01 POWER ISLAND / 01 CCPP / V. Ganapathy-Industrial Boilers and HRSG-Design (2003)

FIGURE 3.9 Superheater piping arrangement for (a) low and (b) high pressure operation.

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

TABLE 3.3 Boiler Performance at Low and High Steam Pressurea

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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 200,000 Btu=ft2 h for oiland gasfired 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 30,000 Btu=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 130,000 Btu=ft3 h for typical packaged oiland 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

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FIGURE 3.10 Furnace construction—membrane wall, tangent tube, and refractory wall.

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.

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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 Ape1e2sðTg4 Tw4Þ ¼ Wf LHV Wghe

FIGURE 3.11 Relating fin tip temperature to heat flux in membrane wall furnace.

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FIGURE 3.12 Furnace outlet temperature for gas and oil firing.

where

Q ¼ 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 s ¼ radiation constant

e1; e2 ¼ emissivity of flame and wall, respectively LHV ¼ 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 steam ¼ 118.71 MM Btu=h. Flue gas flow ¼ 125,246 lb=h at 100% load.

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TABLE 3.4 Boiler Performance—Gas Firinga

aSteam pressure 500 psig; feedwater 230 F, blowdown 1%, amb temp 80 F; RH 60%, fuelstandard 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 area ¼ 1169 ft2, furnace width ¼ 7.5 ft, length ¼ 32 ft, height ¼ 9 ft.

The net heat input to the furnace is

118:71 00::9273992 ¼ 127 MM Btu=h

where 0.992 ¼ 1 7 heat losses, and 0.9273 is the boiler efficiency on LHV basis.

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The furnace exit gas temperature from Fig. 3.12 is 2235 F. It may be shown that the enthalpy of the flue gases at 2235 F 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 44:2 MM Btu=h.

The average heat flux based on projected area is

44:2 106 ¼ 37;810 Btu=ft2 h

1169

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 ¼ OD of furnace tubes

h ¼ membrane 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

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,

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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 h F, fin base temperatures will be a few degrees higher than saturation temperature, assuming that tube-side fouling is minimal.

tt ¼ fin tip temperature, F

h ¼ membrane height, in. (see Fig. 3.11) m ¼ ðhgC=KAÞ0:5

where

hg ¼ gas-side heat transfer coefficient, Btu=ft2 h F

C ¼ perimeter of fin cross section ¼ 2b þ L in. (for heating from one side) where b ¼ fin thickness and L ¼ fin length or furnace length

K ¼ fin thermal conductivity, Btu=ft h F A ¼ cross-section of fin ¼ 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 2200 F. The gas-side heat transfer coefficient is estimated to be 30 Btu=ft2 h F. Fin height ¼ 0.5 in. fin thickness ¼ 0.375 in. Fin base temperature is 600 F. Thermal conductivity of fin is 20 Btu=ft h F. Determine the fin tip temperature.

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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– 8000 Btu=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.

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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 70,000 Btu=ft2 h, while critical heat flux could be in excess of 250,000 Btu=ft2 h. However, one has to perform circulation calculations on all the parallel circuits in the boiler, particularly 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.

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