with isentropic expansion. The internal turbine efficiency then is This efficiency may be reduced to the external turbine efficiency by including mechanical and leakage losses not incident to the steam cycle. Since one horsepower-hour is equivalent to 2544 Btu per hour, the theoretical steam rate of the Rankine cycle in pounds per horsepower-hour is obtained by dividing 2544 by the available energy in Btu. The corresponding value on a pound- per-kilowatt hour basis may be found by dividing 3413 by the available energy. To obtain the actual steam rate at the coupling of the turbine, the theoretical steam rate is divided by the external turbine efficiency, which includes the mechanical losses. To facilitate steam-cycle calculations, standard tables of the thermodynamic properties of steam can be used. The data contained in these tables are plotted on a Mollier diagram (Fig. T-65), which is employed extensively to solve thermodynamic problems relating to steam turbines. Example. Determine the performance of a condensing turbine operating on a Rankine cycle based on the data in Tables T-4 and T-5. h i hh hh = -¢ - 12 12 h R w hh hh = - - 12 1 T-82 Turbines, Steam Turbines, Steam T-83 FIG. T-65 Mollier diagram (by permission from 1967 ASME Steam Tables). (Source: Demag Delaval.) Calculations: Improvements in the Rankine cycle may be obtained by raising the initial pressure and temperature. However, to avoid excessive moisture in the low- pressure stages, the increase in pressure must be accompanied by a corresponding increase in temperature. With present alloy steels the upper limit of the cycle is about 1050°F. The lower limit of the cycle depends on the maximum vacuum obtainable with the available cooling water and rarely exceeds 29 1 / 4 to 29 1 / 2 inHg. The economical limit of the cycle for a particular size of plant may be determined by a study of relative costs and savings. The reheat cycle. The reheat cycle, which is sometimes used for large units, is similar to the Rankine cycle with the exception that the steam is reheated in one or more steps during its expansion. The reheating may be accomplished by passing the partly expanded steam through a steam superheater, a special reheat boiler, or a heat exchanger using high-pressure live steam. The internal thermal efficiency of the cycle is calculated by totaling the available energy converted in each part of the expansion, as shown on the Mollier diagram, and dividing by the total heat supplied in the boiler, in the superheater, and in the reheat boiler or heat exchanger. External turbine efficiency Rankine-cycle steam rate measured steam rate 6.6 10.5 percent===63 Rankine-cycle steam rate lb hp h= - == ◊ 2544 2544 386 66 12 hh . Internal turbine efficiency percent== -¢ - = - - ==h i hh hh 12 12 1322 1054 1322 936 268 386 69 5. Ideal Rankine-cycle efficiency percent == - - = - - = = h R w hh hh 12 1 1322 936 1322 69 386 1253 30 8. Actual enthalpy drop Btu lb=-¢= - =hh 12 1322 1054 268 Isentropic enthalpy drop Btu lb=-= - =hh 12 1322 936 386 T-84 Turbines, Steam TABLE T-4 Initial steam pressure 200 psia Initial steam temperature 600°F Exhaust steam pressure 2 inHg Moisture in exhaust steam 5 percent Exhaust steam temperature 101°F Measured steam rate 10.5 lb/hp·h TABLE T-5 Enthalpy at inlet h 1 1322 Btu/lb Entropy at inlet 1.6767 Btu/°F Enthalpy at 2 inHg and entropy of 1.6767(h 2 ) 936 Btu/lb Enthalpy at 2 inHg and 5 percent moisture (h¢ 2 ) 1054 Btu/lb Enthalpy of saturated at 2 inHg (h w ) 69 Btu/lb In a plant operating with a steam pressure of 1000 lb/in 2 , a steam temperature of 750°F, and an exhaust pressure of 1 inHg absolute with one stage of reheating to 750°F at 175 lb/in 2 in a reheat boiler, the increase in thermal efficiency is about 7 1 / 2 percent. With two reheating stages the improvement over the straight Rankine cycle becomes approximately 10 1 / 2 percent. The main advantage of the reheat cycle is that excessive moisture in the low- pressure stages is avoided without employing a high initial steam temperature. The regenerative cycle. In the regenerative, or feed-heating, cycle, steam is withdrawn from the turbine at various points to supply heat to the feedwater. A considerable gain in economy may be obtained by using this cycle because the extracted steam has already given up part of its heat in doing work in the turbine and because the latent heat of the steam condensed in the feedwater heaters is conserved and returned to the boiler, thus reducing the heat loss to the condenser. The cycle efficiency may be calculated using a method similar to that already mentioned, but in connection with this cycle it is customary to design a flow diagram and to prepare a complete heat balance of the plant. In small and medium-sized plants, one or two extraction heaters may be used in addition to the exhaust heater that serves the steam-driven auxiliaries, and in large plants up to seven heaters may be employed. Additional plant economies result from reduced size of the condenser. From the viewpoint of steam generation, however, the load on the boiler is slightly increased to compensate for the steam extracted to the feed heaters. Furthermore, the higher temperature of the feedwater, while reducing the size of the economizer, also decreases the boiler efficiency by raising the lower level of the combustion gas cycle. This conflict between turbine and boiler cycle efficiencies may be removed by installing an air heater, which restores this lower level and permits the full benefit of the more economic method of regenerative feed heating. Regenerative feedwater heating. The basic principles of this cycle have been discussed previously. There is an optimum temperature to which the condensate can be heated. When this limit is exceeded, the amount of work delivered by the extracted steam is reduced and the benefit to the cycle gradually diminishes. If we assume, as an example, steam conditions of 400 lb/in 2 and 750°F at the throttle and a 29- inHg vacuum, the most favorable feedwater temperature is about 240°F for one stage of feedwater heating, 290°F for two stages, 320°F for three stages, and 330°F for four stages, as shown in Fig. T-66. As the number of heating stages is increased, the savings become proportionately less, as illustrated by the curves. For the steam conditions noted above, the cycle is improved by a maximum of 6 percent with one stage, 7 3 / 4 percent with two stages, 9 percent with three stages, and 9 3 / 4 percent with four stages. For this reason, it is not economically sound to install more than one or two heaters for a small- capacity turbine. Furthermore, the overall plant economy may limit the maximum feedwater temperature. With the condensate heated to a higher temperature because of the increased number of feed-heating stages, the temperature difference available to the economizer, usually provided in the boiler, becomes less; therefore, less heat will be extracted from the flue gases by the economizer. The resulting increase in stack loss and corresponding decrease in boiler efficiency may thus more than outweigh the improvement in the turbine cycle. The use of air preheaters instead of economizers to recover the stack loss makes it possible to obtain the full benefit from the regenerative feed-heating cycle. Regenerative feedwater heating affects the distribution of steam flow through the turbine. The steam required to heat the feedwater is extracted from the turbine at Turbines, Steam T-85 various points, determined by the temperature in the corresponding feed-heating stage. The extracted steam does not complete its expansion to the vacuum at the turbine exhaust; thus somewhat less power is delivered than with straight condensing operation. To obtain equal output, the steam flow to the turbine must therefore be slightly increased, as shown in Fig. T-67, which refers to the same steam conditions as in Fig. T-66. It may be noted from Fig. T-67 that, for instance, with one stage of feedwater heating to the optimum temperature of 240°F, it is necessary to add about 7 1 / 2 percent to the throttle flow and that with two stages the increase is about 10 1 / 2 percent, etc. On the other hand, a certain percentage of the total steam flow is extracted; thus the flow to the condenser is reduced as shown in Fig. T-68. For one and two feed- heating stages in the above example the decrease in steam flow to the condenser is about 8 and 10 1 / 2 percent, respectively, as compared with straight condensing operation. The tube surface and size of the condenser can therefore be reduced by similar amounts. Furthermore, the redistribution of the flow benefits the turbine; the first stages, which usually operate with partial admission, can easily handle more steam T-86 Turbines, Steam FIG. T-66 Reduction in enthalpy consumption due to regenerative feedwater heating (steam conditions: 400 lb/in 2 , 750°F, 29 inHg). (Source: Demag Delaval.) FIG. T-67 Increase in steam flow to turbine due to regenerative feedwater heating (steam conditions: 400 lb/in 2 , 750°F, 29 inHg). (Source: Demag Delaval.) [...]... 30 .0 lb hp ◊ h 192.5 ¥ 0.47 ¥ 0.96 ¥ 0. 98 The rotational loss of a 24-in-pitch-diameter wheel at 36 00 rpm, determined from Fig T-71, is about 6 .3 hp This diagram is based on atmospheric exhaust pressure; therefore, a correction factor must be applied as noted At 10-lb back pressure the specific volume of the steam is about 16 .3 ft3/lb Thus Loss hp = 6 .3 ¥ Steam rate of turbine = 30 .0 ¥ 22 = 8. 5 16 .3. .. 30 .0 ¥ 22 = 8. 5 16 .3 500 + 8. 5 = 30 .5 lb hp ◊ h 500 T-94 Turbines, Steam TABLE T-7 Small units Medium units Large units Large units 150 400 600 900 to to to to 400 lb/in2; 500 to 750°F 600 lb/in2; 750 to 82 5°F 900 lb/in2; 750 to 900°F 35 00 lb/in2; 82 5 to 1050°/F The use of the short method and Fig T-67 results in this case in a steam rate of 31 .4 lb/hp · h, which is about 3 percent higher than that... See Figs T -88 and T -89 Reference and Additional Reading 1 Bloch, H., and Soares, C M., Process Plant Machinery, 2d ed., Butterworth-Heinemann, 19 98 Steam Separators for Steam Drum Applications* Solids in boiler water This information source guarantees less than 1 ppm of total dissolved solids in the outlet steam with 2000 ppm boiler water concentration * Source: Peerless, USA T-109 FIG T -88 Lubrication... Turbines, Steam T- 93 FIG T-71 Rotational loss, average for single-stage turbines (two-row wheel; atmosphere exhaust) (Source: Demag Delaval.) The available energy is 205 Btu; subtracting a 12.5-Btu drop through the governor valve leaves 192.5 net Btu, which corresponds to a theoretical steam —— — velocity C = 2 23. 8 ¥ ÷192.5 = 31 04 ft/s The bucket speed u = 36 00 ¥ 24 ¥ p/60 ¥ 12 = 37 7 ft/s Thus the velocity... the bucket speed is rather low If we assume, for instance, 35 0 ft/s, corresponding to about T- 98 Turbines, Steam 221/2-in pitch diameter at 36 00 rpm, the number of stages required would be about 5; and at 30 0 ft/s with 19-in pitch diameter the number of stages would be 7, etc (Provisional inlet and outlet connections can be determined from Fig T -89 , thus indicating the general overall dimensions of the... other spring-loaded plunger Figure T -87 illustrates a hydraulic trip valve used with a mechanical shaftmounted trip When trip speed is reached, the plunger (or ring) mounted in the trip Turbines, Steam FIG T -84 Extraction governor with master regulator (Source: Demag Delaval.) FIG T -85 Overspeed trip plunger (Source: Demag Delaval.) T-107 T-1 08 Turbines, Steam FIG T -86 Two-ring (shockproof) overspeed... provisional dimensions of a 30 00-hp 36 00-rpm condensing turbine operating at 400 lb/in2, 750°F, and 28 inHg A turbine efficiency of 73 percent is desired; thus, for a size factor of, say, 95 percent, the required efficiency is 77 percent, corresponding to a quality factor of about 7500 The available enthalpy is 460 Btu; consequently the sum of velocity squares is 7500 ¥ 460 = 3, 450,000 Various combinations... inHg absolute Insert: 1450 35 00 psig; 900–1050°F; 1–5 inHg absolute Atmospheric, 100 psig; saturated, 750°F; 1–5 inHg absolute 1450 35 00 psig; 900– 1050°F; 1–5 inHg absolute 100–2400 psig; saturated, 1050°F; 1–5 inHg absolute 100–2400 psig; saturated, 1050°F; 1–5 inHg absolute 400–1450 psig; 750– 1050°F; 1–5 inHg absolute 600 35 00 psig; 600–1050°F; atmospheric, 1000 psig 600 35 00 psig; 600–1050°F; atmospheric,... topping unit matches approximately the 500°F assumed at the existing steam header Based on a total steam flow of 65,000 lb/h and a steam rate of 39 lb/kWh, the increase in power is about 1665 kW at the full-load condition Thus the increase in capacity is 33 .3 percent; likewise, the combined turbine steam rate is 9.75 lb/kWh, an improvement of 25 percent To calculate the corresponding fuel saving, additional... designed for partial admission with the nozzles covering only a part of the full circumference; therefore, the diameter of the wheel may be chosen independently of the bucket height Used as a first stage in a multistage turbine, the impulse stage with partial admission permits adjustment of the nozzle area by arranging the nozzles in separate groups under governor control, thus improving partial-load . efficiency percent == - - = - - = = h R w hh hh 12 1 132 2 936 132 2 69 38 6 12 53 30 8. Actual enthalpy drop Btu lb=-¢= - =hh 12 132 2 1054 2 68 Isentropic enthalpy drop Btu lb=-= - =hh 12 132 2 936 38 6 T -84 Turbines, Steam TABLE T-4 Initial. is about 16 .3 ft 3 /lb. Thus Steam rate of turbine 30 .0 30 .5 lb hp h=¥ + =◊ 500 8 5 500 . Loss hp 1 .3 =¥ = 63 22 6 85 Basic steam rate 192.5 30 .0 lb hp h= ¥¥¥ =◊ 2544 047 096 0 98 Turbines,. rate 6.6 10.5 percent=== 63 Rankine-cycle steam rate lb hp h= - == ◊ 2544 2544 38 6 66 12 hh . Internal turbine efficiency percent== -¢ - = - - ==h i hh hh 12 12 132 2 1054 132 2 936 2 68 38 6 69 5. Ideal Rankine-cycle