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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.
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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.