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3 Steam Generators INTRODUCTION Steam generators, or boilers as they are often called, form an essential part of any power plant or cogeneration system. The steam-based Rankine cycle has been synonymous with power generation for centuries. Though steam parameters such as pressure and temperature have been steadily increasing during the last several decades, the function of the boiler remains the same, namely, to generate steam at the desired conditions efficiently and with low operating costs. Low pressure steam is used in cogeneration plants for heating or process applications, and high pressure superheated steam is used for generating power via steam turbines. Steam is used in a variety of ways in process industries, so boilers form an important part of the plant utilities. In addition to efficiency and operating costs, another factor that has introduced several changes in the design of boilers and associated systems is the stringent emission regulations in various parts of the world. As discussed in Chapter 5, the limits on emissions of NOx;CO; SOx, and particulates have impacted the design and features of steam generators and steam plants, not to mention their costs. Today’s cogeneration systems and power plants resemble chemical plants with NOx; SOx, and particulate control systems forming a major portion of the plant equipment. Oil- and gas-fired packaged boilers used in cogeneration and combined cycle plants have also undergone significant changes during the last few decades. Selective catalytic reduction Copyright © 2003 Marcel Dekker, Inc. systems (SCRs) are used even in packaged boilers for NOx control, adding to their complexity and costs. Steam pressure and temperature ratings of large utility boilers have been increasing in order to improve overall plant efficiency. Several supercritical plants have been built during the last decade. There have been improvements in the design of packaged boilers too. Figure 3.1 shows the general arrangement of a packaged steam generator. The standard refractory-lined packaged boilers of the last century are being slowly replaced by custom-designed boilers with comple- tely water-cooled furnaces (Fig. 3.2). The air heater that was once an integral part of oil- and gas-fired boilers is now replaced by the economizer, which helps to FIGURE 3.1 Package water tube boiler. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. lower NOx levels. To improve efficiency, a few plants are even considering the use of condensing economizers. Though pulverized coal–fired boilers form the backbone of utility plants, fluidized bed boilers are finding increasing application when it comes to handling solid fuels with varying moisture, ash, and heating values; they also generate lower emissions of NOx and SOx. Oil- and gas-fired fire tube boilers (Fig. 3.3) are widely used in small process plants for generating low pressure saturated steam. Though different types of boilers are mentioned in this chapter, the emphasis is on the oil- and gas-fired packaged water tube steam generator, which is fast becoming a common sight in every cogeneration and combined cycle plant. BOILER CLASSIFICATION The terms boiler and steam generator are often used in the same context. Boilers may be classified into several categories as follows: By Application: Utility, marine, or industrial boiler. Utility boilers are the large steam generators used in power plants generating 500–1000MWof FIGURE 3.2 Completely water-cooled furnace design. (Courtesy of ABCO Indus- tries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.3a Fire tube boiler—wetback design. Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.3b Fire tube boiler—dryback design. Copyright © 2003 Marcel Dekker, Inc. electricity. They are generally fired with pulverized coal, though fluidized bed boilers are popping up in some plants. Utility boilers generate high pressure, high temperature superheated and reheat steam; typical para- meters are 2400 psig, 1000=1000F. A few utility boilers generate supercritical steam at pressures in excess of 3500 psig, 1100=1100= 1100F. Double reheat cycles are also in operation. Industrial boilers used in cogeneration plants generate low pressure steam at 150 psig to superheated steam at 1500 psig at temperatures ranging from 700 to 1000F. By Pressure: Low to medium pressure, high pressure, and supercritical pressure. Process plants need low to medium pressure steam in the range of 150–1500 psig, which is generated by field-erected or packaged boilers, whereas large utility boilers generate high pressure (above 2000 psig) and supercritical pressure steam. By Circulation Method: Natural, controlled, once-through, or combined circulation. Figure 3.4 illustrates these concepts. Natural circulation is widely used for up to 2400 psig steam pressure. There is no operating cost incurred for ensuring circulation through the furnace tubes, because gravity aids the circulation process. Controlled and combined circulation boilers use pumps to ensure circulation of a steam–water mixture through the evaporator tubes. Supercritical boilers are of the once- through type. It may be noted that once-through designs can be employed at any pressure, whereas supercritical pressure boilers must be of a once-through design. By Firing Method: Stoker, cyclone furnace, fluidized bed, register burner, fixed or moving grate. By Construction: Field-erected or shop-assembled. Large industrial and utility boilers are field-erected, whereas small packaged fire tube boilers up to 90,000 lb=h capacity and water tube boilers up to 250,000 lb=h are generally assembled in the shop. Depending on shipping dimensions, these capacities could vary slightly. By Slag Removal Method: Dry or wet bottom, applicable to solid-fuel-fired boilers. By Heat Source and Fuel: Solid, gaseous, or liquid fuels, waste fuel or waste heat. Waste heat boilers are discussed in Chapter 2. The type of fuel used has a significant impact on boiler size. For example, coal-fired boiler furnaces are large, because a long residence time is required for coal combustion, whereas oil- and gas-fired boilers can be smaller, as shown in Fig. 3.5. According to Whether Steam is Generated Inside or Outside the Boiler Tubes: Fire tube boilers (Fig. 3.3), in which steam is generated outside the tubes, are used in small plants up to a capacity of about 60,000 lb=h Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.4 Boiler circulation methods. (a) Natural; (b) forced circulation; (c) once-through; (d) once-through with superimposed circulation. 1, Economizer; 2, furnace; 3, superheater; 4, drum; 5, orifice; 6, circulating pumps; 7, separator. FIGURE 3.5 The impact of fuel on furnace size. Copyright © 2003 Marcel Dekker, Inc. of saturated steam at 300 psig or less; they typically fire oil or gaseous fuels. Water tube boilers, in which steam is generated inside the tubes, can burn any fuel, be of any size, and operate at any pressure but are generally economical above 50,000 lb=h capacity. See Chap. 2 for a comparison between fire tube and water tube waste heat boilers. STEAM PRESSURE AND BOILER DESIGN The energy absorbed by steam is distributed among feedwater heating (sensible heat), boiling (latent heat), superheating, and reheating functions. The distribution ratios are a function of steam pressure, as can be seen from steam tables or from Fig. 3.6. If the latent heat is large as in low pressure steam, a large furnace is required for the boiler; as the pressure of steam increases, the latent heat portion decreases and the superheat and reheat energy absorption increases. The boiler design accordingly varies with large surface areas required for the superheaters and reheaters and a small furnace with little or no convective evaporator surface in particular. The sensible heat, which is absorbed in the economizer, is also high at high pressure. The distribution of energy among the various surfaces—the furnace, evaporator, superheater, reheater, and economizer—is somewhat flexible, as will be shown later, but it must be emphasized that steam pressure plays a significant role in determining the sizes of these surfaces. FIGURE 3.6 Distribution of energy in boilers as a function of steam pressure. Copyright © 2003 Marcel Dekker, Inc. In natural circulation units the density differential between the cooler water in the downcomers and the less dense steam–water mixture in the riser tubes of the furnace provides the hydraulic head for circulation of the steam–water mixture through the evaporator tubes. The circulation ratio, CR, which is the ratio of the mixture flow to steam flow, could be in the range of 6–8 in high pressure boilers. In packaged boilers operating at low steam pressure, say 150– 1000 psig, the CR could be higher, ranging from 10 to 20. Note that we are referring to an average value. The circulation ratio will differ for each parallel circuit, depending on its length, tube size, heat flux, and static head available, as discussed in Q7.29. The controlled circulation boiler is operated at a slightly higher steam pressure, around 2500–2600 psig, and flow is ensured through the furnace tubes by a circulating pump; which forces the boiler water through each circuit. The circulation ratio is preselected in the range of about 2–4. This is done to reduce the operating cost associated with the circulating pumps; also, the use of carefully selected orifices ensures the flow of the steam–water mixture through each circuit. Hence a low CR is used in these systems. The once-through unit with superimposed circulation requires the circulating pump during start-up and at low loads when flow through the circuits is not high and later switches to the once-through mode at higher loads. PACKAGED STEAM GENERATORS Packaged boilers are widely used in cogeneration and even in combined cycle plants as auxiliary boilers providing steam for turbine sealing and steam for other uses when the gas turbine trips and the HRSG is not in operation. These boilers are generally shop-assembled and custom-designed. Typically, boilers of up to 250,000 lb=h capacity can be shop-assembled and larger units are field-erected. Steam parameters vary from 150 psig saturated to 1500 psig, 1000F. They typically burn natural gas, distillate fuel oils, and even heavy residual oils. Widely used methods for NOx control are low-NOx burners, flue gas recirculation, and selective catalytic reduction systems (SCRs). Carbon monoxide catalysts are also used if required. Emission control methods are discussed in Chapter 4. Packaged boilers could be further classified as D, A, or O-type depending on their construction, as shown in Fig. 3.7. In the A- and O-type boilers, the flue gases exit the furnace and then make a 180 turn, split up into two parallel paths, and flow through the convection section, then recombine to flow through the economizer. Using a convective superheater in this type of boiler is tricky, because it has to be split into two halves. A radiant design may be located at the furnace exit, but it operates in a harsh environment as discussed later. D-type boilers are widely used in industry. The flue gases generated in the furnace travel though the furnace, make a turn, and go through the convection Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.7 A-, D-, and O-type boiler configurations. 1, Burner; 2, steam drum; 3, mud drum. Copyright © 2003 Marcel Dekker, Inc. bank and then through the economizer to the stack. The gas flow is not split into two parallel paths as in the A- or O-type designs. If a superheater has to be located in the convection bank, the D-type design is the most convenient, because there is no concern with maldistribution in gas flow between parallel paths as with the O- and A-type boilers, which may lead to thermal performance issues. However, the O- and A-type boilers are more suitable as mobile units, because they have balanced weight distribution; rental boilers, which move from location to location, are generally of A- and O-type designs. The gas-fired O-type boiler shown in Fig. 3.8 is another variation of packaged boiler design. In this boiler the flue gases do not make a turn at the furnace end; the gases flow straight beyond the furnace to a convection section consisting of bare and finned tubes; the finned tubes make the convection section compact, thus reducing the overall length of the boiler. The advantage of this design is that the width required is not large, because the width of the furnace determines the width of the unit, whereas in a typical O- or A-type boiler the width of the furnace is added to that of the convection bank, making it difficult to ship the boiler to certain areas of the country or the world. Also, a convective type of superheater can be easily located behind a screen section. The advantages of the convective superheater over a radiant design are discussed later. A recent application for packaged boilers has been in combined cycle plants. These plants require steam for turbine sealing purposes when the HRSG trips, and they need it at short notice, say, within 5–15min. Packaged boilers with completely water-cooled furnace designs are well suited for fast start-ups, as discussed later. Very high steam purity as in utility plants can be obtained in packaged boilers through proper design of steam drum internals. Depending on the application, steam purity in the range of 30–100 parts per billion (ppb) can be FIGURE 3.8 A gas-fired O-type package boiler with extended surfaces. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. achieved. Packaged boiler designs have evolved over the years and have adapted well to the needs of the industry. Standard Boilers Standard boilers, which are pre-engineered packages, are inexpensive and are used in applications that are not very demanding in terms of process or emission limits. Decades ago, various manufacturers had developed so-called standard designs for boilers of 40,000–200,000 lb=h capacity with fixed dimensions of furnace, tubes, tube spacing, lengths, and surface areas. If someone wanted a boiler for a particular capacity that was not listed, the next or closest standard model would be offered. Standard models are less expensive than custom designs because no engineering is required to design and build them. It must be borne in mind that these designs were developed 30–50 years ago when the concept of flue gas recirculation and low-NOx burners were unheard of. They also had a lot of refractory in their design—on the floor, front walls, and rear walls—because completely water-cooled furnace designs had not yet been developed. The concerns with refractory-lined boilers are discussed later. However, emission regulations are forcing suppliers to custom design the boilers. As discussed in Chapter 4, the effect of flue gas recirculation and changes in excess air levels have to be reviewed on a case-to-case basis depending on the NOx and CO levels desired. Hence standard furnace dimensions may or may not be suitable for a given heat input, because the flame shape varies according to the NOx control method used. Flame lengths with low-NOx burners can be wider or even longer than with regular burners. Hence the use of low-NOx burners makes it difficult to select a standard boiler that meets the same need and is also an economical option. The furnace size could be compromised, which may result in flame impingement concerns with the burners used, or the gas pressure drop across the convection surfaces could be very large due to the flue gas recirculation rates used; the efficiency also could be lower due to the higher exit gas temperature associated with the larger flue gas flow. The operating cost due to a higher gas pressure drop is discussed below and in Chapter 4. Often gaseous and oil fuels are fired at excess air ranging from 10% to 20%; flue gas recirculation could be in the range of 10–35%, depending on the NOx level desired. In a few boilers, 9 ppmv NOx has been achieved with the burner operating at 15% excess air and 35% flue gas recirculation rate on natural gas firing. Thus it is possible to have a ‘‘standard’’ steam generator handling nearly 30–40% more flue gases than it was designed for in the good old days when 5–10% excess air was used without gas recirculation: A 100,000 lb=h standard boiler could be operating at gas flow conditions equivalent to those of a 140,000 lb=h boiler if it is not custom-designed. Of course, one could select a larger standard boiler, but it may or may not meet all the requirements of furnace Copyright © 2003 Marcel Dekker, Inc. dimensions, because developers of standard boilers generally increase furnace lengths for higher capacity but not the width or height, due to shipping constraints, particularly when the capacity is large. However, standard boilers are useful where one is not concerned about optimizing all the parameters such as efficiency, gas pressure drop, and emission levels and low initial cost is a primary objective. Packaged steam generators of today are custom-designed with an eye on operating costs and emissions. The furnace design also has undergone major design innovations, the completely water-cooled furnace (Fig. 3.2) being one of them. This design offers several advantages over the refractory-lined boilers designed decades ago. Advantages of Water-Cooled Furnaces Water-cooled furnaces have a number of advantages over other types: 1. The front, rear, and side walls are completely water-cooled and are of membrane construction, resulting in a leakproof enclosure for the flame, as shown in Fig. 3.2. The entire furnace expands and contracts uniformly, thus avoiding casing expansion problems. When refractory is used on the front, side, or rear walls, the sealing between the hotter membrane walls and the cooler outer casing is a concern and hot gases can sometimes leak from the furnace to the outside. This can cause corrosion of the casing, particularly if oil fuels are fired. 2. Problems associated with refractory maintenance are eliminated. Also, there is no need for annual shutdown of the boiler plant to inspect the refractory or repair it, thus lowering the cost of owning the boiler. 3. Fast boiler start-up rates are difficult with refractory-lined boilers because of the possibility of causing cracks in the refractory. However, with completely water-cooled furnaces, start-up rates are limited only by thermal stresses in the drums and are generally quicker. The tubes may be welded to the drums instead of being rolled if the start-ups are frequent. With boilers maintained in hot standby conditions using steam-heated coils located in the mud drum, even 10–15min start-ups are feasible. With a separate small burner whose capacity is 6–8% of the total heat input in operation during boiler standby conditions, the boiler can be maintained at pressure and can be ramped up to generate 100% steam within 3–5min. 4. Heat release rate on an area basis is lower for the water-cooled furnace by about 7–15% compared to the refractory-lined boiler. Some gas- fired boilers designed decades ago still use refractory on the floor; replacing this with a water-cooled floor will increase the effective heating surface of the furnace and lower the heat flux inside the tubes Copyright © 2003 Marcel Dekker, Inc. even further. The furnace exit gas temperature also decrease slightly due to the increased effective cooling surface of the furnace. A lower furnace exit gas temperature decreases the radiant energy transferred to a superheater located at the furnace exit and thus reduces the potential for superheater tube failures. A lower area heat release also helps reduce NOx, as can be seen from the correlations developed by a few burner suppliers. 5. Reradiation from the refractory on the front wall, side walls, and a floor increases the flame temperatures locally, which results in higher NOx formation. Of the total NOx generated by the burner, a significant amount of NOx is formed at the burner flame base, so providing a cooler environment for the flame near the burner helps minimize NOx to some extent. 6. Circulation was one of the concerns about the use of refractory on the floor of even gas-fired boilers because the D tubes are longer than partition tubes of the dividing wall. Heat fluxes in packaged boilers are generally low compared to those of utility boilers. To further protect the floor and roof tubes, a small inclination to the horizontal is used; also, considering the low steam pressure, tube-side velocities, heat flux, and steam quality, departure from nucleate boiling (DNB) has never been an issue, as evidenced by the operation of hundreds of boilers at pressures as high as 1000–1500 psig. The tube-side velocities are also adequate to ensure that steam bubbles do not separate from the water. Hence refractory is not required on the floor or front or rear walls for oil and gas firing. 7. Packaged boilers use economizers as the heat recovery equipment instead of air heaters, which only serve to increase the flame tempera- ture, thus increasing the NOx formation. The gas- and air-side pressure drops are also higher with air heaters, thus adding to the fan size and power consumption. The heat flux inside the furnace tubes is also reduced owing to the smaller furnace duty. Custom-Designed Boilers Custom-designed boilers, as the term implies, are designed from scratch. Based on discussions with the burner supplier and the level of NOx and CO desired, one first selects the type of burner to be used and the emission control strategy. A few options could be considered: Use a large amount of flue gas recirculation (FGR) and a low cost burner, which results in higher operating costs; one may use a large boiler with a wide convection bank to minimize gas pressure drop. Use an expensive burner, which uses fuel or air staging methods and requires little or no flue gas recirculation. A few burners can guarantee Copyright © 2003 Marcel Dekker, Inc. about 20–30 ppmv NOx (at 3% oxygen dry) on gas firing. Installation and operating costs associated with FGR are minimized. One can also consider the possibility of using a selective catalytic reduction (SCR) system along with a less expensive burner, which has a low to nil FGR rate. Steam injection may also be looked into, and the cost of steam versus FGR may be compared. Depending on the NOx and CO levels desired and the fuel analysis, the solution may vary from case to case, and no obvious solution exists for every situation. Thus one arrives at the best option from an emission control viewpoint and then starts developing the boiler design using the excess air and FGR rates for the fuels in consideration; the furnace dimensions to avoid flame impingement on the furnace walls are then arrived at. Assuming a specific exit gas temperature, the boiler efficiency calculations are done to arrive at the air and flue gas flow rates and the amount of flue gas recirculated. This is followed by an evaluation of furnace performance and design of the heating surfaces. The exit gas temperature from the economizer is arrived at and compared with the assumed value; efficiency is recalculated using the computed exit gas temperature, and revised air and flue gas flows are obtained. (Air and flue gas quantities depend on the amount of fuel fired, which in turn depends on efficiency.) Another iteration starting from the furnace is done to fine-tune the performance. The superheater performance is evaluated at various loads to determine whether the surface areas are adequate. If different fuels are fired, these calculations are carried out for all the fuels. Efforts are then made to reduce the fuel consumption and also lower the fan power consumption, which are recurring expenses, by fine-tuning the design of the evaporator and economizer. A large economizer may be used to improve the boiler efficiency if the duration of operation warrants it. The designer also has the ability to change the dimensions of the convection section—for example, the number of tubes wide, length, tube spacing, or even tube diameter—to come up with low gas pressure drop and hence low fan operating cost as shown below. Based on partial load performance and gas temperature profiles, bypass dampers may be required if an SCR system is used. Hence it is likely that the steam parameters of several boilers could be the same but the designs different due to the emission control strategy used and degree of custom designing. A computer program is used to perform these tedious calculations. Example 1 A 150,000 lb=h boiler firing standard natural gas and generating saturated steam at 285 psig with 230F feedwater uses 15% excess air and 15% flue gas recirculation. The exit gas temperature is 323F. Compare the performance of a standard boiler with that of a custom-designed unit. The flue gas flow through the Copyright © 2003 Marcel Dekker, Inc. boiler is 184,300 lb=h. With 80F ambient temperature, the efficiency is 83.38% HHV. The results of the calculations are shown in Table 3.1. The following points may be noted from this table: 1. The efficiency is the same in both designs because the exit gas temperature and excess air are the same. Also, the furnace dimensions are the same. Hence the furnace exit gas temperature is the same in both designs. 2. The convection sections are different. In the standard boiler, we used a standard tube spacing of 4 in. In the custom-designed unit, we reduced the surface area significantly by using fewer rows and also made the convection bank tube transverse spacing 5 in. This reduces the gas pressure drop in the convection bank by 4 in. WC. It also reduces the duty of the evaporator section, as can be seen by the higher exit gas temperature of 683F versus 550F. 3. We added a few more rows to the economizer in the custom-designed unit and made its tubes longer to obtain the same exit gas temperature and also to handle the additional duty. Economizer steaming is not a TABLE 3.1 Reducing Boiler Gas Pressure Drop Through Custom Designing Item Standard boiler Custom boiler Furnace lengthwidth height 32 ft 7ft 11 ft 32 ft 7ft 11 ft Furnace exit gas temp, F 2167 2167 Gas temp leaving evaporator, F 550 683 Exit gas temperature, F 323 323 Boiler surface area, ft2 8,920 6,710 Economizer area, ft2 10,076 14,107 Geometry Evaporator Economizer Evaporator Economizer Tubes=row 16 18 12 18 No. of rows deep 96 12 96 14 Effective length, ft 10 10 10 12 Gas pressure drop, in. WC 11.0 1.7 7.0 1.6 Transverse pitch, in. 4 4 5 4 Copyright © 2003 Marcel Dekker, Inc. concern in packaged boilers due to the small ratio of flue gas to steam flows (this aspect is discussed later). Hence we can absorb more energy in the economizer, which is a less expensive heating surface than the evaporator. The overall gas pressure drop saving of 4 in. WC results in a saving of 31 kW in fan power consumption (see Example 9.06b for fan power calculation). If energy costs 7 cents=kWh, for 8000 h of operation per year the annual saving is 31 0:07 8000 ¼ $17;360: This is not an insignificant amount. Simply by manipulating the tube spacing of the convection bank, we have dramatically reduced the fan power consumption and the size of the fan. Also the boiler cost for the two designs should be nearly the same because the increase in economizer cost is offset by the smaller number of evaporator tubes, which reduces the material costs as well as labor costs. To improve the energy transfer in evaporators one can also use finned tubes if the boiler is fired with natural gas or distillate fuels. For example, if we desire good efficiency but do not want an economizer because of, say, shorter duration of operation or corrosion concerns, we may consider using extended surfaces in the convection bank to lower the evaporator exit gas temperature by about 40–100F, which improves the efficiency by 1–2.5% compared to a standard boiler. 4. Another important point is that surface areas should be looked at with caution. One should not purchase boilers based on surface areas, which is still unfortunately being done. It is possible to distribute energy among the furnace, evaporator, and economizer in several ways and come up with the same overall efficiency and fan power consumption and yet have significantly different surface areas as shown in Tables 3.1 and 3.2. Comparing Surface Areas Example 2 This example illustrates the point that surface areas can be misleading. A boiler generates 100,000 lb=h of saturated steam at 300 psig. Feedwater is at 230F, and blowdown is 2%. Standard natural gas at 10% excess air is fired. Boiler duty¼ 100.8MM Btu=h, efficiency¼ 84.3% HHV, furnace backpressure ¼ 7in. WC It is seen from Table 3.2 that boiler 2 has about 10% more surface area than boiler 1 but the overall performance is the same for both boilers in terms of operating costs such as fuel consumption and fan power consumption. Also the Copyright © 2003 Marcel Dekker, Inc. energy absorbed in different sections is different, hence comparing surface areas is difficult unless one can do the heat transfer calculations for each surface. It has become a common practice (with the plethora of spreadsheet users) to compare surface areas of boilers and generally select the design that has the higher surface area. Surface areas should not be used for comparing two boiler designs for the following reasons: 1. Surface area is only a part of the simple equation Q ¼ UA DT , where U ¼ overall heat transfer coefficient, A¼ surface area, DT ¼ log-mean temperature difference, and Q¼ energy transferred. However, the Q and DT could be different for the two designs at different sections as shown in the above example. Hence unless one knows how to compute U ;A values should not be compared. 2. Even if DT remains the same for a surface, U is a function of several variables such as the tube size, spacing, and gas velocity. With finned tubes, the heat transfer coefficient decreases as fin surface area increases, as discussed in Q8.19. Hence unless one is familiar with TABLE 3.2 Comparison of Boilers with Same Efficiency and Backpressure Itema Boiler 1 Boiler 2 Heat release rate, Btu=ft3 h 90,500 68,700 Heat release rate, Btu=ft2 h 148,900 116,500 Furnace length, ft 22 29 Furnace width, ft 6 6 Furnace height, ft 10 10 Furnace exit gas temp, F 2364 2255 Evaporator exit gas temp, F 683 611 Economizer exit gas temp, F 315 315 Furnace proj area, ft2 (duty) 802 (36.6) 1026 (40.4) Evaporator surface, ft2 3972 (53.7) 4760 (52.1) Economizer surface, ft2 8384 (10.5) 8550 (8.3) Geometry Evaporator Economizer Evaporator Economizer Tubes=row 11 15 10 15 Number deep 66 14 87 10 Length, ft 9.5 11 9.5 10 Economizer, fins=in. ht thickness (serration) 3 0.75 0.05 0.157 50.75 0.05 0.157 Transverse pitch, in. 4 4 4.375 4 Overall heat transfer coeff 18 7.35 17.0 6.25 aDuty is in MM Btu=h, fin dimensions in inches, heat transfer coefficient in Btu=ft2 h F. Copyright © 2003 Marcel Dekker, Inc. all these issues, a simplistic tabulation of surface areas can be misleading. EFFECT OF STEAM PRESSURE ON BOILER DESIGN AND PERFORMANCE Another example of custom designing is shown in Example 3. In this example, we are asked to design a boiler for a lower pressure of operation for the first few years with the idea of operating at a higher steam pressure after that. Example 3 An interesting requirement was placed on the design of a boiler. The 175,000 lb=h boiler was to generate steam at 150 psig and 680F for the first few years and then operate at 650 psig and 760F. The piping and superheater changes had to be minimal when the time came for modifications. Operating a steam generator at two different pressures is a challenging task, particularly when a superheater is present. The reason is that the large difference in specific volume of steam affects the steam velocity inside the superheater tubes and the steam-side pressure drop, which in turn affect the flow distribution inside the tubes. The ratio of specific volume between the 150 and 650 psig steam is about 4. Hence for the same steam output, we could have a 4 times higher steam velocity at the lower pressure if the flow per tube were the same. Also, if the pressure drop at 650 psig were, say, 30 psi, it would be about 120 psi at the lower operating pressure if flow per tube were the same. Hence it was decided to manipulate the streams and steam flows as shown in Fig. 3.9. In the low pressure operation, there would be two inlets to the superheater from opposite ends of the headers as shown in Fig. 3.9a. This would make the velocity and pressure drop inside the tubes more reasonable. The total length of tubing traveled by steam in the low pressure option would be nearly half that of the high pressure case, which also reduces the pressure drop. Part of the steam is in parallel flow and part in counterflow. At high gas temperatures, as in this case, the difference in performance between parallel and counterflow superheaters is marginal. In the high pressure case, all the steam flows through the superheater tubes in counterflow. Because the specific volume is small, the steam can flow as shown with a reasonable steam velocity and without increasing the pressure drop. The performance in both, cases is shown in Table 3.3. Thus with a minimal amount of reworking, the piping could be changed when high pressure operation is begun. The superheater per se was untouched, and only the nozzle connections were redone. This boiler will be in operation for several years. If custom designing were not done, the capacity at low pressure mode would have to be limited to Copyright © 2003 Marcel Dekker, Inc. about 50–60% of the boiler capacity in order to avoid unreasonable steam velocity or pressure drop values. The main steam line has two parallel valves in the low pressure mode and will be converted to single-valve operation in the high pressure mode. BOILER FURNACE DESIGN The furnace is considered the heart of the boiler. Both combustion and heat transfer to the boiling water occur here, so it should be carefully designed. If not, several problems may result, such as lower or higher steam temperature if a FIGURE 3.9 Superheater piping arrangement for (a) low and (b) high pressure operation. TABLE 3.3 Boiler Performance at Low and High Steam Pressurea Low pressure High pressure Steam flow, lb=h 175,000 175,000 Steam temperature, F 680 760 Steam pressure, psig 150 650 Pressure drop, psi 23 46 aFeedwater¼ 230F; excess air¼ 15%; FGR¼17%; natural gas. Copyright © 2003 Marcel Dekker, Inc. superheater is used; the heat flux should be such as to avoid from DNB concerns. Circulation inside the tubes should be good. There could be incomplete combustion, which leads to lower efficiency and, coupled with a poor burner design, higher emissions of NOx and CO. Also, the flame should not impinge on the walls of the furnace enclosure. Hence it is always good practice to discuss emission control needs with potential burner suppliers who can model the flame shape and ensure that the furnace dimensions used can avoid flame impingement issues while ensuring the desired emission levels. In boilers fired with fuels that produce ash, the furnace is sized so that the furnace exit gas temperature is below the ash softening temperature. This is to avoid potential slagging problems at the turnaround section. Slag or molten deposits from various salts and compounds in the ash can cause corrosion damage and also affect heat transfer to the surfaces. The gas pressure drop across the convection section is also increased when the flow path is blocked by slag deposits. One of the parameters used in furnace sizing is the area heat release rate. This is the net heat input to the boiler divided by the effective projected area. This factor determines the furnace absorption and hence the duty and heat flux inside the tubes. Typically it varies from 100,000 to 200,000Btu=ft2 h for oil- and gas- fired boilers and from 70,000 to 120,000 Btu=ft2 h for coal-fired units. The volumetric heat release rate is another parameter, which is obtained by dividing the net heat input by the furnace volume. This is indicative of the residence time of the flue gases in the furnace and varies from 15,000 to 30,000Btu=ft3 h for coal-fired boilers. For oil and gaseous fuels it is not as significant a parameter as for fuels that are difficult to burn such as solid fuels. However, this parameter ranges from 60,000 to 130,000Btu=ft3 h for typical packaged oil- and gas-fired boilers. From the steam side, the circulation of the steam-water mixture in the tubes should be good. As discussed in Q7.30, several variables affect circulation, including static head available, steam pressure, tube size, and steam generation. The circulation is said to be adequate when the heat flux does not cause DNB conditions for the steam quality in consideration. Packaged boilers have a low static head, unlike field-erected industrial boilers, and also have longer furnace tubes. However, packaged boilers operate at low pressures, on the order of 200– 1200 psig, unlike large utility boilers, which operate at 2400–2600 psig, and circulation is better at lower pressures. Today’s boilers use completely welded membrane walls for the furnace enclosure (Fig. 3.2). Earlier designs were of tangent tube construction or had refractory behind the tubes (Fig. 3.10). With the refractory-lined casing, it is difficult to maintain a leakproof enclosure between the refractory walls and the water-cooled tubes, as a result flue gases can leak to the atmosphere, leading to corrosion, at the casing interfaces, particularly on oil firing. Balanced draft Copyright © 2003 Marcel Dekker, Inc. furnace design is used to minimize this concern, where the furnace pressure is maintained near zero by using a combination of forced draft and induced draft fans. The tangent tube design is an improvement over the refractory-lined casing. However, it has the potential for leakage across the partition wall. During operation the tubes in the partition wall are likely to flex or bend due to thermal expansion, paving the way for leakage of combustion gases from the furnace to the convection bank, resulting in higher CO emissions and also higher exit gas temperature from the evaporator and lower efficiency. Present-day boiler designs use forced draft fans, and the furnace is pressurized to 20–30 in. WC, depending on the backpressure. If SCR and CO catalysts are used, the back-pressure is likely to be even higher. With such a large differential pressure between the furnace and the convection bank, a leakproof combustion chamber is desired to ensure complete combustion. If gas bypassing occurs from the furnace to the convection side, the residence time of the flue gases in the furnace is reduced, thus increasing the formation of CO. Another concern with leakage of hot furnace gases from the furnace to the convection bank is the impact on superheater performance; the steam temperature is likely to be lower. The present practice is to use membrane walls. These consist of tubes welded to each other by fins as shown in Figs. 3.2 and 3.10. A gastight enclosure is thus formed for the combustion products. The partition wall is also leakproof, hence gas bypassing is avoided between the furnace and convection sections. This ensures complete combustion in the furnace enclosure. Typical designs at low pressures use 2 in. OD tubes at intervals of 3.5–4 in. depending on membrane tip temperature. Three-inch tubes have also been swaged to 2 in. and used at 4 in. FIGURE 3.10 Furnace construction—membrane wall, tangent tube, and refractory wall. Copyright © 2003 Marcel Dekker, Inc. pitch. This ensures a lower membrane temperature as well as reasonable ligament efficiency in the steam and mud drums. At pressures up to 700–750 psig, membranes using 2 in. tubes on 4 in. pitch have been found to be adequate due to the combination of low heat flux in the furnace and low saturation temperature, as evidenced by the operation of several hundred boilers. The 1 in. long membrane with appropriate thickness does not result in excessive fin tip temperatures or thermal stress concerns. At higher pressures, one may use 0.5 in. 0.75 in. long membranes. Figure 3.11 shows how fin tip temperatures vary with heat flux and membrane length. The furnace process is extremely complicated, because today’s burners have to deal with various aspects of burner designs such as staged fuel or staged air combustion, flue gas recirculation, and other NOx control methods; hence furnace performance should be arrived at on the basis of experience, field data, and calculations. The furnace exit gas temperature is the most important variable in this evaluation and is a function of heat input, flue gas recirculation rate, type of fuel used, effective cooling surface available, and excess air used. A gas-fired flame has less luminosity than an oil flame, so the furnace exit temperature is higher, as shown in Fig. 3.12. A coal-fired flame has an even higher furnace exit gas temperature. An oil flame is more luminous and the furnace absorbs more energy, resulting in higher heat flux in the furnace tubes. Energy Absorbed by the Furnace The energy transferred to the furnace is obtained from the equation Q ¼ Ape1e2sðT 4g T4wÞ ¼ Wf LHV Wghe FIGURE 3.11 Relating fin tip temperature to heat flux in membrane wall furnace. Copyright © 2003 Marcel Dekker, Inc. where Q¼ energy transferred to the furnace, Btu=h Tg ¼ average gas temperature in the furnace, R he ¼ enthalpy of flue gases corresponding to the furnace exit gas temperature Te, Btu=lb Tw ¼ average furnace wall temperature, R Ap ¼ effective projected area of the furnace, ft2 s¼ radiation constant e1; e2 ¼ emissivity of flame and wall, respectively LHV¼ lower heating value of the fuel, Btu=lb Wf ;Wg ¼ fuel and flue gas quantity, lb=h The emissivity of the flame may be determined by using methods discussed in Q8.08. The effective projected area includes the water-cooled surfaces and the opening to the furnace exit plane. If refractory is used on part of the surfaces, a correction factor of 0.3–0.5 has to be used for its effectiveness. Once the furnace duty is arrived at, the heat flux inside the tube may be estimated. Heat flux inside the tubes is a very important parameter because it affects the boiling process. Example 4 Determine the energy absorbed by the packaged boiler furnace firing natural gas for which data are given in Table 3.4. At 100% load, boiler duty or energy absorbed by steam¼ 118.71MM Btu=h. Flue gas flow¼ 125,246 lb=h at 100% load. FIGURE 3.12 Furnace outlet temperature for gas and oil firing. Copyright © 2003 Marcel Dekker, Inc. The net heat input to the furnace is 118:71 0:992 0:9273 ¼ 127 MM Btu=h where 0.992¼ 17 heat losses, and 0.9273 is the boiler efficiency on LHV basis. Net heat input Effective furnace area ¼ 127 10 6 1169 ¼ 108;900 Btu=ft2 h TABLE 3.4 Boiler Performance—Gas Firinga Load (%) 25 50 75 100 Boiler duty, MM Btu=h 29.14 50.09 89.03 118.71 Excess air, % 30 15 15 15 Fuel input, MM Btu=h 34.68 69.79 105.69 141.89 Heat rel rate, Btu=ft3 h 16,055 32,310 48,931 65,691 Heat rel rate, Btu=ft2 h 29,646 59,660 90,349 121,297 Steam flow, lb=h 25,000 50,000 75,000 100,000 Steam temperature, F 711 740 750 750 Economizer exit water temp, F 328 334 356 374 Boiler exit gas temp, F 525 587 666 739 Economizer exit gas temp, F 254 271 298 327 Air flow, lb=h 32,954 58,665 88,843 119,275 Flue gas, lb=h 34,413 61,602 93,290 125,246 Dry gas loss, % 3.71 3.58 4.08 4.62 Air moisture loss, % 0.1 0.1 0.1 0.12 Fuel moisture loss, % 10.48 10.55 10.67 10.79 Casing loss, % 1.2 0.6 0.4 0.3 Margin, % 0.5 0.5 0.5 0.5 Efficiency, % HHV 84.01 84.67 84.24 83.66 Efficiency, % LHV 93.12 93.85 93.37 92.73 Furnace back pressure, in. WC 0.8 2.61 6.21 11.49 aSteam pressure 500 psig; feedwater 230F, blowdown 1%, amb temp 80F; RH 60%, fuel- standard natural gas. Flue gas analysis (vol%): CO2 ¼ 8:29, H2O ¼ 18:17, N2 ¼ 71, 0:07;O2 ¼ 2:46. Boiler furnace projected area¼1169 ft2, furnace width¼7.5 ft, length¼ 32 ft, height¼9 ft. Copyright © 2003 Marcel Dekker, Inc. The furnace exit gas temperature from Fig. 3.12 is 2235F. It may be shown that the enthalpy of the flue gases at 2235F is 661.4 Btu=lb based on the flue gas analysis. (See Appendix, Table A8.) The furnace duty from (5)¼ 127 106 125; 246 661:4 ¼ 44:2MM Btu=h. The average heat flux based on projected area is 44:2 106 1169 ¼ 37;810 Btu=ft2 h However, what is of significance is the heat flux inside the boiler tubes, not the heat flux on a projected area basis. We can relate these two parameters as follows: qpSt ¼ qcðpd=2 þ 2hÞ where qp ¼ heat flux on projected area basis St ¼ transverse pitch of membrane walls, in. qc ¼ heat flux on circumferential area basis, Btu=ft2 h d ¼OD of furnace tubes h¼membrane height, in. Once qc is obtained, we can relate it to qi, the heat flux inside the tubes, as follows: qcd ¼ qidi where di ¼ tube inner diameter, in. Simplifying qi ¼ qpStðd=diÞ pd=2 þ 2h In our example, qp ¼ 37;810 Btu=ft2 h, St ¼ 4 in:; h ¼ 1 in:; d ¼ 2, di ¼ 1:706 in. Then qi ¼ 37;810 ð2=1:706Þ 4 3:14 2=2 þ 2 1 ¼ 34;500 Btu=ft 2 h Note that if we did the same calculation for oil firing, the heat flux would be higher, because the furnace exit gas temperature is lower. Heat flux inside tubes is an important parameter, because allowable heat fluxes are limited by circulation rates. Large heat flux inside tubes can lead to departure from nucleate boiling conditions. Estimating Fin Tip Temperatures Fin tip temperatures in boilers of membrane wall design depend on several factors such as cleanliness of the water or tube-side fouling, fin geometry, and heat flux, Copyright © 2003 Marcel Dekker, Inc. which is a function of the load and gas temperature. Assuming that membranes are longitudinal fins heated from one side, the following equation may be used to determine the fin tip temperature: tg tb coshðmhÞ ¼ tg tt where tg ¼ gas temperature, F tb ¼ fin base temperature, F Due to the high boiling heat transfer coefficients, on the order of 3000– 10,000 Btu=ft2 hF, fin base temperatures will be a few degrees higher than saturation temperature, assuming that tube-side fouling is mini- mal. tt ¼ fin tip temperature, F h¼membrane height, in. (see Fig. 3.11) m¼ðhgC=KAÞ0:5 where hg ¼ gas-side heat transfer coefficient, Btu=ft2 hF C ¼ perimeter of fin cross section¼ 2b þ L in. (for heating from one side) where b¼ fin thickness and L¼ fin length or furnace length K ¼ fin thermal conductivity, Btu=ft hF A¼ cross-section of fin¼ bL C=A for long fins ¼ ð2b þ LÞ=bL ¼ L=bL ¼ 1=b Example 5 In a boiler furnace, gas temperature at one location is 2200F. The gas-side heat transfer coefficient is estimated to be 30Btu=ft2 hF. Fin height¼ 0.5 in. fin thickness¼ 0.375 in. Fin base temperature is 600F. Thermal conductivity of fin is 20Btu=ft hF. Determine the fin tip temperature. Solution: Using the above equation, we have Tg ¼ 2200F; tb ¼ 600F; hg ¼ 30; h ¼ 0:5 in:; b ¼ 0:375 in:; K ¼ 20 mh ¼ 0:5 12 30 12 20 0:25 0:5 ¼ 0:3536 or coshð0:3536Þ ¼ 1:063 Tt ¼ 2200 2200 600 1:063 ¼ 695F Copyright © 2003 Marcel Dekker, Inc. THE BOILING PROCESS When thermal energy is applied to furnace tubes, the process of boiling is initiated. However, the fluid leaving the furnace tubes and going back to the steam drum is not 100% steam but is a mixture of water and steam. The ratio of the mixture flow to steam generated is known as the circulation ratio, CR. Typically the steam quality in the furnace tubes is 5–8%, which means that it is mostly water, which translates into a CR in the range from about 20 to 12. CR is the inverse of steam quality. Circulation calculations and the importance of heat fluxes are discussed in Q7.29. Nucleate boiling is the process generally preferred in boilers. In this process, the steam bubbles generated by the thermal energy are removed by the flow of the mixture inside the tubes at the same rate, so the tubes are kept cool. Boiling heat transfer coefficients are very high, on the order of 5000– 8000Btu=ft2 h F as discussed in Q8.46. When the intensity of thermal energy or heat flux exceeds a value known as the critical heat flux, then the process of nucleate boiling is disrupted. The bubbles formed inside the tubes are not removed adequately by the cooler water; the bubbles interfere with the flow of water and form a film of superheated steam inside the tubes, which has a lower heat transfer coefficient and can therefore increase the tube wall temperatures significantly as illustrated in Fig. 3.13. It is the designer’s job to ensure that we are FIGURE 3.13 Boiling process and DNB in boiler tubes. Copyright © 2003 Marcel Dekker, Inc. never close to critical heat flux conditions. Generally, packaged boilers operate at low pressures compared to utility boilers and therefore DNB is generally not a concern. The actual heat fluxes range from 40,000 to 70,000Btu=ft2 h, while critical heat flux could be in excess of 250,000Btu=ft2 h. However, one has to perform circulation calculations on all the parallel circuits in the boiler, particu- larly the front wall, which is exposed to the flame, to ensure that there is adequate flow in each tube. In the ABCO D-type boiler, carefully sized orifices are used to limit the flow of mixture through the D headers while ensuring flow through all the tubes in the front wall. Ribbed or rifled tubes are sometimes used as evaporator tubes. These tubes ensure that the wetting of the tube periphery is better than in plain tubes. They have spiral grooves cut into their inner wall surface. The swirl flow induced by the ribbed tubes not only forces more water outward onto the tube walls but also promotes general mixing between the phases to counteract the gravitational stratification effects in a nonvertical tube. Ribbed or twisted tubes can handle a much higher heat flux, often 50% higher than plain tubes. They are expensive to use but offer a safety net in regions of high heat flux, particularly in very high pressure boilers. In fire tube boilers, the critical heat flux may be estimated as shown in Q8.47. Again owing to the low pressure of steam, the allowable heat flux to avoid DNB is much higher than the actual values; hence tube failures are rare unless tube deposits or scale formation is severe. As discussed later in this chapter, maintaining good boiler water chemistry, ensuring proper blowdown, and adding chemicals to maintain proper alkalinity and pH in the boiler should minimize scale formation and thus prevent tube failures. BOILER EFFICIENCY CALCULATIONS The boiler efficiency is an important variable that is impacted by the type of fuel, its analysis, the exit gas temperature, excess air used, and ambient reference conditions. The major losses due to flue gases and the method of computing efficiency are discussed in Q6.19. With rising fuel costs, plant engineers should try to aim for higher efficiency if the plant is base-loaded and operates continuously. Often less efficient and less expensive units are purchased owing to lack of funds, and this practice should be reviewed. One should look at the long-term benefits to the end user. Similarly, the fan operating costs should also be evaluated. A design with high gas pressure drop in the boiler may be less expensive, but if one considers the long-term operating costs, it may not be the better choice. Table 3.5 shows the effect of excess air and exit gas temperatures on boiler efficiency and cost of operation. It is important to operate at as low an excess of air as possible; however, as discussed in Chapter 4, limits on NOx and CO may force the burners to use higher values of excess air. Copyright © 2003 Marcel Dekker, Inc. As shown in Tables 3.4 and 3.7, the efficiency of packaged boilers varies with load. This information may be used as a planning tool as discussed, particularly when the plant has HRSGs in addition to steam generators. Combination Firing Boiler efficiency calculations are done using ASME PTC 4.1 methods, as shown in Q6.19. When a combination of fuels is fired, the calculations can be involved. The results from a program developed are shown in Fig. 3.14. They show the performance of a boiler firing two different fuels at the same time. Based on the exit gas temperature and measured or predicted oxygen for the flue gas mixture, one can simulate the excess air and obtain the performance with individual fuels first and then obtain the combined effect on air and gas flows, flue gas analysis, combustion temperatures, heat losses, and efficiency. BURNERS The fuel burner is an important component of any boiler. Burner designs have undergone several iterations during the last decade. Burner suppliers such as Coen and Todd are offering burners that result in single-digit NOx emissions and very low CO levels, competing with the SCR system presently used in the industry for single-digit NOx emissions. However, these burners use a large amount of flue gas recirculation, and flame stability at low loads is a concern. Development work is going on to improve on these results. Fuel or air staging and TABLE 3.5 Effect of Excess Air and Exit Gas Temperature on Efficiencya Excess air (%) 5 20 5 20 Exit gas temp, F 300 300 400 400 Vol% CO2 9 7.97 9 7.97 H2O 19.57 17.56 19.57 17.56 N2 70.53 71.31 70.53 71.31 O2 0.89 3.16 0.89 3.16 Efficiency, % HHV 84.81 84.22 82.64 81.79 % LHV 94.11 93.46 91.71 90.70 Flue gas, lb=h 96,160 110,000 98,680 113,210 Annual fuel cost, MM$=yr 2.854 2.873 2.928 2.959 a Steam flow¼100,000 lb=h, 300psig sat, feedwater temp¼ 230F, 2% blowdown, ambient temp¼ 80F, relative humidity¼60%, boiler duty¼100.8MM Btu=h, fuel cost¼$3=MM Btu. Copyright © 2003 Marcel Dekker, Inc. steam injection are the other methods used by burner suppliers to control NOx. Today single burners are used for capacities up to 300–350MM Btu=h on gas or oil firing. Often more than one fuel is fired in the burner. When different gaseous fuels are fired in a burner, the fuel gas pressure has to be adjusted at the burner inlet to ensure proper fuel flow. Example 6 Let us say that a burner is firing 5MM Btu=h on LHV basis using a fuel of lower heating value, 1400Btu=ft3, and molecular weight 25.8 at a pressure of 30 psig. Assuming the nozzles remain the same, what should be done when a fuel of FIGURE 3.14 Efficiency calculations for simultaneous firing of fuels. Copyright © 2003 Marcel Dekker, Inc. heating value 700Btu=ft3 whose molecular weight is 11.6 is fired, the duty being the same? Solution: The gas pressure should be adjusted; otherwise it would be difficult to control the heat input. The pressure drop across the nozzles is related to the flow of fuel as follows (Subscripts 1 and 2 refer to fuels 1 and 2): DP1 ¼ KW 2=MW ¼ KQ2MW where Q¼volumetric flow W ¼mass flow MW¼molecular weight K is a constant¼ 30=Q21 MW Basically we are converting the pressure drop equation from mass to volumetric flow. Because the heat input by both fuels is the same, Q1 LHV1 ¼ Q2 LHV2 where LHV is the lower heating value of the fuel, Btu=ft3. DP2 ¼ 30 Q21 MW1 Q22 MW2 Rewriting Q2 in terms of Q1 and simplifying, we have DP2 ¼ 30 700 1400 2 25:8 11:6 ¼ 17 psi Thus we should have a lower fuel gas pressure to ensure the same heat input. COMBUSTION CONTROLS The function of a combustion control system is to ensure that the steam generation matches the steam demand. When the demand exceeds the supply, the steam pressure will decrease and vice versa. Although a few utility boilers generate steam at sliding pressures, packaged boilers typically generate steam at fixed pressure. The control system immediately adjusts the fuel input to maintain the steam pressure. The following methods are typically used for combustion control. Single-Point Positioning: This is a simple and safe system for combustion control. A common jackshaft is modulated by a power unit based on variations in drum pressure and is mechanically linked to both the fuel control value and the air control damper. This system is limited to small boilers, typically below 100,000 lb=h, that have an integral fan mounted Copyright © 2003 Marcel Dekker, Inc. on top of the wind-box and are fired by a single fuel of nearly constant heating value. Fuel heating values should not vary, and only one fuel can be fired at a time. When low CO values are desired such as less than 70 ppmv, an oxygen trim is added. Parallel Positioning System: This system is used on large boilers where a remote fan supplies air to the wind-box. It has separate pneumatic power units for controlling air and fuel. Full Metering with Cross Limiting: This system is expensive but is recommended for accurate air=fuel ratios, for keeping oxygen levels optimized, and for its firing precision. Fuel and air flows are measured continuously and are adjusted as required to maintain the desired air=fuel ratio. Air leads on load increases, and fuel leads on load decreases. This system allows simultaneous firing of two or more fuels. When emission levels are stringent and a large flue gas recirculation rate is used, this method is used and offers better control over the combustion process. As far as the boiler is concerned, a three-element-level control system is generally used to control the drum water level. Other controls would include steam temperature and master pressure control. Figure 3.15a and 3.15b show typical schemes of gas-side and steam-side instrumentation and controls, respec- tively, used in packaged boilers. FAN SELECTION Packaged steam generators of today use a single fan for up to 250,000 lb=h of steam. The furnaces of oil- and gas-fired boilers are pressurized, hence the fan parameters should be selected with care. Estimating the flow or head inaccurately can force the fan to operate in an unstable region or result in the horsepower being too high and the operation inefficient. The density of air should be accurately estimated, so elevation and ambient temperature conditions should be considered. In some cold locations, a steam–air preheat coil is used to preheat the air before it enters the fan, and this adds to the pressure drop. When flue gas recirculation is required, usually the flue gases from the boiler exit are sucked in by the fan, which handles the resistance of the entire system. The density of the mixed air is lower, owing to the higher temperature of the air mixed with the flue gases. The fan should be selected for the lowest density case, as explained in Q9.06, because the mass flow of air is important for combustion and not the volumetric flow. The effect of gas density on fan performance is shown in Fig. 3.16a. Large margins on flow and head should not be specified, because this leads to oversizing of the fan and can force the fan operating point to the extreme right of the curve in Fig. 3.16b, where the horsepower can be extremely high; a lot of energy is also wasted. Inlet vane control is typically used for controlling the flow Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.15a Scheme of boiler controls—gas side. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.15b Scheme of boiler controls—steam side. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. of air; this system typically operates stably between 20% and 100% vane opening, which does not translate into a large flow difference, as can be seen from Fig. 3.16c. Hence a small margin on flow and head is preferred—about 15% margin on flow and 20–25% on head is adequate; otherwise one may have to use a variable-speed drive or frequency modulation for control, which is expensive. Underestimating the fan head can also cause the fan to operate in the unstable FIGURE 3.16 (a) Fan performance and range of operation. (b) Effect of system resistance on fan horsepower. (c) Effect of vane position on flow reduction in fans. Copyright © 2003 Marcel Dekker, Inc. region as shown in Fig. 3.16a. Curve 2 in Fig. 3.16b is the estimated curve, and the actual curve 1 is to the left, close to the unstable region with positive slope. It also delivers less flow than required. The fan operating point must preferably be in the negatively sloping portion of the head versus flow curve; otherwise the fan could operate in the unstable region, causing surges and vibration. The flue gas recirculation lines must be properly sized; typical air and flue gas velocity in ducts is about 40 ft=s. The flue gas recirculation line is usually connected to the fan inlet in gas and distillate oil–fired boilers. This increases the size of the forced draft fan. The higher gas pressure drop in the boiler due to the increased mass flow should also be considered when selecting the fan. A separate recirculation fan is used occasionally when heavy fuel oils containing sulfur are fired and the flue gases are admitted into the burner wind-box. If the flue gases were allowed to mix with the cold air at the fan inlet, the mixture temperature could fall below the acid dew point, possibly leading to corrosion. The fan inlet duct and downstream ductwork must have proper flow distribution. Pulsations and duct vibrations are likely if the inlet airflow to the fan blades is not smooth and the maldistribution in velocity is large. Similarly, the ductwork between the fan and wind-box should be designed to minimize flow maldistribution to ensure proper airflow to the burner. SUPERHEATERS The superheater is an important component of a packaged boiler. The degree of superheat could be very high, with steam temperatures up to 1000F, or as low as FIGURE 3.16 Continued. Copyright © 2003 Marcel Dekker, Inc. 50F. With a very low degree of superheat, one can locate the superheater behind the evaporator and ahead of the economizer. In this case, the superheater may require a large surface area due to the low log-mean temperature difference, but extended surfaces may be used (if distillate oils and gaseous fuels are fired) to make it compact. Radiant superheaters, which are typically located in the furnace exit region, are widely used by several boiler manufacturers. Radiant superheaters have to be designed very carefully because they operate in a much harsher environment than convective superheaters, which are located in the convective zone behind screen tubes as shown in Fig. 3.17a. Radiant superheaters are located at the furnace exit or in the turning section (Fig. 3.17b). The furnace exit gas temperature is a difficult parameter to estimate. Variations in excess air, flue gas recirculation rates, and burner flame patterns can affect this value and the temperature distribution across the furnace exit plane. The gas temperature in operation could be off by 100–150F from the predicted value. The turning section is also subject to nonuniformity in gas flow and turbulence, which can affect the superheater performance. Thus its duty can be either underestimated or over- estimated by a large margin. The convective superheater is shielded behind screen tubes as shown in Fig. 3.17a and often operates at 1800–1900F in comparison with the 2200–2300F for radiant designs. Because it operates at lower tube wall temperatures, its life can be longer, but it requires a greater surface area because of the lower log-mean temperature difference. However, owing to the lower operating temperatures, a convective superheater can use a lower grade material than the radiant design, and this helps balance the cost to some extent. Also, its location behind screen tubes helps reduce the gas flow nonuniformity to a great extent; hence predicting its performance is easier and more reliable than predicting the performance of the single-stage radiant superheater. FIGURE 3.17 Location of convective and radiant superheater. 1, Superheater; 2, burner; 3, screen evaporator. Copyright © 2003 Marcel Dekker, Inc. Several boilers operate at partial loads of less than 60% for large periods. The radiant superheater, by its nature, absorbs more enthalpy at lower loads, hence the steam temperature increases at lower loads. Convective heat transfer depends on mass flow of flue gases, so as the load decreases, the gas flow and temperature decrease at the superheater region, and therefore the steam tempera- ture and the tube wall temperatures drop with load. Also if at 100% load the steam-side pressure drop in a radiant superheater is 50 psi, then at 30%, it will be about 5 psi, which can lead to concerns about steam flow distribution through the tubes when it is receiving more radiant energy per unit mass of steam. Coupled with nonuniform gas flow distribution at low loads and low gas velocities, the radiant superheater poses several concerns about its tube wall temperatures and hence its life. The convective superheater is located behind several rows of screen tubes that shield it from furnace radiation. Gas flow entering the superheater is well mixed; hence it is easier to predict its performance and tube wall temperatures. As mentioned earlier, its surface area requirement may be more, but one is assured of low tube wall temperatures and hence longer life. The steam temperature in a convective superheater generally decreases as the load falls off, whereas in a radiant design it remains within a small range over a larger load range. Hence the convective design has to be sized to ensure that the required steam temperature is achieved at the lowest load, which can increase its size and cost. The choice of whether to use a radiant or a convective superheater is based on the experience of the supplier. Because the surface area requirements are significantly different due to the different log-mean temperature differences, this is yet another reason that a comparison of surface areas can be misleading. If heavy oil is fired in the boiler, the problems associated with slagging and high temperature corrosion pose concerns for the longevity and operability of radiant superheaters as discussed below, so convective superheater designs are preferred in such cases. Packaged boilers use limited space compared to utility or field-erected boilers; with high gas velocities and slagging potential in the furnace exit region, the radiant design is vulnerable. Even with a convective superheater design, care should be taken to use retractable soot blowers, and there should be adequate space provided for cleaning and maintenance. Steam Temperature Control The steam temperature in packaged boilers is often controlled from 60% to 100% load by using a two-stage superheater design with interstage attemperation as shown in Fig. 3.18. Steam temperature can also be maintained from 10% to 100%; however, this calls for a much larger superheater surface area. Deminer- alized water should be used for attemperation, because it does not add solids to Copyright © 2003 Marcel Dekker, Inc. the steam. The solids in the feedwater used for attemperation should be in the same range as the final steam purity desired, which could be as low as 30–100 ppb. If solids are deposited inside the superheater, the tubes can become overheated, particularly if operated at high loads and high heat flux conditions. The convective superheaters are generally oversized at 100% load as explained earlier. The quantity of water spray is larger at higher load. In the radiant design, the steam temperature remains nearly flat over the load range because the radiant component of energy increases at lower loads and decreases at higher loads. Thus many radiant superheaters do not use a two-stage design. However, reviewing other concerns such as possible overheating of tubes and higher tube wall temperatures, the choice is left to the user. When demineralized water is not available, a portion of the saturated steam from the drum is taken and cooled in a heat exchanger, preheating the feedwater as shown in Fig. 3.18. The condensed water is then sprayed into the attemperator between the two stages of the superheater. Often, in order to balance the pressure drops in the two parallel paths, a resistance is introduced into each path or the exchanger is located vertically up, say 30–40 ft above the boiler, to provide additional head for the spray water control valve operation. Spraying downstream of the superheater for steam temperature control is not recommended, because the steam temperature at the superheater exit increases with load, thus increasing the superheater tube wall temperature, which can lead to tube failures. For example, if 800F is the final steam temperature desired, the steam temperature at the superheater exit may run as high as 875–925F, which will diminish the life of the tubes over a period of time. Also, the water droplets FIGURE 3.18 Steam temperature control methods. Copyright © 2003 Marcel Dekker, Inc. may not evaporate completely in the piping and the steam turbine could end up with water droplets and the solids present in the water, leading to deposits on turbine blades. Design Aspects Figures 3.19a and 3.19b show an inverted loop superheater commonly used in packaged boilers, and Fig. 3.19c shows a horizontal tube design with vertical headers. Superheaters operate at high tube wall temperatures; hence their design should be carefully evaluated. The convective superheater design located behind several rows of screen section operates at lower tube wall temperatures than the radiant design, though the steam temperatures may be the same. Figure 3.20 shows the results from a computer program for a superheater located very close to the furnace section and beyond several rows of screen tubes. Option a shows the results for a packaged boiler generating 150,000 lb=h of steam at 650 psig when a 14-row screen section is used. The gas temperature entering the superheater is 1628F. For the steam temperature of 758F, the superheater tube wall temperature is 856F. The surface area used is 1833 ft2. In option b, a nine-row screen section is used. The gas temperature entering the superheater is 1801F. The superheater tube wall temperature is 882F. However, owing to the higher log-mean temperature difference, the surface area required is smaller, namely 1466 ft2. It can be shown, as discussed under life estimation below, that the difference in the life of the superheater for a 26F difference for alloy steel tubes such as T11 can be several years. By the same token, one may wonder about the life of the radiant design with a gas inlet temperature of 2187F. Tube sizes are typically 1.5–2 in. OD, and materials used range from T11, T22, and T91 to stainless steels, depending upon steam and tube wall temperatures. Generally, bare tubes are used; however, I have designed a few packaged boilers, which are in operation in gas-fired boilers, using finned superheaters to make the design compact. Steam velocity inside the tubes ranges from about 50 ft=s at high steam pressure (say 1000–1500 psig) to about 150 ft=s at low pressure (150–200 psig). The turndown conditions and maximum tube wall temperatures determine the number of streams used and hence the steam pressure drop. In inverted loop superheaters, the headers are inside the gas path and are therefore protected by refractory. A few evaporator tubes are provided in the superheater region to ensure that steam blanketing does not occur at the mud drum and that steam bubbles can escape from the mud drum to the steam drum. Flow distribution through tubes is another concern with superheater design. If long headers are used, multiple inlets can reduce the nonuniformity in steam flow distribution through the tubes as shown in Fig. 3.21. Inlet and exit connections from the ends of headers should be avoided because they can Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.19a Inverted loop superheater arrangement. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. result in flow distribution problems. In arrangement 1, the inlet and exit connections are on opposite ends, causing the greatest difference in static pressure at the ends of the headers, and should be avoided. Arrangement 2 is better than 1 because the flow distribution is more uniform. However, arrange- ment 3 is preferred, because the central inlet and exit reduce the differential static pressure values by one-fourth, so the flow maldistribution is minimal. FIGURE 3.19b An inverted loop superheater. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.19c Horizontal tube superheater arrangement. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.20 Results from boiler program showing effect of screen section on superheater performance. Option a: More screen rows; option b: fewer screen rows. Copyright © 2003 Marcel Dekker, Inc. Two temperatures are of significance in the design of superheater tubes. One is the tube midwall temperature, which is used to evaluate the tube thickness per ASME code. (The published ASME stress values have increased during the last few years and therefore the latest information on stress values should be used in calculating the tube thickness.) The outer wall temperature determines the maximum allowable operating temperature, sometimes known as the oxidation limit. Table 3.6 gives typical maximum allowable temperatures for a few materials. One can vary the tube thickness to handle the design pressure, but if the outermost tube temperature gets close to the oxidation limit, we have to review FIGURE 3.21 Flow nonuniformity due to header arrangements. TABLE 3.6 Maximum Allowable Temperatures Material Composition Temp (F) SA 178A (erw) Carbon steel 950 SA 178C (erw) Carbon steel 950 SA 192 (seamless) Carbon steel 950 SA 210A1 Carbon steel 950 SA 210C Carbon steel 950 SA 213-T11 1.25Cr-0.5Mo-Si 1050 SA 213-T22 2.25Cr-1Mo 1125 SA 213-T91 9Cr-1Mo-V 1200 SA 213-TP304H 18Cr-8Ni 1400 SA 213-TP347H 18Cr-10Ni-Cb 1400 SA 213-TP321H 18Cr-10Ni-Ti 1400 SB 407-800H 33Ni-21Cr-42Fe 1500 Copyright © 2003 Marcel Dekker, Inc. the design. In large superheaters, different materials and tubes of different sizes may be used at different sections, depending on the tube midwall and outer wall temperatures. In all these calculations one has to consider the nonuniformity in gas flow, gas temperature across the cross section, and steam flow distribution through the tubes. Because of their shorter lengths, a few tubes could have higher flow and starve the longer tubes. Life Estimation High alloy steel tubes used in superheaters and reheaters, unlike carbon steel, fail by creep rupture. Creep refers to the permanent deformation of tubes that are operated at high temperatures. Carbon steel tubes operate in the elastic range where allowable stresses are based on yield stresses, whereas alloy tubes operate in the creep-rupture range, where allowable stresses are based on rupture strength. The life of superheater tubes is an important datum that helps plant engineers plan tube replacements or schedule maintenance work. When a new superheater tube is placed in service, it starts forming a layer of oxide scale on the inside. This layer gradually increases in thickness and also increases the tube wall tempera- ture. Therefore, to predict the life of the tubes, information on the corrosion or the formation of the oxide layer is necessary. The corrosion of oxide formation also reduces the actual thickness of the tubes and increases the stresses in the tubes over time even if the pressure and temperature are the same. The data on oxide formation were once obtained by cutting tube samples and examining them but are now obtained through nondestructive methods. There are also methods to relate the oxide layer thickness with tube mean wall temperatures over a period of time. Creep data are available for different materials in the form of the Larson Miller parameter, LMP. This relates the rupture stress value to temperature T and the remaining lifetime t, in hours. LMP ¼ ðT þ 460Þð20 þ log tÞ Every tube in operation has an LMP value that increases with time. LMP can be related to stress values and the relationship then used to predict remaining life. However, there are charts that give what is called the minimum and the average rupture stress versus LMP, and one can compute different life times with the different values. Also, it can be seen that even a few degrees difference, say 10F in metal temperatures, can change the lifetime by a large amount, which shows how complex and difficult it is to interpret the results. Figure 3.22 shows the relationships between LMP and minimum rupture stress values for T11 and T22 materials. Copyright © 2003 Marcel Dekker, Inc. Example 7 Assume that a superheater of T11 material operates at 1000F and at a hoop stress of 6000 psi. What is the predicted time to failure? From Fig. 3.22, the LMP at 6000 is 36,800. Solution: From the above equation, we can see that 36; 800 ¼ ð1460Þð20 þ log tÞ; or t ¼ 160;500 h If a tube had operated at this temperature for 50,000 h, its life consumed would be 50,000=160,500¼ 0.31, or 0.69 of its life would remain. If after this period of 50,000 h, it operated at, say, 1020F and at the same stress level, then 36;800 ¼ ð1480Þð20 þ log tÞ; or t ¼ 73;250 h and the number of operating hours at this temperature would be 0.69 73,250¼ 50,728 h. One can see from the above how sensitive these numbers are to tempera- tures and stress values. Hence we have to interpret the results with caution backed up by operational experience. Simplistic approaches to replacement of tube bundles are not recommended. It should also be noted that if the average rupture stress is used instead of the minimum value, the lifetime would be much higher, casting more uncertainty in these calculations. FIGURE 3.22 Larson–Miller parameters for T11 and T22 materials. Copyright © 2003 Marcel Dekker, Inc. ECONOMIZERS Economizers are used as heat recovery equipment in packaged boilers instead of air heaters because of NOx concerns as discussed in Chapter 4. They are also less expensive and have lower gas pressure drops across them. Economizers for gas firing typically use serrated fins at four to five fins per inch. For distillate fuel, about 4 fins=in, solid fins are preferred. For heavy oil, bare tubes or a maximum of 2–3 fins=in. are used, depending upon the dirtiness of the flue gas and the ash content of the fuel. Economizers are generally of vertical gas flow and counterflow configura- tion with horizontal tubes as shown in Fig. 3.23. The water-side velocity ranges from 3 to 7 ft=s. Small packaged boilers, below 40,000 lb=h capacity, use circular economizers that can be fitted into the stack. Another variation is the horizontal gas flow configuration with vertical headers and horizontal tubes. Generally, steaming in the economizer is not a concern, as discussed earlier. Feedwater temperatures of 230–320F are common, depending on acid dew point concerns. The feedwater is sometimes preheated in a steam–water exchanger if the deaerator delivers a lower feedwater temperature than that desired to avoid acid corrosion in the case of oil-fired boilers. BOILER PERFORMANCE ASPECTS Plant engineers are interested in knowing how a given boiler performs at various loads. The variables affecting its performance are the fuel, amount of excess air, FGR rate, and steam parameters. Tables 3.4 and 3.7 show how boiler performance varies with load on gas and oil firing. Figure 3.24 shows the results in graph form. The following observations can be made: 1. As the load increases, the boiler exit gas temperature increases. This is due to the larger flue gas mass flow transferring energy to a given heating surface. The water temperature leaving the economizer is higher at loads owing to the higher gas temperature entering the economizer. The approach point (difference between saturation and water temperature entering evaporator) is lower at higher loads. Steaming in the economizer is not a concern in steam generators because the approach point is quite large at full load and increases at lower loads. The ratio of gas flow to steam generation is maintained at 1.2–1.3 at various loads. Hence the economizer does not absorb more energy at low loads as in the case of HRSGs. 2. The boiler efficiency increases as the load increases, peaks at about 50– 70% of load, then drops off. The two major variables affecting the heat losses are the casing heat losses and heat loss due to flue gases. Q6.24 discusses this calculation. As the load increases, the flue gas heat losses Copyright © 2003 Marcel Dekker, Inc. FIGURE 3.23a Economizer in a packaged boiler. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. increase due to the higher exit gas temperature. The casing loss decreases as a percentage but, as explained in Q6.24, in terms of Btu=h it remains the same because the evaporator operates at saturation temperature, so heat losses in Btu=h are unaffected by boiler load except if ambient temperature or wind velocity changes. Thus the combination of these losses results in a parabolic shape for efficiency as a function of load. 3. The steam temperature generally increases with load owing to the convective nature of the superheater. If a radiant design were used, it would decrease slightly at higher loads. 4. It may also be seen that the gas temperature leaving the evaporator decreases as the load decreases. If an SCR is used between the evaporator and the economizer, the gas temperature should be main- tained in the range of typically 650–780F; hence one may have to use a gas bypass system to obtain a higher gas temperature at low loads. Chapter 4 shows the arrangement of dampers to achieve this purpose. 5. The steam temperature on oil firing is lower than that in gas firing. This is due to the better absorption of energy from the oil flames in the FIGURE 3.23b Photo of an economizer. (Courtesy of ABCO Industries, Abilene, TX.) Copyright © 2003 Marcel Dekker, Inc. TABLE 3.7 Boiler Performance—Oil Firing Load (%) 25 50 75 100 Boiler duty, MM Btu=h 28.94 58.26 89.03 118.71 Excess air, % 30 15 15 15 Fuel input, MM Btu=h 32.98 65.95 101.25 135.9 Heat rel rate, Btu=ft3 h 15,266 30,531 46,875 62,918 Heat rel rate, Btu=ft2 h 28,188 56,376 86,554 116,176 Steam flow, lb=h 25,000 50,000 75,000 100,000 Steam temp, F 694 710 750 750 Economizer exit water temp, F 324 329 350 368 Boiler exit gas temp, F 526 588 671 748 Economizer exit gas temp, F 254 269 296 325 Air flow, lb=h 32,064 56,728 87,096 116,903 Flue gas, lb=h 33,731 60,061 92,212 123,771 Dry gas loss, % 3.95 3.83 4.36 4.95 Air moisture loss, % 0.1 0.1 0.11 0.13 Fuel moisture loss, % 6.58 6.62 6.69 6.77 Casing loss, % 1.2 0.6 0.4 0.3 Margin, % 0.5 0.5 0.5 0.5 Efficiency, % HHV 87.67 88.35 87.93 87.35 Efficiency, % LHV 93.67 94.39 93.95 93.33 Furnace back pressure, in. WC 0.8 2.45 5.81 10.76 Steam pressure¼ 500psig, oil firing. HHV¼ 19,727; LHV¼ 18,463Btu=lb. Flue gas analysis (vol%): CO2 ¼ 10:76, H2O ¼ 11:57, N2 ¼ 73:63, O2 ¼ 2:51. FIGURE 3.24 Boiler performance versus load. Copyright © 2003 Marcel Dekker, Inc. furnace, which results in a lower furnace exit gas temperature and lower gas temperature at the superheater in oil firing. Hence the steam temperature is lower. However, if we wanted to maintain the same steam temperature on both oil and gas firing, we would have to size the superheater so that it makes the steam temperature in the oil-firing case and then control it in gas firing by attemperation. Performance Without an Economizer If we look at Table 3.4 for performance of a boiler at, say, 100% load, we see that the gas temperature leaving the evaporator is 739F and leaving the economizer it is 327F. Now if the economizer is removed from service, can we assume that the exit gas temperature will still be 739F? The answer is No, for the following reasons: 1. The boiler efficiency drops significantly, by at least (7397 327)=40¼ 10.3%. Hence the efficiency will be at best 83.667 10.3¼ 73.36% HHV. 2. The boiler fuel input increases by this ratio. The new heat input is (118.71=0.7336)¼ 161.8MM Btu=h versus (118.71=0.8366)¼ 141.9MM Btu=h. Hence the flue gas flow, which is proportional to heat input, will be higher by 161.8=141.9¼ 1.14 or 14%, or about 1.14 125,246¼ 142,800 lb=h. 3. The furnace heat input and heat release rate will also be higher due to the lower efficiency and hence higher furnace exit gas temperature. The combination of higher gas flow and higher gas inlet temperature to the convection bank will increase the exit gas temperature from the evaporator from 739F to a slightly higher value. Therefore another iteration will have to be performed to arrive at the exit gas temperature based on the revised efficiency and fuel input. The exit gas temperature could be close to 770–780F. 4. Because of the larger flue gas flow and higher operating temperature in the evaporator bank, the gas pressure drop will also be higher; it could be as much as in the earlier case or even more. Hence, the assumption that removing the economizer will reduce the total gas pressure drop is incorrect. One has to do the performance calculations before arriving at any conclusion. Why the Economizer Does Not Steam in Packaged Boilers Unlike HRSGs, packaged boilers, fortunately, do not have to deal with the issue of steaming. The reason is illustrated in Fig. 3.25, which shows the temperature Copyright © 2003 Marcel Dekker, Inc. profiles of the economizer of the boiler whose performance is given above and an HRSG. Because of the small ratio of gas to water flow in packaged boilers, the temperature drop of the flue gas has to be large for a given water temperature increase. If the water temperature increases by, say, 145F, the gas temperature drop is given by 1:23 0:286 ðT1 T2Þ ¼ 145 or T1 T2 ¼ 412F whereas in an unfired HRSG, the gas temperature drop of only 105F accom- plishes a water temperature increase of 248F! Thus it is easy for the water to reach saturation temperature in HRSGs. Thus in spite of the fact that the gas entering temperature is quite large in packaged boilers (due to the high furnace exit gas temperature), the water temperature does not increase significantly. If the water temperature approach is large at 100% load, it will be even larger at partial loads, because the gas temperature entering the economizer decreases. Performance with Oil Firing Steam generators have been fired with both distillate fuel oils and residual oils. The design of the boiler does not change much for distillate oil firing compared to gas firing. The fouling factor used is moderately higher, 0.003–0.005 ft2 h F=Btu, compared to 0.001 ft2 h F=Btu for gas firing; rotary soot blowers located at either end of the convection section are adequate for cleaning the surfaces for distillate oil firing. With heavy fuel oils, retractable soot blowers are required. Economizers FIGURE 3.25 Economizer temperature pick-up in boiler versus HRSG. Copyright © 2003 Marcel Dekker, Inc. also use rotary blowers in oil-fired applications. Solid fin tubes of a fin density of three or four per inch may be used if distillate fuels are used, but if heavy oil is fired it is preferable to use bare tubes or at best 2–3 fins=in. The emissions of NOx will be higher on the basis of fuel-bound nitrogen, because it can contribute to nearly 50% of the total NOx. Flue gas recirculation has less effect on NOx in oil firing than in gas firing. With residual fuel oil firing, there are several aspects to be considered. 1. High temperature corrosion due to the formation of salts of sodium and vanadium in the ash has been a serious problem in with heavy oil boilers fired. The furnace exit region is a potentially dirty zone prone to deposition of molten ash on heating surfaces. The use of superheaters in such regions presents serious performance concerns. Retractable steam soot blowers are required, with access lanes for cleaning. Tubes should preferentially be widely spaced at the gas inlet region to avoid bridging of tubes by slag. Vanadium content in fuel oil ash should be restricted to about 100 ppm to minimize corrosion potential. 2. Superheater materials used in heavy oil firing applications should consider the high temperature corrosion problems associated with sodium and vanadium salts. The metallurgy of the tubes should be T22 or even higher if the tube wall temperature exceeds 1000F. A large corrosion allowance on tube thickness is also preferred. This is yet another reason for preferring a convective superheater design to a radiant superheater. 3. Steam temperatures with oil firing will be lower than on gas firing as discussed above. 4. Furnace heat flux will be higher in oil firing than in gas firing. Therefore one has to check the circulation and the furnace design. 5. One of the problems with firing a fuel containing sulfur is the formation of sulfur dioxide and its conversion to sulfur trioxide in the presence of catalysts such as vanadium, which is present in fuel oil ash. Sulfur trioxide combines with water vapor to form sulfuric acid vapor, which can condense on surfaces whose temperature falls below the acid dew point. Q6.25 illustrates the estimation of dew points of various acid vapors. Sulfuric acid dew points can vary from 200 to 270F depending on the amount of sulfur in the fuel. If the tube wall temperature of the economizer or air heater falls below the acid dew point, condensation and hence corrosion due to the acid vapor are likely. I have seen a few specifications where a parallel flow arrange- ment was suggested for the economizer to minimize acid dew point corrosion. Because the feedwater temperature governs the tube wall temperature and not the flue gas temperature, only maintaining a high Copyright © 2003 Marcel Dekker, Inc. water temperature avoids this problem, as shown in Q6.25c. One could use steam to preheat the feedwater or use the water from the exit of the economizer to preheat the incoming water in a heat exchanger. Experience and research show that acid corrosion potential is maxi- mum not at the dew point but at slightly lower values, about 15–20C below the dew point. Hence one may use a feedwater temperature even slightly lower than the dew point of the acid vapor in order to recover more energy from the waste gas stream. In waste heat boiler econo- mizers, other acid vapors such as hydrochloric acid or hydrobromic acid may be present. The dew points of these are much lower than that of sulfuric acid, as discussed in Q6.25, so care must be taken in the design of economizers or air heaters in heat recovery applications. Table 3.7 shows the boiler performance with distillate oil firing. The efficiency on LHV basis is nearly the same as for gas firing, but on HHV basis there is a difference. The flue gas analysis with 15% excess air is shown. The flue gases have less water vapor but more carbon dioxide than flue gases from natural gas combustion. Effect of FGR on Boiler Performance Flue gas recirculation is widely used as a method of NOx control because it reduces the flame temperature and thus lowers NOx formation as discussed in Chapter 4. The effect of FGR on boiler performance is quite significant. Not only is the gas temperature profile across the boiler different, but the steam tempera- ture and gas pressure drop are also affected. Table 3.8 shows the performance of a 150,000 lb=h boiler with and without FGR. The following points may be noted: 1. The flue gas quantity increases with FGR; hence the backpressure increases at all loads. 2. The steam temperature is higher with FGR in both 100% and 50% load cases, but the difference is greater at low loads. 3. The furnace exit gas temperature is lower with FGR, and the gas temperature across the superheater is higher at 50% load than at 100%. Thus load plays a big role in the temperature profiles. 4. The efficiency naturally drops due to the higher stack gas temperature at both 100% and 50% loads. Relating FGR and Oxygen in the Wind-Box Flue gas recirculation affects the oxygen in the wind-box by diluting it. One may measure the oxygen values to evaluate the FGR rate used. Copyright © 2003 Marcel Dekker, Inc. Example 8 A boiler firing natural gas at 15% excess air uses 119,275 lb=h of combustion air, and about 14,000 lb=h of flue gases is recirculated. Determine the oxygen levels in the wind-box. Let us assume that the air is dry and is 77% by weight nitrogen and 23% oxygen. Then the amount of nitrogen in air¼ 0.77 119,275¼ 91,842 lb=h, and that of oxygen¼ 27,433 lb=h. The flue gas analysis (vol%) is CO2 ¼ 8:29;H2O ¼ 18:17;N2 ¼ 71:07, and O2 ¼ 2:47. To convert to percent by weight (wt%) basis, first obtain the molecular weight: MW ¼ ð8:29 44 þ 18:17 18 þ 71:07 28 þ 2:47 32Þ=100 ¼ 27:61 % CO2 ¼ 8:29 44=27:61 ¼ 13:21 Similarly, H2O ¼ 11:84 wt%;N2 ¼ 72:07, and O2 ¼ 2:88. TABLE 3.8 Effect of FGR on Boiler Performance Load (%) 100 100 50 50 Excess air, % 15 15 15 15 FGR, % 0 15 0 15 Combustion, temp, F 3,230 2,880 3,230 2,880 Furnace exit temp, F 2,350 2,188 2,007 1,956 Gas temp to superheater, F 1,695 1,630 1,323 1,334 Gas temp to evaporator, F 1,250 1,240 944 973 Gas temp to economizer, F 630 645 543 555 Gas temp leaving economizer, F 300 315 263 270 Flue gas flow, lb=h 185,500 215,000 88,900 104,000 Efficiency, % HHV 84.26 83.9 85.1 84.9 Steam flow, lb=h 150,000 150,000 75,000 75,000 Steam temp, F 748 756 686 711 Economizer exit water temp, F 338 355 318 333 Boiler backpressure, in. WC 6.2 7.8 2.0 2.5 Feedwater temp, F 228 228 228 228 Fuel: standard natural gas; 1% blowdown; steam pressure¼650psig. Copyright © 2003 Marcel Dekker, Inc. The individual constituents in the mixture of 14,000 þ 119,275¼ 133,275 lb=h of gases are CO2 ¼ 0:1321 14;000 ¼ 1849:4 lb=h H2O ¼ 0:1184 14;000 ¼ 1658 lb=h N2 ¼ 91;843 þ 0:7207 14;000 ¼ 101;922 lb=h O2 ¼ 27;433 þ 0:0288 14;000 ¼ 27;836 lb=h Converting this to percent by volume basis as we did earlier, we have CO2 ¼ 0:9 vol %;H2O ¼ 1:98;N2 ¼ 78:37; and O2 ¼ 18:75 SOOT BLOWING Soot blowing is often resorted to in coal-fired or heavy oil–fired boilers. In packaged boilers, both steam and air have been used as the blowing media, and both have been effective with heavy oil firing. Rotary blowers are sometimes used with distillate oil firing. Steam-blowing systems must have a minimum blowing pressure of 170–200 psig to be effective. The steam system must be warmed up prior to blowing to minimize condensation. The steam must be dry. Increasing the capacity of a steam system is easier than increasing that of an air system. With an air system, the additional capacity of the compressor must be considered. Also, because steam has a higher heat transfer coefficient than air, more air is required for cooling the lances in high gas temperature regions compared to steam. Moisture droplets in steam can cause erosion of tubes, and often tube shields are required to protect the tubes. The intensity of the retractable blower jet is more than that of the rotary blower jet, and its blowing radius is larger, thus cleaning more surface area. However, one must be concerned about the erosion or wear on the tubes. Sonic cleaning has been tried on a few boilers. In this system, low frequency high energy sound waves are produced when compressed air enters a sound generator and forces a diaphragm to flex. The resulting sound waves cause particulate deposits to resonate and dislodge from the surfaces. Once dislodged, they are removed by gravity or by the flowing gases. Typical frequencies range from 75 to 33Hz. Sticky particles are difficult to clean. The nondirectional nature of the sound waves minimizes accumulation in blind spots where soot blowers are ineffective. Piping work is minimal. Sonic blowers operate on plant air at 40–90 psi and sound off for 10 s every 10–20min. Copyright © 2003 Marcel Dekker, Inc. WATER CHEMISTRY, CARRY OVER, AND STEAM PURITY Good water chemistry is important for minimizing corrosion and the formation of scale in boilers. Steam-side cleanliness should be maintained in water tube as well as fire tube boilers. Plant engineers should do the following on a regular basis: 1. Maintain proper boiler water chemistry in the drum according to ABMA or ASME guidelines by using proper continuous blowdown rates. The calculation procedure for the blowdown rate based on feedwater and boiler water analysis is given in Q5.17. 2. Ensure that the feedwater analysis is fine and that there are no sudden changes in its conductivity or solids content. 3. Check steam purity to ensure that there are no sudden changes in its value. A sudden change may indicate carryover. 4. Watch superheated steam temperatures, particularly in boilers with large load swings. If slugs of water get carried into the steam during large load swings, the deposits are left behind after evaporation, potentially leading to tube failure. An indication of slugging, which is likely in boilers with small drums, is a sudden decrease in steam temperatures due to entrainment of water in the steam. In the process of evaporating water to form steam, scale and sludge deposits form on the heated surfaces of a boiler tube. The chemical substances in the water concentrate in a film at the evaporation surface; the water displacing the bubbles of steam readily dissolves the soluble solids at the point of evaporation. Insoluble substances settle on the tube surfaces, forming a scale and leading to an increase in tube wall temperatures. Calcium bicarbonate, for example, decomposes in the boiler water to form calcium carbonate, carbon dioxide, and water. Calcium carbonate has limited solubility and will agglomerate at the heated surface to form a scale. Blowdown helps remove some of the deposits. Calcium sulfate is more soluble than calcium carbonate and will deposit as a heat-deterrent scale. Most scale-forming substances have a decreasing solubility in water with an increase in temperature. In boilers that receive some hardness in the makeup water, deposits are generally compounds of calcium, sulfate, silica, magnesium, and phosphate. Depending on tube temperatures and heat flux and the solubility of these compounds as a function of temperature, these compounds can form deposits inside the boiler tubes. These scales, along with sludge and oils, form an insulating layer inside tubes at locations where the heat flux is intense. Alkalinity and pH of the water also affect the scale formation. Salts such as calcium sulfate and calcium phosphate deposit preferentially in hot regions. Boilers are consid- ered generally clean if the deposits are less than 15mg=cm2. Boilers having more than 40mg=cm2 are considered very dirty. The least soluble compounds deposit Copyright © 2003 Marcel Dekker, Inc. first when boiling starts. Calcium carbonate deposits quickly, forming a white friable deposit. Magnesium phosphate is a binder that can produce very hard, adherent deposits. Insoluble silicates are present in many boilers. The presence of sodium hydroxide, phosphate, or sulfate may be considered proof that complete evaporation has occurred in the tubes, because these are easily soluble salts. Sludge or easily removable deposits accumulate at the bottom of the tubes in the mud drum and should be removed by intermittent blowdown, generally once per shift. Based on conductivity readings, the frequency may be increased or decreased. Continuous blowdown is usually taken from the steam drum a few inches from above the waterline, where the concentration of solids is the highest. Any boiler water treatment program should be reviewed with a water chemistry consultant, because this program can vary on a case-to-case basis. Generally the objective is to add chemicals to prevent scale formation caused by feedwater hardness constituents such as calcium and magnesium compounds and to provide pH control in the boiler to enhance maintenance of a protective oxide film on boiler water surfaces. There are methods such a phosphate-hydroxide, coordinated phosphate, chelant treatment, and polymer treatment methods. In medium and low pressure boilers, all these methods have been used. Carryover of impurities with steam is a major concern in boilers having superheaters and also if steam is used in a steam turbine. Carryover results from both ineffective mechanical separation methods and vaporous carryover of certain salts. Vaporous carryover is a function of steam density and can be controlled only by controlling the boiler water solids, whereas mechanical carryover is governed by the efficiency of the steam separators used. Total solids carryover in steam is the sum of mechanical and vaporous carryover of impurities. The steam purity requirements for saturated steam turbines are not stringent. Because the saturated steam begins to condense on the first stage of the turbine, water-soluble contaminants carried with the steam do not form deposits. Unless the steam is contaminated with solid particles or acidic gases, its purity does not significantly affect the turbine performance. However, there can be erosion concerns due to water droplets moving at high speeds. With superheated steam, steam purity is critical to the turbine. Salts that are soluble in superheated steam may condense or precipitate and adhere to the metal surfaces as the steam is cooled when it expands. Deposition from steam can cause turbine valves to stick. Reduced efficiency and turbine imbalance are the other concerns. Deposition and corrosion occur in the ‘‘salt zone’’ just above the saturation line and on surfaces in the wet steam zone. The solubility of all low volatility impurities such as salts, hydroxides, silicon dioxide, and metal oxides decreases as steam expands in the turbine and is lowest at the saturation line. The moisture formed has the ability to dissolve most of the salts and carry them downstream. The critical region for deposition in turbines operating on super- heated steam is the blade row located just upward of the Wilson line. Copyright © 2003 Marcel Dekker, Inc.