EPA-600/2-77-190
September 1977
Environmental Protection Technology Series
EFFECTS OF COMBUSTION MODIFICATIONS
FOR NOX CONTROL ON UTILITY
BOILER EFFICIENCY AND
COMBUSTION STABILITY
Industrial Environmenta.l Research Laboratory
Office of Research and Development
U.S. Environmental Protection Agency
Research Triangle Park, North Carolina 27711
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RESEARCH REPORTING SERIES
Research reports of the Office of Research and Development, U.S. Environmental Protection
Agency, have been grouped into five series. These five broad categories were established to
facilitate further development and application of environmental technology. Elimination of
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2. Environmental Protection Technology
3. Ecological Research
4. Environmental Monitoring
5. Socioeconomic Environmental Studies
This report has been assigned to the ENVIRONMENTAL PROTECTION TECHNOLOGY
series. This series describes research performed to develop and demonstrate instrumenta-
tion, equipment, and methodology to repair or prevent environmental degradation from point
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required for the control and treatment of pollution sources to meet environmental quality
standards
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This document is available to the public through the National Technical Information Service.
Springfield, Virginia 22161.
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EPA-600/2-77-190
September 1977
EFFECTS OF COMBUSTION MODIFICATIONS
FOR NOX CONTROL ON UTILITY BOILER
EFFICIENCY AND COMBUSTION STABILITY
by
Owen W. Dykema
The Aerospace Corporation
Environment and Energy Conservation Division
El Segundo, California 90245
Grant No. R803283-02
ROAP No. 21ADG-089
Program Element No. 1AB014
EPA Project Officer: Robert E. Hall
Industrial Environmental Research Laboratory
Office of Energy, Minerals, and Industry
Research Triangle Park, N.C. 27711
Prepared for
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Research and Development
Washington, D.C. 20460
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PREFACE
This study is a continuation of previous Aerospace Corporation
efforts reported in “Analysis of Test Data for NO Control in Gas- and Oil-
Fired Utility Boilers” (EPA-650/2 -75_Oj2, January i975) and in “Analysis of
Test Data for NO Control in Coal-Fired Utility Boilers” (EPA-600/2_76..274,
October 1976). The data published in the earlier report was used as a basis
for the analyses reported herein.
This study, as well as the two previous studies, was conducted
for the U. S. Environmental Protection Agency, Combustion Research Branch,
Industrial Environmental Research Laboratory, Research Triangle Park,
North Carolina, during the second year of a three-year continuing grant.
(The first study was conducted under a separate EPA Grant No. R-802366 for
this same EPA office. ) The first two studies concerned the effects of com-
bustion modifications on NO emissions, whereas the effort reported here
concerns possible side effects of these combustion modifications on plant
efficiency and combustion stability. In the third and final year of the current
grant, all of the previous studies will be combined and simplified into an over-
all design model capable of indicating design modifications to minimize NO
emissions while maintaining low emissions of other air pollutants, high plant
efficiency and stable combustion.
A brief introduction is contained in Section I. Conclusions and
recommendations are contained in Section II while Section III contains a brief
summary of results. Section IV describes the analysis of the effects of corn-
bustion modifications on plant efficiency. Section V describes the development
and application of a model of air-side feed system coupled modes of combus-
tion instability in utility boilers.
11
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CONTENTS
PREFACE
FIGURES
TABLES
NOMENCLATURE
ACKNOWLEDGMENTS
I. INTRODUCTION 1
II. CONCLUSIONS AND RECOMMENDATIONS 3
2. 1 Conclusions 3
2. 2 Recommendations 4
III. SUMMARY 6
3. 1 Plant Efficiency 6
3. 2 Combustion Stability 8
IV. PLANT EFFICIENCY ii
4. 1 Combustion Modifications for NO Control 12
x
4. 2 Efficiency Losses 13
4. 3 Results 16
4. 3. 1 Efficiency Corrections 17
4. 3. 2 Incomplete Combustion 18
LIL
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CONTENTS (Continued)
4. 3. 3 Effects of Load Variations 19
4. 3. 4 Other Possible Effects 23
V. COMBUSTION STABILITY 24
5. 1 Analytical Modeling .24
5. 1. 1 Burner Air Flow Response Z4
5. 1. 2 Furnace Pressure .,.. • .,. 27
5. 1. 3 Feed System Coupled Instability . . . • .. • ,. ,... . .. 31
5.2 Results . .. 31
5. 2. 1 Validity of the Analysis • - 33
5. 2. 2 Further Parametric Calculations • . . . . 48
REFERENCES 58
APPENDIX A. EFFICIENCY LOSSES 59.
APPENDIX B. ANALYSIS OF A COMBUSTION-AIR
FEED SYSTEM-COUPLED MODE 0F’ -
INSTABILITY IN A UTILITY BOILER 71
iv
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FIGURES
1. Effects of boiler load variations on efficiency and
NO emissions 22
x
2. Model of the dynamics of a burner in a utility boiler 25
3. Active burner flow and combustion model 29
4. Model of an air feed system coupled mode of
combustion instability in a utility boiler 32
5. Typical liquid rocket engine model of a low frequency
(chug) feed system coupled mode of combustion
instability (simplified from Figure 4) 34
6. Schematic of the boiler analyzed-defining the
acoustic mode directions, lengths and reflection
surfaces 36
7. Comparison of results of the current analysis with
those of conventional rocket engine analysis 38
8. Comparison of analytical predictions with
experimental observations 40
9. Variation of the maximum gain within the low and
intermediate frequency ranges for burners located
at various vertical positions 42
10. Effects of acoustic wave reflection efficiency 45
ii. Effects of the fraction of combustion completed
within an active burner on combustion stability 47
12. Analytical expressions used to approximate the
function F(r) 50
13. Burner air/fuel equivalence ratio with/without
flames in burners 52
V
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FIGURES (Continued)
14. Effects of the degree of burners-out-of-service on
instability at the unstable frequency. 53
15. Effects of the degree of burners-out-of-service on
instability at various frequencies 54
16. Experimental C0 2 /0 2 data for all firing configurations
in one boiler type ‘. 65
vi
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TABLES
1. Effects of Combustion Modifications On Plant Efficiency
Losses Due to Incomplete Combustion 20
2. Geometry and Full Load Operating Conditions 35
3. Anti-Nodes and Nodes in the Vertical Resonant Modes
oftheFurnaceCavity 43
vii
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NOMENCLATURE
FORTRAN NOTATION
A ’R = Air-fuel ratio, by weight.
AK1, AK2, AK3 = Intermediate constants specific to a given fuel, defined by
Eqs. (A—12), (A—13), (A—i4).
AK3L, AK3H = Intermediate constants specific to a given fueland to the Low
and High heats of combustion.
CES, CESI, CESZ Constants in the derived function relating sensible heat
losses to AFR and FLOAD. Eq. (A -24).
ESC = Efficiency data after correction for losses due to electrical equipment,
to water vapor and sensible heat in the flue gases and to incomplete
combustion.
LOAD (FLOAD) = Electrical load generated by the plant, megawatts (fraction
of rated load).
NUM (DENOM) = Intermediate functions used to describe the numerator
(denominator) of the complex function SMTD. Eqs. (A-3), (A-4),
(A -5).
QLIC, QLH2O, QLSEN = Heat losses due to incomplete combustion, to water
vapor in the flue gases and to sensible heat in the flue gases, re-
spectively. Eqs. (A-17), (A-i8), (A -24).
QHHC (QHHCM) = High heat of combustion, including allowance for incom-
plete combustion (maximum high heat of combustion, with complete
cornbu stion).
SMTD = The total number of moles (dry) in the combustion products.
WFT = The total weight flow rate of fuel.
* = Designates the FORTRAN ‘ t multiply”.
VLIL
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CHEMICAL COMPOUND NOTATION
C, H, 0, N, S = Designates the moles (a, b, c, d, e, respectively) of atomic
Carbon, Hydrogen, Oxygen, Nitrogen, and Sulfur, respectively,
in a given fuel “molecule”. Eq. (A- 1).
CO, CO 2 , (COZth), H 2 0, N 2 , NO, °2, SO 2 = Designates molecular Carbon
Monoxide, Carbon Dioxide (theoretical for complete combustion),
Water, Nitrogen, Nitric Oxide, Oxygen, and Sulfur Dioxide,
respectively.
a, b, c, d, e = Designate moles of each atomic species in a given fuel mole-
cule. C, H, 0, N, S, above and Eq. (A-I).
ARABIC NOTATION
A, B, C, D, E, F, G, H = Designates the moles of C, CO, CO 2 , H 2 O, O ,
N , NO, and SO , respectively, in the products of combustion.
Eq. (A-i).
A’, B’, D’ = Intermediate constants in the development of the function for
QLSEN. Eqs. (A-20), (A-21).
Ab = Cross-sectional flow area for air in a burner.
C = Capacitance within a burner. Eq. (B-21).
Cpfg = Specific heat at constant pressure of the boiler flue gases.
Ch = Fraction of combustion completed within a burner.
FA, FB, FC, FD, FE, FF, FG = Functions used to simplify Eq. (B-73).
Eqs. (B-78) through (B-84).
Fa(I) (1 = 1 - 5) = Arbitrary functions used to schematically designate the
furnace pseudo-acoustics, in five of the six directions, in Figure 4.
Fb = A complex function partially describing the time-varying response of the
total weight flow rate issuing from a burner to changes in the furnace
pressure at the burner exit. Eqs. (B-39), (B-40), (B-41).
F , Ffe, Frr, F = Functions used to simplify Eq. (B-73). Eqs. (B-85)
through (B-88).
F = A complex function describing the time-varying response of the furnace
p pressure at a burner exit to changes in the total weight flow rate
issuing from the burner. Eqs. (B-73) through (B-77).
ix
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F(r) = A function describing changes in local furnace pressures resulting
from changes in the local combustion air-fuel ratio (through changes
in the ‘combustion temperature and molecular weight of the corn-
bustion products).
Fj, F 2 , F 3 = Functions used to simplify Eqs. (B-42), (B -43). Eqs. (B-46),
(B-47), (B_48).
Kj(Kiavg) = Constant describing damping during acoustic wave travel and the
efficiency of wave reflection at solid boundaries (an average of K 1
in the six directions of wave travel). Eq. (B-67).
K . The efficiency of reflection of acoustic waves at :solid boundaries.
K 3 = An empirical coefficient relating the fraction of combustion cornpleted
within a burner (Ch) to the weight flow rate of air through the
burner (“b)• Eq. (1).
L = Inertance of the air within a burner.
Lb = Length of a burner.in the flow direction.
Lh = Distance from a burner exit to a solid boundary, in all six directions.
M(MWf) = Molecular weight (of a fuel “molecule’ t ).
Mag = The magnitude of a complex function.
P = Pressure; furnace (Pf); furnace response (Pf 0 ); furnace input, or driving
pressure (Pfj); open loop response (P 0 ); open loop input, or driving
pressure (P 1 ); constant windbox pressure wb pressure at the
exit of the boiler radiant section ( fe) intermediate pressures in a
burner (P 1 , P 2 , P 3 ); constant ambient’pressure amb
= Sensible heat leaving the boiler with the hot flue gases (time rate).
= Heat entering the boiler, in chemical form, in the fuel (time rate).
out = Electrical energy, expressed in heat units,, leaving the plant (time
rate).
R = Flow resistance; inlet to a burner (R 1 ); burner exit region (due to
presence of flame) (Rf); linearized R 1 and Rf (R j, Rffl; linearized
resistance to the flue gas flow from one burner in leaving the
radiant section (exit) of the boiler (Re); linearized resistance to the
total flue gas flow leaving the radiant section of the boiler (RbP);
steady state resistance in the burner exit region due to the presence
of a flame (R 3 ).
x
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R = Universal gas constant.
0
Re, Im = The Real and Imaginary components of a complex function.
S = The LaPlace operator.
T = Temperature; of the flue gases entering the stack (leaving the air pre-
heater) (T 5 ); ambient (Tamb).
Tj, T 2 , T 3 , T 4 , Tja, T3a = Time constants in the expression for the
dynamic response of a burner. Eqs. (B-29) through (B-33).
Ve = The control volume, at a burner exit, in which the volume flow rate
and air-fuel ratio perturbations issuing from the burner are con-
verted to furnace pressure perturbations.
X = The moles of air burning with one mole of fuel.
a = The acoustic velocity in the air within a burner.
c = The acoustic velocity in the furnace gases.
e = Designates an exponential term (natural).
g = The acceleration of gravity.
i = An index.
j r
n = The exponent of the expression for the fraction of combustion completed
within a burner. Eq. (i).
= The total number of burners in the boiler.
r = The weight air-fuel ratio; in the burner (rb).
t = Time.
= Weight flow rate; air at the burner inlet (h); air leaving the burner (‘ ‘b )
fuel leaving the burner (constant) (Wp ); total flow in and out of the
control volume in the furnace (win, 1O ) total flow leaving the
boiler radiant section ( vt); total fuel flow rate into the boiler (‘ ‘f).
w 5 = Weight of gases stored in the control volume in the furnace.
xi
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GREEK NOTATION
I Hc( HCh) = The heat of combustion (the high heat).
AH;f(aHW) = The heat of formation of the fuel (of water).
0 b 0 = The phase angle in the response of burner flow to the driving furnace
p pressure (of furnace pressure to the driving, burner flow).
T = Time delay; from the burner exit to the region of concentrated combus-
tion (combustion time delay) at the steady- state burner flow v lócity
(Tc); for acoustic wave travel from a burner exit to the exit from
-the radiant section of the furnace (Te); for acoustic wave travel from-
a burner exit to and from a solid boundary,, in the (i) direction (T 1 )•
a = An acoustic wave damping coefficient in plane wave travel in the furnace.
Eq. (B-67).
o = Designates a dynamic perturbation quantity (infinitesimally small
amplitude).
w = Frequency, in radians per second.
SUPERSCRIPTS
— = Time-invariant quantities.
‘ cii
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AC KNOW LEDGMENTS
Sincere appreciation is acknowledged for the guidance and
assistance provided by Mr. Robert E. Hall of the Combustion Research
Branch, Industrial Environmental Research Laboratory, Research Triangle
Park, North Carolina, who was the U. S. Environmental Protection Agency
Project Officer during the conduct of this study.
A special acknowledgment is also due, once again, to the Los
Angeles Department of Water and Power (LA DWP) for its earlier coopera-
tion in making available the data upon which this study is based. That data
was originally released by LA DWP for The Aerospace Corporation study of
the effects of combustion modifications on NO emissions from natural gas -
and oil-fired utility boilers, and was published in EPA-650/Z-75-O1Z,
January 1975.
Acknowledgment is also due Mrs. Sandra Barnes of The
Aerospace Corporation for her assistance in computer programming and
operation.
x l i i
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SECTION I
INTRODUCTION
Earlier studies (1), (2), and (3) by The Aerospace Corporation
were oriented toward analyzing the effects of combustion modifications on
NO emissions from natural gas-, oil-, and coal-fired utility boilers. Large
samples of data from tests run on full-scale utility boilers for the purpose of
empirically investigating these effects were used in the Aerospace analyses.
Resulting data correlations, and such extrapolations which could be safely
made, indicated no inherent limitations in reduction of NO emissions with
the use of staged combustion modification techniques (i. e., burners out of
service and/or NO ports). A question arose concerning what kinds of real,
practical limitations might exist in application of these techniques and, there-
fore, what limitations might exist in reduction of NO emissions.
In the data sample used in the earlier study (1) of ten natural
gas-and oil-fired utility boilers, potential limitations were observed involving
possible excessive loss of plant efficiency and combustion instability. Among
the data available to that study were measured levels of boiler electrical load
and fuel flow rates. This enabled a gross calculation, from measured values,
of an overall boiler or plant efficiency, defined as the dimensionless ratio of
the electrical output of the plant to the heat input represented by the chemical
energy of the fuel (in compatible units). This measure of efficiency reflects
all of those losses which can be affected by combustion modifications as well
as some which cannot (i. e., losses in the steam cycle). Accurate, measured
data on coal flow rates were not available to the study of NOx control in coal-
fired boilers.
A preliminary, direct correlation of this overall plant effi-
ciency against NO emissions indicated that, if indeed the efficiency was some
direct function of NOx emissions, efficiency losses of 7 to 10 percent-could
result from attempts to reduce NO emissions from the uncontrolled levels
to 100 parts-per-million (ppm). Such a loss, of course, would make the NO
reduction extremely expensive both in terms of cost and excessive energy
requirements. It was not expected that such a direct relation existed between
efficiency and NOx emissions but it was not known whether the changes in
hardware and/or operating conditions necessary to reduce NOx would also un-
avoidably cause significant efficiency losses.
I
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One task of this study, then, was to investigate the causes of
the observed efficiency variations and to relate them to modifications neces-
sary to reduce NO emissions. The end result of this task was to determine
if combustion modifications made for the express purpose of NO reduction
had any significant effect on plant efficiency.
A second observation from the data sample used in the study(1)
was that certain staged combustIon’ configurations exhibited high vibrations,
or combustion instability, and/or flame liftoff (flames detaching from the
burner exit and moving out into the furnace). The available data on combus-
tion instability was investigated, to a limited extent, to evaluate possible
mechanisms which might explab i the instabilities. It, was concluded that the
vibrations resulted from a feedback coupling between furnace pressure and
air flow through the burners (defined as an air-side feed system coupled -
mode of combustion instability). No effort was made in that study to investi-
gate the phenomenon beyond this preliminary identification. It did appear,
however, that these instabilities were associated with fuel-rich operation of
the burners, in the staged-combustion technique for NOx, control. Thus, this
type of problem could represent a significant limitation to maximum imple-
mentation of the stag ed-combustion technique and a real, practical limit on-
NO ’reduction. These :instabilities .were observed only in the data from
natural gas-fired boilers. Although none were observed in the data from oil-
fired boilers available to this study, such instabilities have occurred in other
oil-fir ed.boil r.s. No’ .documentation on similar instabilities in coal-fired
boilers-was ‘available.’ - - - -
I , -
-A second task of this study, then, was to analyze such air-si4
feed system coupled-modes of combustion instability in utility boilers of-the
types used to fire natural gas and oil fuels and to apply this analysis to the,
boilers in the available data sample. The end result of this task was to de--
velop and verify a method of analysis which could be used by others to re-
solve specific cases and to determine if certain combustion modifications
made for the express purpose of NO ‘reduction lead inevitably to combustion
instability. ‘ - •, , -
Beyond the study just described, the third and final year of this
grant will integrate and simplify the analyses for NOx control, plant efficiency
and combustion stability (and other data observations) conducted under apxe-
v-bus EPA grant (1) and the first two years of the current EPA grant. An. -
example design application, for a new, oil-fired boiler, will be shown.
Other studies have indicated that erosion or corrosion of the -
boiler water walls (principally in—coal-fired boilers) might also represent a,
limitation on the use of combustion modification techniques for NO control.
This possibility is being experimentally investigated by other contractors and
agencies (for example, Ref. 4).
2
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SECTION II
CONCLUSIONS AND RECOMMENDATIONS
2. 1 CONCLUSIONS
The major conclusions of this study can be summarized as
follows:
a. Combustion modifications, made for the purpose of NO
control, do not significantly affect overall plant efficiency.
b. Use of the staged-combustion technique for NO reduction,
involving very fuel-rich first-stage combustion, results in
a tendency toward unstable combustion and high boiler
vibrations.
c. A method of analysis of such air- side feed system coupled
modes of combustion instability was developed which can
be used to provide stable operating conditions even with
very fuel-rich first stage combustion.
Results of the evaluation of the effects of combustion modifica-
tions on plant efficiency showed that there are no significant losses which
can be charged directly against these modifications. The majority of the
efficiency variations observed in the data were attributed to load variations.
Load variations are not considered combustion modifications for the express
purpose of NO control. No significant efficiency losses could be charged
against operation with burners out of service or with the use of NO ports.
There was some indication that reduction in combustion air temperature,
which would reduce NO emissions, might also significantly reduce plant
efficiency, but the available data were not in a form that could be used to
support such a conclusion.
A model and analysis technique were developed for air-side
feed system coupled modes of combustion instability. The model was based
on conventional analyses of such modes in the liquid rocket industry and, in
the limit, reduced to such conventional rocket models. Comparison with
limited data on instabilities in gas-fired boilers in the data sample (1) showed
3
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reasonably good agreement. This not only tended to verify the analysis but
also to verify the conclusion (1) that the observed high boiler vibrations were
indeed air-side feed system coupled modes of combustion instability. Further
evaluation of such modes of instability showed that very fuel-rich burner
operation, for the purpose of NO reduction, results in a tendency toward
unstable combustion and high boiler vibrations. The evaluation also showed,
however, that such instabilities can be controlled by recognition of the charac-
ter of the problem and by proper design. Proper design for stability can in-
clude consideration of such design parameters as the flame anchoring tech-
nique and the burner air-side pressure drop. Combustion instability,
therefore, is not considered to represent an inherent limit on NO reduction
by combustion modification techniques but rather an undesirable side effect -
which must be accounted for by special design analyses and/or techniques.
2.2 RECOMMENDATIONS
Conclusions of this study relate to the effects of combustion
modifications on t o possible sources of limitation- on NO reduction:
excessive loss of plant efficiency and combustion instability. In general,
neither: area appears to represent inherent limits. Both evaluations, how-
ever, di4 reveal other problems which could represent such limits.
After reviewing the results of this and other studies, further
investigations are rec9mmended in the following areas:
a. Effects of combustion air temperature on plant efficiency
b. Effects of in-flame flue gas recirculation on plant efficiency
c. Effects of combustion modification techniques on the plant
efficiency of coal-fir:ed boilers
d. Burner designs which provide soundly stable flame nchor.ing-
in burners operated with very fuel-rich mixtures.
In the efficiency study there was some indication that reduc--
tion in combustion air temperature, sometimes considered as a means of
reducing NOx emissions, could significantly reduce plant efficiency. The
data available to this study were not adequate to evaluate this possibility and
further study is recommended. Also, combustion modification techniques
other than those involving burners out of service and NO ports, such as
in-flame fiue gas recirculation, were not tested and their effects on plant
efficiency were not studied.
In the combustion stability task, it was shown that some full-
scale utility koilers are currently being operated with sufficient burners out
of service (to control NOR) that the remaining active burners are operating
4
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very near fuel-rich flammable limits. Not only is it important that each
case be evaluated, and perhaps re-designed, to assure combustion stability,
but additional study should be devoted to designing burners specifically to
provide soundly stable flame anchoring when the overall burner air-fuel
ratio is near or below the flammable limit. Without such design it is con-
ceivable that the problem of providing a stable pilot flame, and reliable
subsequent flame propagation, in a very fuel-rich burner could represent
a real, practical limit to NO reduction by techniques involving burners out
of service or NO ports.
5
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SECTION III
SUMMARY
This study consisted of two tasks, both largely addressing
potential limitations in utility boilers to NO reduction by combustion modi-
fication techniques. These limitations were related to possible excessive
loss of plant efficiency and to combustion instability. Both tasks involved
large samples of data obtained from natural gas- and oil-fired boilers which;
were reported previously by Dykema (1).
3. 1 PLANT EFFICIENCY
The basic data used as a starting point in the study of efficiency
were measured values of electrical load (plant output) and fuel flow rates
(heat input), in compatible units. Efficiency losses and efficiency corrections
were calculated and evaluated in six areas: a. electrical generating and con-
trol equipment; b. uncondensed water vapor in the flue gases; c. excess
sensible heat entering the stack in the flue gases; d. incomplete combustion;
e. variations in expansion efficiency through the steam turbine with varying
boiler load; and f. an ttothertl category remaining after the overall plant effi-
ciency data were corrected for the first five, identifiable losses.
Losses due to the electrical equipment and to uncondensed
water vapor are clearly not functions of any combustion modifications for NO
control. As a result, only a simple correction was derived. The efficiency X
data were corrected to remove these effects so that the remaining data would
be more sensitive to the effects of combustion modifications, if any.
Combustion modifications could affect the excess sensible heat
loss. A calculation for these losses was derived and evaluated over the range
of operating variables represented in the data. It was concluded that efficiency
losses from this source greater than about one-half percent would be reflected
in increased combustion air temperatures. Evaluation of measured com-
bustion air temperatures showed rio significant differences between the nom-
inal and modified operating conditions. As a result, it was concluded that
combustion modifications do not significantly affect sensible heat losses.
Again, these losses were calculated and the remaining efficiency data cor-
rected to remove variations due to sensible heat.
6
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Losses due to incomplete combustion were thought to be the
most likely to increase with combustion modifications. Strongly fuel-rich
combustion in the first stage could result in a number of energetic hydrocar-
bons, carbon monoxide (CO), and increased carbon loss with the particulate
emissions. The measure of this loss used here was the relation of the CO 2
concentrations measured in the flue gases to the theoretical CO 2 concentra-
tions calculated from stoichiometry, at the overall boiler air-fuel ratio cor-
responding to the measured oxygen (02) level. Zero loss due to incomplete
combustion would be indicated by CO 2 concentrations which contain all of the
carbon in the fuel.
No evidence was found in either the direct measured levels of
CO 2 or in a calculated efficiency correction (depending on the measured CO 2
level) to indicate that combustion modifications for the purpose of NOx con-
trol affect losses due to incomplete combustion. Of course, this loss is one
which is particularly susceptible to correction by the boiler operator. If
high CO levels or smoke are observed, the boiler operator could change the
level of excess air, or some other variable, to eliminate these malfunctions.
Perhaps the correct conclusion here, then, might be that no evidence was
found that combustion modifications for the purpose of NOx formation affect
losses due to incomplete combustion to the extent that they are not readily
correctable by the boiler operator. Sensible heat losses, which vary with
boiler excess air (if that is the parameter used by the boiler operator to
eliminate high CO or smoke) would vary only slightly.
Variations in boiler load are necessary to satisfy the varying
electrical demand. It is well known that NO concentrations in the flue gases
usually decrease significantly with decreasing load. Two major variables
which change with boiler load are the combustion air temperature and the ex-
pansion efficiency through the steam turbine. The former affects NOx emis-
sions and the latter affects plant efficiency. Thus it was expected that both
NO and plant efficiency would decrease with boiler load. Although load vari-
ation is not considered a practical NOx control technique, it was necessary
both to derive an efficiency correction for load-related losses and to evaluate
what fraction of the observed efficiency variations results from this parameter.
An empirical correlation of efficiency data showed the parabolic
curve typical of turbine efficiency variations, with maximum efficiency at the
most common operating condition, about 80 percent of rated load. Load re-
duction from 80 to 40 percent of rated resulted in an efficiency decrease of
more than six percent. Thus, excluding data scatter, almost all of the plant
efficiency variations observed in the available data sample are due to load
variations, leaving little other variation which might be charged against com-
bustion modifications made for the express purpose of NO control.
It is possible that the observed parabolic form of the efficiency
data with load does not represent steam turbine efficiency variations alone.
7
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It is then further possible that part of the loss at low loads could result from
the reduced combustion air temperatures. If combustion air temperatures
alone were reduced at fixed load, then, to reduce NO emissions, significant
efficiency losses might result. This would be a case where the combustion
modification desired, for the purpose of NO control, would significantly re-
duce plant efficiency. Unfortunately, the data available to this study were not
adequate to delineate the separate effects of combustion air temperatures on
both efficiency and NO emissions. The appropriate conclusion of this study
is still that the observed (significant) efficiency variations with load are not
the result of a combustion modification made (at least in the data of this study)
for the express purpose of NO control.
Finally, after correcting the original overall plant efficiency
data for the five losses discussed above, the remaining data variations’ were
examined for any other losses which might result from combustion modifica-
tions. Although there was some indication of a direct relation of efficiency
to the combustion temperatur-e in the fuel-rich first stage, the effect could
not be larger than about 0. 8 percent. This variation is small compared to the
data scatter. The final efficiency data, corrected for all of the losses dis-
cussed-herein, had a standard deviation of 2.2 percent.
3.2 COMBUSTION STABILITY
Models and analysis techniques of feed system coupled modes
of instability developed in the rocket industry were used as a basis for the
development of a method of analysis for such feedback systems coupled to
the air flow system in. utility boilers. A major modification of those models’
and techniques was necessary to adequately describe the coupling between’
burner flow rate perturbations and resonances in the three coordinates of -
the furnace cavity. This modification greatly complicated the analysis but
it was shown that, in the limit (primarily, in the case where the furnace
cavity dimensions are very small), the analysis developed here becomes
identical with that long used in the rocket industry.
Basically, the model can be described as follows. The boiler
windbox is taken as a large, constant-pressure plenum from which combus-
tion air enters the burners. The air’ in the burners has compressibility and
inertia. Resistance to air flow through the burner is in two parts, a con-
stant resistance at the burner inlet (through the air registers) and a variable
resistance near the burner exit which is a function of the degree of initial
combustion within the burner.
The pressure drop a-cross the burner from the windbox to the
furnace is very small, measured in inches of water. As a result, this pres-
sure drop and the resulting air flow rates are quite sensitive to variations in
furnace pressures at the burner exit. Perturbations in furnace pressure at
the burner exit cause perturbations in the air flow -rate through the burner.
8
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The pressure drop across the fuel injectors (orifices) however is quite large,
usually being measured in pounds per square inch (psi). Perturbations in
furnace pressure have negligible effects on fuel flow rates. As a result of
constant fuel flow rates mixing with varying air flow rates, the air-fuel ratio
leaving the burner is also varying.
The effect, in turn, of varying air flow rates and air-fuel ratios
entering the furnace is in two parts. The total volume flow rate perturbations
begin immediately upon leaving the burner to mix with the gases in the fur-
nace, decelerating and generating acoustic waves which propagate away in
all six directions. Combustion takes place at some later time (the combus-
tion time delay), farther out in the furnace. Heat release rates vary only as
a function of the varying air-fuel ratio. When the burner is operating very
fuel-rich (a combustion modification for the purpose of NO control) the heat
release rate varies strongly with air-fuel ratio. The varying heat release
rates also create acoustic waves which propagate away in all six directions.
After appropriate time delays for acoustic wave travel to the
limits of the furnace cavity, and reflection off solid boundaries (at some
efficiency) the waves return (at different times) to the burner exit and add
together to create the furnace pressure variations which, in turn, cause
further variations in air flow rates and air-fuel ratios coming from the burner.
Acoustic waves which travel to the exit of the radiant section of the boiler
cause varying rates of flow of gases out of the radiant section into the back
pa s S.
The resulting analytical model resolved, in the limit, to the
conventional rocket engine model (a verification of the basic model). It was
also verified reasonably well by comparison with the available data. The
analytically predicted instability frequencies of 11 and 43-45 Hz agreed well
with measured frequencies of about 12.5 and 40-50 Hz. The model properly
predicted a strong instability at 11 Hz. Although the frequencies in the 40-
50 Hz range were properly predicted, the model did not clearly predict an
instability in this frequency range. The model also properly showed that
the vertical distribution of burners in the burner array prevented a possible
instability in the intermediate frequency range, at about 30 Hz. En general,
the agreement with data is considered good for this kind of analysis.
Further study of the characteristics of air-side feed system -
coupled modes of instability such as this in utility boilers showed two signifi-
cant results relative to combustion modifications made for the purpose of
NO control. If the burners are operated very fuel-rich, near to the rich
flammable limit, the variation of the heat release rate resulting from air-
fuel ratio perturbations is very large (the gain in the loop is high) and the
boiler tends to become unstable. Boiler vibrations were consistently shown
to become significant when more than about 25 percent of the burners were
9
-------
operated air-only, for purposes of NO control. It was shown that the burner
air-fuel ratio, under such operation, approaches the natural gas fuel-rich
fla.rnmable limit. Unless care is taken to reduce the loop gain, combustion
instability could become a very real, practical limit to NO reduction.
The second observation relates to the anchoring of the flame
in the burner. If the fraction of combustion which is completed within the
burner is a. function of the flow velocity in the burner, the flow rate through
the burner becomes more sensitive to furnace pressure perturbations and
combustion vibrations are amplified. All of the cases of high vibration oh- -
served in this data sample from natural gas-fired boilers (and one, undocu-
mented case with oil fuels) involved relatively poorly anchored flames in the
burner. Aerodynamic (vortex) flame stabilization appears particularly sensi-
tive to variati ons in flow velocity through the hurner. Again, fuel-rich burner
operation creates a more difficult flame a.nchoring problem and, in turn, a
greater possibility of unstable combustion. Unless recognized and properly
corrected, tEis could also represent a limit on NO reduction.
In general, however, thé analysis of air-side feed system
coupledmodes of instability in utility boilers shows that, while modifications
to the combustion in existing boilers can create unstable combustion and thus
limit NO reducti9n, such instabilities can be prevented by analysis and
proper design. The worst case, in designing for stable combustion with
burners operating near the rich flammable limit, must be resolved for other.
reasons. Simply to maintain a flame in the furnace the burners would have
to be designed with somewhat less fuel-rich flame ignition and stabilization
zones, with subsequent controlled mixing and flame,propagation,into the more
fuel-rich regions. These design changes are simultaneously those which
minimize combustion instability. Therefore, there i s no reason to conclude
that combustion instability represents an inherent limit to NO reduction.
10
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SECTION IV
PLANT EFFICIENCY
During the study described by Dykema (1), data were
accumulated on electrical load (the useful plant output) and fuel flow rates
(heat input) for most of the test conditions involving natural gas and oil fuels.
These are the two parameters of real, practical interest from the standpoint
of energy efficiency; the electrical power which can be generated for a given
heat input from the fossil fuel. In the overall study, of which this task is a
part, the interest has been, and is, in the control of NO emissions. Energy
efficiency, then, represents a constraint on NO reduction techniques in that
implementation of a NO reduction technique, by definition, shall not signifi-
cantly reduce the energy efficiency of a plant.
The initial impetus for the efficiency study derived from the
observation that the overall plant energy efficiency varied over a range of
as much as 12 percent. A simple, direct computer correlation was attempted
at that time, relating efficiency directly to the measured NO data, using a
second-order polynomial. The resulting correlation coefficients were moder-
ately low (0.34 and 0.58 with oil and natural gas fuels, respectively). Those
results tended to indicate that techniques used to reduce NO had no signifi-
cant effect on plant efficiency. The correlation coefficients were sufficiently
high, however, to suggest further investigation. If there were a direct rela-
tion, the derived correlation equations indicated that if the NO levels were
reduced from the uncontrolled levels (440 ppm for oil and 660 ppm for gas)
to 100 ppm, reductions in efficiency of 7.0 and 9.6 percent, respectively,
could result. Such efficiency losses would indeed be considered significant.
If such losses were shown to result from the modifications made for the pur-
pose of NO reductions, then that technique for NO control could not be con-
sidered acceptable.
The objective of this task, then, was to identify whether (and
how much) combustion modifications made for the purpose of NO control
might affect the energy efficiency of a utility boiler. If the dimensionless
ratio of electrical output (load) to heat input from the fuel is considered an
acceptable measure of the overall energy efficiency of a utility boiler, then
a large sample of appropriate data was available (1). In using that data for
this study it was assumed that if there were no need to control NO emissions
ii
-------
the boilers would be fired such that: (a) all burners were active, (b) no NO
ports were used, (c) overall boiler excess air would be only slightly higher
than that necessary to assure acceptably low carbon monoxide (CO) emissions,
and (d) boiler load would be varied as necessary to meet electrical output
requirements. Tests involving burners out of service (air-only) and air flow
through NO ports, then, represent tests of combustion modifications (from
the normal mode) made for the purpose of NO reduction.
4. i COMBUSTION MODIFICATIONS FOR CONTROL
It seems clear that the formation of the rather trace quantities
(parts per million) of NO in a boiler cannot be the direct cause of any mea-
surable changes in overall plant efficiency. Instead, NO emissions vary in
response to variations in some other parameter. The variations in that
parameter, then, may also cause the plant energy efficiency to change.. The-re
are a number of operational and ge6metric parameters in boilers variations
of which can have significant effects’ on NO emissions. Some of these param-
eters must be varied in norn al operation and in design simply to satisfy the
main purpose of a utility boiler; the demand for electrical power. Some of
these parameters, however, can be deliberately varied for the sole purpose
of reducing NO emissions. It is the simultaneous effect of va’riations in
these latter parameters on plant efficiency that is of interest here. In order
to use all of the data available it was necessary to correct for, or weed out,
the effects on plant efficiency of variations not for the purpose of NO, control
so that the effects of those that are made for the purpose of NO control can
be seen and evaluated.
For example, we know that strong reduction in boiler load will,
in almost all cases, strongly reduce NO emissions, with a simultaneous
significant effect on plant efficiency. , Although some load sharing with other
boilers can take advantageof this observation to minimize overall NOx, nor-
mally the load ona given boiler must be varied to accommodate the smaller
variations in electricaldemand regardless of the effect on NO emissions.
Load variation, then, is not consideredäNO control technique in this study.
The installation of NO ports, however, is a geometric and operational vari-
able made for the primary purpose of NO control. It is important in this
study to try to evaluate, from experimental data, the simultaneous effect of
NOx ports on plant efficiency. -
Similarly, the study (1) showed that, at least with natural gas
and oil fuels, the effect of overall boiler excess air on NO emissions was
small. In any case, all indications are that the plant efficiency is maximum
with minimum excess air, compatible with acceptable emissions of CO and
smoke. As a result, control of excess lair, at least for this study (involving
natural gas Wand oil fuels), is not considered a significant NO control
technique. .
12
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Reduction in combustion air temperature is known to be an
effective NO control technique if the NO emissions are largely thermally
generated. dthough combustion air temperatures varied somewhat in the
data (i), they varied only as a result of, and in conjunction with, boiler load
(or total fuel and air flow). There are no data in Ref. 1 directly showing the
effect on NO emissions of combustion air temperature alone. Some efforts
were made (1) to analytically separate this effect. Because this modification
can have a significant effect on NO emissions, and might be deliberately
varied to control NO , a similar effort is made here to evaluate the magnitude
of the effects of com ustion-air temperature reduction on plant efficiency.
Although load variation is not considered here as a NO control
technique, it does have a significant effect on efficiency. In order to use all
available data to determine the effects of NO control techniques, it was neces-
sary to account for, or correct for, efficiency variations resulting from load
variations, as well as from all other hardware and operating condition vari-
ations which are not specifically for the purpose of NO control.
4.2 EFFICIENCY LOSSES
Six areas of efficiency losses were considered in this study.
Three considered heat losses from the combustion gases (heat not transferred
to the steam) and two considered efficiency losses in the steam and electrical
processes. The final area was an Itothertt category. Of these six areas,
four are reasonably well-known and can be calculated: electrical generating
and control equipment; uncondensed water vapor in the flue gases; excess
sensible heat entering the stack in the flue gases; and incomplete combustion.
Data were not available with which to calculate variations in the energy ex-
pansion efficiency through the steam turbine. A reasonable functional form
of these losses could be assumed however and quantified from the data. After
subtracting these five losses from the overall plant efficiency, the remaining
category could still show some effects of combustion modifications which were
not accounted for in the calculated losses.
Obviously, efficiency losses in the electric generating and con-
trol equipment, at a given electrical load, are independent of how the heat is
being supplied to the steam in the boiler. Thus, these losses could not be a
function of combustion modifications made for the purpose of NO control.
This loss is calculated, and the efficiency corrected, only to increase the
sensitivity of the corrected efficiency data to other variations.
It appears desirable to maintain the temperature of the flue
gases entering the stack well above the dew point of the water vapor to avoid
the corrosion which can result from condensation of water containing dilute
acids on cooler surfaces (such as the air pre-heater). As a result, the heat
losses due to uncondensed water vapor in flue gases are unavoidable and will
13
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exist regardless of any combustion modifications. Actually this loss appears
in the data of this study only because the initial efficiency calculation, follow-
ing convention, involved the high heat of combustion rather than the low heat.
Again, this loss is calculated only to increase the sensitivity of the corrected
efficiency to other variations.
One way in which combustion modifications could reduce plant
efficiency would be by reducing the heat transferred to the steam cycle and
increasing the sensible heat losses in the hot flue gases entering the stack.
These increased heat losses could occur both through increased heat capacity
(specific heat) and the temperature of the flue gases. Since the heat capacity
of the flue gas is largely determined by the fraction of inert nitrogen in the air,
it is almost totally a function of the overall boiler air-fuel ratio, or the ex-
cess air. Small changes in the composition of the combustion products have
a negligible effect on the heat capacity. While it must be accounted for in
calculating sensible heat losses, it was calculated in this study simply as a
function of the overall boiler air-fuel ratio and, as a result, was not a function
of any combustion modifications.
On the other hand, increased losses due to combustion modifi-
cations could be reflected in higher flue gas temperatures at the same boiler
air-fuel ratio. Unfortunately, the temperature of the flue gases leaving the
pre-heater (and entering the stack) was not recorded in many of the tests
available to this study. Review of such temperature data as were available
indicated that they could fit.reasonably well as a.linear function of the plant
electrical load. Combining this temperature function with the heat capacity -
function discussed above, the efficiency losses for all of the data were calcu-
lated and examined. This showed that, for a given boiler, efficiency losses
(not heat losses) from this source ranged from about 1.9 percent at minimum
load to about 3. 1 percent at rated load. This range largely resulted from flue
gas temperature variations of about 39 K (70°F). These flue gas temperature
variations are clearly reflected in the combustion air temperatures, physically
through the pre-heater. Flue gas temperature variations of about 16 K (30°F)
or higher, corresponding to one-half percent or more efficiency losses, are
easily detectable through combustion air temperature variations. The latter
measurements were available to this study for most data. Further review
of combustion air temperature data showed no consistent variations greater
than 16 K (30°F) as a result of combustion modifications (at.any given load and
level of excess air). Therefore, it was assumed that any effects of combus-
tion modifications on sensible heat losses must amount to less than about one-
half percent efficiency variation. The efficiencies calculated from measured
data were only accurate to within 1 to 2 percent. As a result, further effort
to detect any efficiency variations due to sensible heat losses was terminated.
As with the two previous efficiency losses, however, the overall efficiency
was corrected for these sensible heat losses, again to make the remaining
efficiency variations more sensitive to other effects.
14
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The area of efficiency loss which appears the most likely to
be significantly affected by combustion modifications is that identified here
as incomplete combustion. The most effective general technique for NO
control demonstrated in the data is the so-called two-stage combustion. In
this approach, the first stage is operated fuel-rich, and appreciable cooling
of the products of this fuel-rich combustion is accomplished before the re-
maining air is introduced to provide fuel-lean combustion in the second stage.
One can easily visualize that if the first stage were operated too fuel-rich the
flame might actually be extinguished (and the efficiency drop to zero). Also,
the products of fuel-rich combustion still contain considerable chemical
energy (for example, in high concentrations of carbon monoxide). If these
products are cooled too much before the remaining excess air is added, fur-
ther reaction to minimum energy products (for example CO to CO 2 ) could be
inhibited, with a resulting large loss in efficiency. Thus, efficiency losses
due to incomplete combustion were of most interest in evaluating the effects
of combustion modifications.
The final major loss which clearly affects the efficiency data
variations is that due to steam expansion losses through the turbine. These
losses could be calculated directly if the flow rates and thermodynamic prop-
erties of the steam at the turbine control valves and at all exits from the
turbine system were known. Unfortunately, these data were not gathered in
the study reported in Ref. 1. It is known, however, that the need to vary the
power output of the constant speed turbine to match the desired electric gen-
erating load results in operation at less than maximum turbine efficiency
over most of the load range. Losses resulting from this mismatch are
clearly independent of any combustion modifications.
It is possible that changes in the heat release and heat transfer
profiles through the boiler, besides changing the exit flue gas temperature
(total heat transferred), could also change the overall thermodynamic cycle
efficiency. For example, staged combustion could reduce the combustion
temperature and the heat transferred in the radiant section of the boiler while
raising both in the area of the superheater tubes. In such a case it might be
possible that the resulting average temperature difference (gas-to-water or
steam) might be ‘lower and the resulting steam cycle efficiency higher.
The large variations in efficiency losses with load through the
steam turbine were at least partially accounted for in this study by assuming
that these losses would follow some parabolic function of load, with the maxi-
mum efficiency probably occurring (by design) at the most common load,
about 80 percent of rated load. A single expression for this loss, as a func-
tion of load alone, was empirically determined for all data on a given boiler
(turbine design), regardless of the fuel burned or the combustion modification
in use. These losses were then subtracted from the overall efficiency data.
Any effects of combustion modifications on thermodynamic cycle efficiency,
then, would be more clearly observed in the remaining efficiency data.
15
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In this discussion, it has been implied that the effect of a given
combustion modification on efficiency could be seen directly in the difference
in measured efficiency with and without the modification. It must be remem-
bered, however, that the utility boilers involved in the tests were also gen-
erating electricity for commercial use during the tests. It is normal practice
for the boiler operator, not always intimately.involved in the’environmental
testing in progress, to make certain operational adjustments to maintain
boiler operation as close to optimum as possible. Thus if a given modification
is made to the combustion to evaluate its potential to control NOR,. and the
change results in a severe decrease in plant efficiency, the boiler operator
(or even some automatic control equipment) may adjust operating conditions
to recover all or part of this loss. For example, if a given configuration of
active burners and’burñers out of service, resulted in higher temperature flue
gases entering the stack; the boiler operator might increase the rotational
speed of the air preheater to transfer more heat to the incoming air and re-
cover this loss. Wherever possible, then, the modified and nonmodified
(normal) operating conditions need also to be examined to determine if operat-
ing conditions are significantly different. Since there were not sufficiently
detailed data in Ref. 1 to evaluate all possible operational changes, a con-
clusion that a given combustion modification did not significantly affect -
efficiency must be considered toimply also that the efficiency was at least
not affected beyond the ability of the operator to restore operation to normal
by changing. some other patt of the system.
4.3 RESULTS
Evaluation of the effect on plant efficiency of combustion modi-
fications made for the purpose of NO control was conducted in three steps:
a. The overall plant efficiency data were corrected to remove
the influences, of the four efficiency losses which could be
calculated directly. The calculated losses due to incom-
plete combustion were then evaluated to determine if com-
bustion modifications had any significant effect. The -other
three areas of efficiency losses were not evaluated further.
b. An empirical relation was established, from the data, to
account for efficiency variations due to variations in ex-
pansion efficiency through the steam turbine. Although.
this loss cannot be charged. to combustion modifications
made for the purpose of NOx control, the simultaneous
effects of load on efficiency and NO were evaluated.
c. Finally, the plant efficiency data, already corrected for
the losses in a., were further corrected for the loss vari-
ations from b. and the resulting data evaluated for any
other effects of combustion modification.
16
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Results described herein do not concern losses in the electrical
generating and control equipment or those due to uncondensed water vapor or
excess sensible heat in the flue gases. These latter corrections are described
briefly below.
4. 3. 1 Efficiency Corrections
As described in Appendix A, a simple three percent load loss
was assumed in the electrical generating and control equipment for all plants
and for all loads. This percentage was somewhat arbitrarily chosen as a
typical loss for such systems. The main observation, however, is that this
loss is relatively small and variations from this nominal level should be
negligible.
Heat losses due to uncondensed water vapor in the flue gases
are a constant fraction of the heating value of the fuel for the natural gas and
low sulfur oil fuels used in the test data (1). These heat losses represent
9.63 and 5.60 percent of the higher heating value of these fuels, respectively.
The overall plant efficiency, following convention, used the higher heating
value. In correcting the plant efficiency data for the uncondensed water vapor,
a loss certainly not chargeable to combustion modifications, we are, in effect,
using the low heat of combustion. As in all of these corrections the efficiency
correction does not amount to the full 9.63 or 5.60 percent heat loss, because
more than half of this heat would be lost due to the low steam cycle efficiency.
For a typical gas-fired boiler the efficiency correction for uncondensed water
vapor amounted to increasing the remaining efficiencies by 4. 1 to 5.2 percent.
The sensible heat losses are not a constant fraction of the heat-
ing value of the fuel but also depend on the air-fuel ratio at which the fuel is
being burned. These losses are greater at high air-fuel ratios, or levels of
excess air, because more total flue gas leaves the boiler at any given tem-
perature. This is somewhat offset by a decrease in flue gas heat capacity
with increasing air-fuel ratio but the overall effect is still for these losses
to increase with air-fuel ratio, or excess air.
For all of the data of Ref. 1 from one facility (six boilers)
firing natural gas and oil fuels, the efficiency corrections for sensible heat
losses ranged from 1.6 to 3.1 percent. The data from the other facility
(two boilers) showed corrections in the range of 2. 5 to 3. 8 percent. Since
these calculated corrections involved functions only of boiler load and overall
boiler excess air they could not be expected to show any effects of combus-
tion modifications, such as burners out of service,’ so no significant attempt
was made to further analyze the calculated corrections. One clear reault is
the conclusion that reduction in excess air increases efficiency. The NO
analysis of Ref. 1 has already indicated that, at least with those utility
boiler data, reduction in excess air slightly increases NO with natural
gas and slightly decreases NO with oil fuels. In neither case are the
17
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efficiency or NO variations large. Coal-fired boilers, however, should show
the same trend of efficiency with excess air but decreasing excess air can
significantly reduce NOx. In general, then, decreasing excess air increases
the plant efficiency and generally decreases NO emissions. As discussed
earlier, however, reductions in excess air are not really considered a com-
bustion modification made primarily for the purpose of NO control.
4. 3. 2 Incomplete Combustion
An area of efficiency losses which can be reasonably calcu-
lated and can be significantly affected by combustion modifications involves
incomplete combustion. As shown in Appendix A, these losses can be esti-
mated from flue gas measurements of the carbon monoxide, carbon dioxide
and oxygen. These data can then be evaluated to determine, directly, if com-
bustion modifications significantly affect this loss.
As discussed in Appendix A, the measure of the completeness
of combustion used in this study is essentially a carbon loss calculation. in
the few tests where unburned hydrocarbons were measured, they appeared to
be negligible. In all cases CO was negligible, at least as a source of signifi-
cant energy loss. The assumption was made, therefore, that all carbon from
the fuel which does not appear as CO 2 in the flue gas analyses is lost (does
not yield any heat). Complete combustion (zero loss) is defined as the case
where the measured CO 2 is equal to that, calculated from stoichiometry cor-
responding to the mea&ured °2
Figure 16, in Appendix A, already shows that the combustion
modifications, such as result from the use of burners out of service or NO
port techniques, have no discernible effect on losses due to incomplete com-
bustion. The measured levels of CO 2 consistently appear slightly less than
one-half percent below those calculated from stoichiometry, at any measured
level of 02. The significant observation is that measured CO 2 data from
boilers fired with all burners active and NO port closed (nominal design con-
dition) are the same, within the scatter of the data, as those where the com-
bustion was modified (for NO control) with NO ports open, one-third of the
burners operated air-only or both.
The °2 and CO 2 data were used to calculate a correction to
the overall plant efficiency, to remove the influence of this loss from the
overall data so that the remaining data could be analyzed further. These cal-
culated corrections were also reviewed to again evaluate whether there are
any significant differences between losses due to incomplete combustion when
the boilers were operated in the nominal condition and when these same
boilers were operated with modified combustion. The data examined here
were the differences between the average calculated efficiency corrections for
this loss under these different operating conditions.
18
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The resulting data are shown in Table 1. Of the sixteen possible
combinations of boiler types, burner configurations, NO port operation and
fuels, the effects of combustion modifications on the efficiency correction for
incomplete combustion were well within the accuracy of the data (data scatter)
in thirteen of the combinations, ranging from an increased loss of 0.6 percent
to a decrease in this loss of 0.8 percent. No data were available from two
combinations and one showed an increased loss of 2.4 percent. This latter
case, from the smallest, single-wall boiler, apparently resulted from the fact
that only four tests were available of the nominal configuration and these in-
volved CO 2 measurements made by a different company, using different
apparatus, than those made under the modified configurations (or on any other
boilers or with the gas fuel). As a result the 2.4 percent loss is not con-
sidered valid. It certainly is not consistent with all of the other thirteen
observations.
This evaluation of the effects of combustion modifications on
efficiency corrections agrees with the previous conclusion directly from the
C0 2 /0 2 data. This conclusion, and the conclusion of this study, is that com-
bustion modifications for the purpose of NO control do not significantly affect
efficiency losses due to incomplete combustion.
4. 3. 3 Effects of Load Variations
As discussed several times to this point, the variation of load
is considered a necessary operational function of an electric generating plant
and, as such, is not considered a combustion modification made for the pur-
pose of NO control. It is well-known that a reduction in boiler load normally
results in a reduction in NO concentrations measured in the flue gas (although
not necessarily in NO emissions expressed as grams per unit heat input).
Although the necessary data were not in the sample available to the previous
study (1), it was suggested here that the reduction in measured NO concen-
trations with load might be due to the simultaneous observed reduction in
combustion air temperatures with load. Reduced combustion air tempera-
tures reduce initial combustion product temperatures which, in turn, reduce
the rate of NO formation by the thermal mechanism. Control of combustion
air temperatures is a valid combustion modification which could be made for
the purpose of NO control.
Unfortunately, in the data sample used (1), reductions in boiler
load were always accompanied by reductions in combustion air temperatures.
Since many other changes take place when the load is changed, there was no
suitable, valid way in that study (1) to evaluate the temperature effect alone.
Similarly, there is no good, valid way in this study to investigate the effect
of combustion air temperature alone on plant efficiency. It is again necessary,
therefore, to look only at the combined effects of all parameters which vary
with load and it is still necessary to define that load variation is not a
19
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TABLE 1. EFFECTS OF COMBUSTION MODIFICATIONS ON PLANT EFFICIENCY
LOSSES DUE TO INCOMPLETE COMBUSTION
Difference in Efficiency Corrections, %4
Boiler
No.
Firing
Type
NO
Parts
Rated
Load, MW
No.
Burners
I
Natural Gas
Oil____
,
A*
B
C
D
A
B
C
D
1
2
Single-.
wall
No
180
240
240
- 360’
16
12
12
24
-
•
+0.6
-0.5
-
-0.3
+0.1
-
-0.6
-
-
-
-
+0.1
-0.1
O.1
-
-
+0.4
+0.2.
-
+0.6
N.A.
-
+2.4
-
-
- -
-
-0.3
N.A.
-0.8
Opposed
Yes
Averages +0. 1
-0.1
-0.6
0
Averages Both Gas and Oil
+0.3
+0.6
+2.4
-0.6
+0.
2
+0.
2
+0.9
-0.
3
0
Firing
Config.
Fraction Burners Air
NO Ports
.
Nominal
. .
Modified
.
Nominal
.
Modified
A
B
C
D
None
None
None
None
None
1/6
1/4
1/3
Closed
Closed
Closed
Closed
Open
Open
and
Closed
*Difference in Efficiency Correction = The correction for losses due to incomplete com-
bustion in the modified operating configuration minus the correction in the nominal
configuration.
Firing Configurations **
-------
combustion modification made for the purpose of NO reduction. The attempt
here is only to develop a method of correcting for the major effects of effi-
ciency which are not affected by combustion modifications.
For this purpose it was assumed that the expansion efficiency
through the steam turbine should be the major effect of load on plant efficiency.
It was further assumed, subject to experimental verification, that this effi-
ciency variation should be described by some parabolic curve (second-order
polynomial) of efficiency as a function of load. It was expected that the maid-
mum would occur near the most common operating point (about 80 percent of
rated load). The appropriate curve should be the same for a given boiler
type but independent of fuel type (natural gas or oil) and certainly independent
of combustion modifications resulting from the use of burners out of service
or NO ports. The actual magnitude of the efficiency, of course, could vary
as a function of many other variables, but any remaining magnitude variations
could be evaluated in the final corrected efficiencies data.
The data used to develop this load correlation were the original
plant efficiencies corrected for all four of the other calculated losses described
in Sections 4. 3. 1 and 4.3. 2. A number of data samples were correlated to
evaluate differences in boilers, fuels and other operating conditions. The final
correlation developed for the largest boiler in the sample is shown in Figure 1.
This same correlation curve, reduced by a constant three percent, also cor-
related all of the data from the rest of the boilers in the sample. The correla-
tion coefficient for all of the data, all boilers, both fuels and all operating con-
ditions, was 0.744. This is considered au adequate correlation.
The efficiency curve shown in Figure 1 shows a maximum of
50. 5 percent, at about 84 percent of rated load, and drops to as low as
44. 3 percent at 40 percent of rated load, a decrease of more than six percent.
Since load variations down to about 43 percent of rated load were part of the
data, it is likely that most or all of the efficiency variations observed in the
data, excluding data scatter, result from load variations. It follows, then,
that little other efficiency variations remain which could be ascribed to com-
bustion modifications made for the purpose of NO control.
The NO data from the same plant whose efficiency variations
with load are shown in Figure 1, were correlated in the previous study (1).
These correlations, for both fuels, are also shown in Figure 1. It can be
seen that, while NO concentrations from both fuels decrease with load (as
does the efficiency), neither follows the parabolic shape of the efficiency
curve. If anything can be interpreted from this comparison, it would appear
that the NO concentrations probably follow the combustion air temperature
variations and the efficiency follows the steam turbine performance, both
of which are, in turn, associated with load variations. Without further data
however these separate, and perhaps little related effects cannot be verified.
21
-------
700
600
EFFIC
500
50
- -
NO -NATURAL GAS
0
t’J
L’J
/
L)
W
C-)
U-
U-
46 ’
C_)
>-
-C)
Au <
I-
f -r i
lI.
— —
_ - —
— —
NOR-OIL
0.5
0.7 0.8
FRACTION OF RATED LOAD
42
0.9
1.0
Figure -1.
Effects ofboilerload variation on efficiency and NO emissions.
x-
-------
In any case, it is clear from Figure 1 that, at least for these
boilers, load reduction does indeed reduce NO concentrations in the flue
gases but it also substantially reduces plant efficiency. The efficiency re-
ductions shown correspond to the majority of the efficiency variations ob-
served in the data. There is not sufficient data in the available sample to
determine whether the two effects are separate and, therefore, whether this
same NO reduction could be attained without effecting the efficiency.
The correlation of corrected efficiency versus load discussed
here was finally used to further correct the efficiency for a final evaluation.
That final evaluation is discussed in Section 4.3.4.
4. 3.4 Other Possible Effects
After correcting the original plant efficiency data for the four
losses discussed in Sections 4. 3. 1 and 4. 3. 2 and for the load effects discussed
in Section 4. 3. 3, very little variation remained outside of the expected data
scatter. Estimated precision of the original (instantaneous and simultaneous)
measurements of fuel flow and electrical load, and of the various other mea-
surements which were used in efficiency corrections, resulted in the conclu-
sion that the precision of the final, corrected efficiency values could not be
better than about two percent. This was about the range of what appeared to
be the data scatter in the final efficiency data. Nevertheless, some effort was
made to evaluate any other pos sible effects of operation with burners out of
service, with NO ports, or both.
Several attempts were made to evaluate possible mechanisms
by which these combustion modifications might affect efficiency. The one,
if any, which appeared to have a discernible effect was based on the assump-
tion that cooler combustion in the radiant section of the boiler, resulting
from fuel-rich first stage operation (in turn resulting from burner and NO
port variations) might cause an efficiency loss. An expression was developed
which related the average combustion temperature in the active burner region
of the boiler when the active burners were operated fuel-rich to that same
temperature under nominal boiler operating conditions. Reasonable data cor-
relation coefficients were obtained using this term but the maximum possible
efficiency loss with staged combustion was less than one percent. Consider-
ing the estimated range of data scatter, and this very small potential effi-
ciency loss, it was concluded that no further significant effects of combustion
modifications, made for the purpose of NO control, could be observed in
the data.
After correcting the calculated efficiency data for all of the
losses discussed in this report (including the one of this section) the re-
maining data had a standard deviation of 2. 2 percent.
23
-------
SECTION V
COMBUSTION STABILITY
During the previous study(1) data available from a number
of tests in natural gas-fired utility boilers were examined to determine if a..
general mechanism could be identified which might account foi cases of ob-
served combustion instability, flame liftoff or both. Results of that exami-
natidri led to the tentative cónclusioi s hat (a) th combustion instability
wars vç ’ry likely an air-feed system coupled mode, and (b) the feedback cou-
I I
pling between air flow velocity through a burner and the degree of combustion
within the burner has a strong effect not only on the steady-state air-fuel
ratio within the bu rner, but also on both combustion instability andflame
liftoff. The experimental obsërvatfons on which these conclusions are basea
are discussed in the, earlier study (1). It was not within the scope of that
study, however, to do more than evaluate potential ñiechanisms from study
of available data Development of an analytical model of the suspèçted insta-’
bility mechanism and verification of analytical predictions a ainst this data,
were deferred to the subject study. This section reports these analytical
efforts.
5. 1 ANALYTICAL MODELING
The analysi’s Is divided into ’thr è components; (a) response
air flow through a burner to perturbations in furnace pressure at the burner
exit, (b) effects of pertàrbations in total flow and air-fuel ratio on furnace
pressure, and ’(c) coupling of the above two responses into a complete descrip-
tion of the combustion-air feed system coupled mode of instab ility. Details
of the entire analysi’s are contained in Appendix B.
5. 1. 1 Burner Air Flow Response
‘Figure 2 shows a schematic of the model of air flow through
a burner whi h was used as the basis for this part of the analysis. Major
assumptions include:
a. The windbox dimensions are small èompared to those of.
the’ furnace cavity. This results in the further assumption
that the air pressure in the windbox is uniform throughout
the windbox.
24
-------
CONSTANT ________
CONSTANT INLET INERTIA
WINDBOX (L12)
(R 1 )
Wb
FURNACE
PRESSURE
* VARIABLE
FLAME
RES I STANCE
(Rf)
Figure 2. Model of the dynamics of a burner in a utility boiler.
-------
b. The windbox volume is large compared to the volume flow
of air in and out of the windbox, at least over the period
of one cycle over the frequency range of about lOto 100 hertz.
This together with the previous assumption, results in the
further assumption that the windbox is a true plenum and its
pressure is constant both in time and space.
c. The resistance to air flow through the burner is concentrated
in only two places; the inlet loss through the air registers
and the exit loss resulting from the requirements for in-
creased momentum due to the presence of a flame within
the burner.
d. The axial length of the burne’rs is small compared to the
wavelengths of the oscillations of interest (up to 100 hertz).
As a result, the burner air can be treated as two simple in-
compressible lumps with all of the burner capacitance con-
centrated between. -
Assumption a. is certainly tr’ue in the horizontal direction
parallel to the burner flow axes since the windbox depth is normally less
than one-quarter of the furnace depth. Similarly, in the vertical direction,
the windbox height is normally 30-45 percent of the furnace height. In the
horizontal direction perpendicular to the burner flow axes (width), the wind-
box width may be quite comparable to the furnace width and assumption a.
may not be valid. The windbox, however, is actually cluttered with burners,
gas and oil feed lines, register control arms and linkages and other control
and instrumentation lines. As a result, transverse flow and/or resonances
within the windbox would be strongly damped.
In order to assure, with reasonable confidence, that all
burners are equally fed air from the common windbox, the windbox flow
area (and, therefore, volume) must be large relative to the sum of the
bu-rner flow areas. While some pressure measurements taken over a sin-
gle windbox indicate a marginal flow area in the windbox, it is probably suf-
ficient to justify assumption b.
Most burners are designed with ir registers which are in-
tended to impart some degree of swirl to the entering air flow. To generate
this tangential flow component, some pressure head must be converted to
velocity head. This is the definition of flow resistance. In addition, the air-
only flow cross-section within the burner is again cluttered with fuel feed
lines, diffusers and ignitors, while the flow c;oss-section after the fuel is
introduced is relatively free of flow obstruction. it is reasonable to lump
these pressure losses within the burner air-only flow region into a single
inlet resistance, as described in assumption c. Also, if there is a signifi-
cant flame within the burner, it would tend to be located primarily near the
26
-------
burner exit, within a short burner length, where the resistance to unheated
air-flow alone would be small. This part of assumption c., then, also seems
reasonable.
Assumption d. is simply a normal method of modeling com-
pressible flow in a pipe for purposes of dynamic analysis. Although not
assumed beforehand, it may even be adequate to model the air within the
burner as a single incompressible lump, at least for analyses within the
frequency range of interest.
Details of the development of the analytical expression for
burner response are contained in Appendix B. Major results are discussed
in Section 5.2.
5.1.2 Furnace Pressure
In Section 5. 1. 1, the variations in air flow through a burner
in response to perturbations in furnace pressure were described. As this
varying air flow (now with a constant fuel flow mixing into it) enters the fur-
nace it causes various three-dimensional perturbations in flow within the
furnace cavity.
Initially, any excess in flow velocity will be quickly damped
by momentum exchange with the relatively stagnant gases already in the fur-
nace. This deceleration can result in generation of acoustic waves which
would then propagate away from the burner exit in all directions at the local
speed of sound.
Combustion may be occurring simultaneously with this flow
deceleration. The mixing required for flow deceleration largely occurs at
the periphery of the main flow coming out of the burner (core flow) while that
required to complete combustion may be either within the core flow (combus-
tion at the air-fuel ratio of the burner) or at the periphery (combustion at
the furnace air-fuel ratio), or both. As a result, the average time delay to
complete combustion in most cases will be longer than, or at least equal to,
that required to damp the unreacted flow perturbations. These relative
delays will depend on the burner design, the fuel state and the difference
between the burner air-fuel ratio and that represented by the furnace gases
at the burner exit.
It seems clear, however, that the time delay to damp the
unreacted burner flow perturbations is short compared to the period of the
very low frequency oscillations of interest in full-scale utility boilers.
Therefore, it is reasonable to set the time delay for damping of burner flow
oscillations to zero and to assign a variable time delay (i ) to represent the
average time required for complete combustion.
27
-------
The effect of combustion is to change the local gas density,
both by raising the temperature and by changing the molecular weight between
the reactants and the products. These changes can also generate three-
dimensional acoustic waves which again propagate away at the local speed of
sound. Thus, the effect of a, single perturbation in flow out of a burner can
be to generate acoustic waves b 9 th instantaneously at the burner exit (flow
perturbation damping) and (usually), at some later time and spacial location
further out in the furnace (combustion). This model of dynamic burner flow
behavior, is shown schematically in Figure 3.
As discussed thus far, the effects of burner flow-perturba-
tions in generating furnace pressure perturbations are complicated, with
two sources of acoustic perturbations occurring at different times and.-spacial
locations, and three-dimensional acoustic response of the furnace cavity.
Consideration of complete three-dimensional wave equation solutions for the
furnace cavity, with driving at two (variable) sources would greatly compli-
cate analysis of.the overall coupled instability loop and very likely force a
nonlinear computer solution. In the light of the approximations and assump-.
tions discussed in Section 5. 1.1,. and other poorly known inputs, such as the
appropriate average combustion time delay, a complete solution is not j isti-
fied here. Perhaps such improvements could be incorporated at a later date
if necessary to obtain adequate useful solutions. Several simplifying assump-
tions are made here,,. both to provide ameans of coupling the dynamic, react-
ing burner flcw to the acoustics of the furnace cavity as well as to develop -
manageable .ps,eudo-acoustics for thethree-dimensional furnace cavity. The
major assumptions are discussed below..
a. Damping of burner flow perturbations is assumed to occur
in a furnace cavity control volume at the burner exit. The
time delay for this damping is short compared to the shortest
period of oscillation of interest and is, therefore, taken as
zero.
b. Although the .burner mass flow perturbations are damped
put immediately upon entering the furnace, the air-fuel ratio
perturbations (resulting from varying burner,air flow rates
mixed with a constant fuel-flow rate) remain and are carried’
farther out into the furnace cavity at a constant flow velocity.
After an average time delay, Tc the air-fuel ratio pertur-
bations result in perturbations in local gas temperatures
and molecular weights during (concentr,ated) combustion. It
is recognized that the combustion is actually distributed both
,in time and space and that the assumption of concentrated
combustion is in disagreement with the observation discussed
in Section 5. 1. 1 that some of the combustion has already
been completed even before the burner flow enters the furnace
28
-------
TIME DELAY TO
COMBUSTION
______ CORE
FLOW
1r
EXIT
— CONTROL
VOLUME I
ACOUSTIC WAVES
GENERATED BY
DAMPING OF BURNER
FLOW PERTURBATIONS
FLOW OF
UN REACTED
FUEL-AIR
MIXTURE
— — — — — —
COMBUSTION
CONTROL VOLUME
ACOUSTIC WAVES
GENERATED BY
VARYING AIR-FUEL RATIO
AND CONCENTRATED
COMBUSTI ON
AIR
‘ .0
— — — —
FUEL
‘At
fa
BURNER
RESPONSE
WINDBOX
(constant
pressure
plenum)
— —
AIR
FUEL
— —
Figure 3. Active burner flow and combustion model.
-------
cavity (partial combustion within the burner). This conflict
does not introduce significant error, since Section 5. 1. 1
treats the beginning of combustion while this section con-
cerns the average of complete combustion.
c. If the mass flow and combustion perturbations were intro-
duced into an infinite reservoir, only small acoustic pres-
sure perturbations would be generated. Instead, they would
tend to produce large oustic velocity perturbations. [ This
is because the specific acoustic impedance (the complex
ratio of acoustic pressure to velocity) of the gases in the
furnace cavity is low.] This observation is approximated
here by the assumption that any changes in pressure which
might tend to be generated by burner mass flow into, or com-
bustion withina control volume in the furnace cavity are
exactly negated by a proportionate mass flow out of the con-
trol volume, in acoustic waves. This implies zero specific
acoustic impedance.
d. In a finite enclosure, acoustic waves generated by burner
mass flow perturbations or combustion, according to assump-
tion c., can travel throughout the enclosure, reflect off solid
surfaces and return to the source. The impedances at the
surfaces of the control volumes, then, are no longer the spe-
cific impedance of the gas but depend on the location of the
control volume with respect to solid surfaces, the damping
of the acoustic waves in traveling through the gas and the
efficiency of reflections off the solid surfaces. Pressures
can now be generated within the control volume. This obser-
vation is approximated here by summing the flows into a con-.
trol volume which result from returning, reflected acoustic
waves. Pressures within the control volume result, through
the perfect gas law, only from these acoustic flows.
e. It is assumed that the spherical, three-dimensionai acoustic
waves generated within a control volume can be resolved into
independent plane waves traveling along the positive and nega-
tive Cartesian coordinates (Six directions). With such pseudo-
acoustics, allowance can also easily be made for acoustic
damping in wave travel as well as fox imperfect reflections
off solid surfaces.
f. It is recognized that the plane waves are taken to travel from
the control volume in which they were generated to a solid
surface, reflect and return to the volume. For this analysis,
the two control volumes described in assumptions a. and b.
and their acoustic interactions are considered, but acoustic
30
-------
interactions between burners are not considered. The
generated acoustic flow perturbations are divided equally
among the six plane waves. Each plane wave requires a
specific time delay to return to the source. The one excep-
tion to this is that wave which travels directly from the exit
control volume back into the burner and into the windbox. A
zero reflection is assumed for this wave.
Perhaps assumptions a. through f. should not really be called
assumptions. Rather they are simplifying approximations of complicated,
poorly known physical processes which are intimately involved in combustion
stability. The separation of the total driving effects of perturbations in a
gaseous, reacting flow into mass and reaction effects separated in time and
space is an extension of an observation first suggested by Dykema (4) and
developed further in applications to chemical laser combustors injecting
gaseous propellants. The remaining assumptions, c. through f. are neces-
sary to derive a simple set of driven, pseudo-acoustics for this three-
dimensional case which are compatible with the state-of-the-art of under-
standing of these and other related phenomena.
Based on the above assumptions and discussion, an expres-
sion for the response of furnace pressure (at the exit of a burner) to per-
turbations in the burner flow rate was derived. The details of that deriva-
tion are shown in Appendix B.
5. 1.3 Feed System Coupled Instability
A block diagram of the response of the total, feed system
coupled mode of instability is shown in Figure 4. The burner response
shown in the figure is that response discussed in Section 5. 1. 1 and described
by Eqs. (33) or (39) in Appendix B. The remainder of the forward loop shown
in the figure represents the furnace response discussed in Section 5. 1.2 and
described by Eqs. (73) and (74) or (75) of Appendix B. The total, open loop
response is simply the product of the burner response and the furnace re-
sponse. Because the feedback is negative, the conditions for the closed loop
to be unstable are that the phase angle between the output and input pressures
(Pf 0 /Pf 1 ) must be 180 degrees and the magnitude, or gain, of the open loop
should be greater than one. -
5.2 RESULTS
Results of interest here are of two types: (a) comparisons
with other analyses and with data to evaluate the validity of the analysis,
and (b) parametric studies of the effects of certain design variables on the
analytically predicted stability of a boiler. The usefulness of the latter
depends, of course, on the degree of agreement in the former.
31
-------
AIR/FUEL RATIO
EFFECTS IN
COMBUSTION
FURNACE
PSEUDO-
ACOUSTICS
Nomenclature defined on page viii
Figure 4.
Model of an air feed system coupled mode of combustion
instability in a utility boiler.
Wbd
COM BUST ION
TIME DELAY
“wb
=, K e1S
-------
5. 2. 1 Validity of the Analysis
5.2. 1. 1 Comparison with Previous Analyses
In Appendix B, the final open loop response of the feed system
coupled mode of instability analyzed here is compared to more conventional
analysis from the liquid rocket engine field. It is shown that the analytical
solution derived in this study yields that applicable to a liquid rocket engine
when appropriate simplifying assumptions are made. Since the rocket engine
chug analyses, such as those presented in previous studies (5), (6) and (7)
have been scrutinized, verified and used for many years. the analysis
developed here can also be considered generally valid, except where this
analysis represents an extension or modification of such analyses to fit the
current case.
A block diagram of a typical, simple feed system coupled
mode of instability in a liquid rocket engine is shown in Figure 5. Compari-
son of this diagram with that of Figure 4 broadly indicates the degree of ex-
tension or modification made in this study. The seven simplifying assump-
tions which reduce the block diagram of Figure 4 to that of Figure 5 are:
a. There is no flame in the burner (injector orifice) (Rfj = 0)
b. The burner (orifice) capacitance is negligible (C = 0)
c. The acoustic wave travel time from the region of con-
centrated combustion to the combustor exit is negligible
(STe = 0)
d. There is no effect of mixture ratio variations on heat
release rates and subsequent pressure variations (F(r) 0)
e. Wave travel times from the burner exit (injector face) to all
reflecting surfaces, r , are negligible
f. All acoustic damping and losses due to inefficient reflections
are negligible (K = 1)
g. Perturbations in mass flow from a burner (an orifice) db not
immediately appear as mass addition into the cornbustor
volume (implies liquids) but are delayed until the concen-
trated reaction occurs, at a time Tc later.
Further comparison of the current analysis with that of a
conventional liquid rocket engine analyéis can be made by comparing the
magnitude-frequency responses of the two analyses for a typical boiler
33
-------
COM BUST i N
:TIME DELAY
Figure 5.
Typical li uid rocket engine mod l of a low frequency (chug)
feed system coupled mode of combustion instability
(simplified from Figure 4).
wb
-------
configuration (defining some of the boiler parameters in terms of their
equivalent rocket engine parameters). For all parametric calculations,
the geometry of the radiant section of the boiler and the definition of the
distances required for wave travel and the reflecting surfaces are defined
in the schematic of Figure 6. The remainder of the geometry and nominal
full load operating conditions of this typical natural gas-fired utility boiler
are given in Table 2.
Table 2. Geometry and Full Load Operating Conditions
Lb = 1.5rn(5ft)
Ab = 0.5 rn 2 (5.24 ft 2 )
R. = 0.083 sec 2 /N-m 2 (0.0343 sec 2 /lbf-ft 2 )
a = 477 m/sec (1566 ft/sec)
c = 900 rn/sec (2950 ft/sec)
= 2.39 N (0.537 lbf)
= 0.9 (for all i)
=0
¶ = 0.04 sec
c
= 96.21 N/sec (21.63 lbf/sec)
Wfb = 10.43 N/sec (2. 344 lbf/sec)
No. of air-only burners = 8
No. of burners, total = 24
Location of air-only burner being investigated:
= 31.8 m (104.4 ft) L 4 = 6.86 m (22.5 ft)
L 2 = 4.15 m (13.6 ft) L 5 = 1.58 m (5.2 ft)
L 3 = 3. 20 rn (10. 5 ft) L 6 = 7. 56 m (24.8 ft)
L 7 = 9. 14 m (30.0 ft)
35
-------
K.
BOILER
RADIANT
SECTION
(fu mace)
(front view)
L 3 . L 4 .
-Q
BURNER
K 2
K 4
BURNER
Figure 6.
Schematic of the boiler analyzed-defining the acoustic
mode directions, lengths and reflection surfaces.
Kr 3
CON CE,NTRATED
COMBUSTI ON
ZONE
Kr 6
L 1
$
36
-------
To compare the more conventional liquid rocket engine feed
system coupled mode analysis to that derived in this study, all of the above
typical geometry and nominal operating conditions (except Tc = 0. 05 sec)
were used to calculate the magnitude of the open loop response over a range
of frequencies up to 50 Hz. These results are shown in Figure 7. Noted on
the magnitude-frequency curves are the frequencies where the open loop
phase shift is 180 degrees. If the magnitude of the response is greater than
one at these frequencies, the mode is unstable. The magnitude at other fre-
quencies has little meaning other than to indicate the potential for instability
if the phase shift around the open loop were proper (180 degrees). This par-
ticular calculation does not indicate the frequency ranges over which the
phase shift could ever be 180 degrees.
Clearly the two calculations yield similar results except that
the subject calculation appears to superimpose the furnace cavity acoustics
on the simpler rocket engine chug analysis. The superppsition, however,
is actually opposite to what might be expected. Figure 7 shows the frequen-
cies of the various natural resonances of the furnace cavity. The maxima
in the response magnitude curve all fall between, rather than at the resonant
frequencies. This is the result of the damping effect of the air mass addi-
tion at the exit of the burner. As discussed in Section 5. 1. 2, a decrease in
furnace pressure at a burner exit results in an increase in air flow out of
the burner. At any of the resonant frequencies of the furnace cavity, or at
very low frequency, this increased air flow from the burner will increase
the local pressure, effectively with no time delay, which in turn will “fill
in the trough” and damp the oscillation. Between the resonant frequencies
this effect of mass flow perturbations would be reversed, actually amplify-
ing the open loop response. Figure 7 shows all of these effects, with the
minima in the open loop response occurring at zero frequency (out-of-phase)
and at all of the resonant frequencies while the maxima occur at off-
resonant frequencies.
Figure 7 also shows that a simple, conventional chug analysis
would indicate only one instability, a strong one (gain of 3. 2) at 6. 7 Hz.
Based on gain alone, all frequencies below about 24 Hz could be unstable
while all above would be stable. The subject analysis indicates a similar
relatively strong instability (gain of 1. 4) but at a frequency nearly twice as
high (11 Hz) as the simpler analysis. The subject analysis also exhibits a
gain curve such that instabilities could occur at any frequency below about
22 Hz. Unlike the simpler analysis, however, the subject analysis indicates
that the furnace cavity acoustics can modify the gain curve such that insta-
bilities are possible above the 22-24 Hz range. Since this latter prediction
results from the modifications introduced in this study, an experimentally
observed instability in this boiler in the 40-50 Hz range would tend to further
verify those modifications. It will be seen in the following paragraphs that
an instability was observed in this frequency range.
37
-------
NOMINAL OPERATING CONDITIONS
TC
I-J
I—.
LU
1<
I jLi
I
I
I C J
-J
I
0.05 sec
Figure 7.
Comparison of results of the current analysis with those of
conventional rocket engine analysis.
6
1
C-)
C-,
C
(I )
-
+
-J
-J
I-
LU
U-
-J
L U
U-
-
1
0
1=
.c J
±
—J.
c’J
-J
+
- I
LU
‘I)
0
0
LU
0
0
0
-J
LU
C
LU
U-
0
LU
I-
00
CONVENT I ONAL
ROCKET ENGINE
\ ANALYSIS
IS
.10
38
-------
5.2.1.2 Comparison with Data
The primary data normally used to verify stability analyses
are the measured frequencies and relative magnitudes of instabilities, com-
pared to those analytically predicted for those hardware and operating con-
ditions. Obviously, no operator of a full-scale utility boiler wants to operate
his boiler unstably in order to develop data to verify an analytical model.
The potential danger and financial loss should the instability get out of con-
trol is too large. Fortunately for this study, one such case of a violent
instability occurred in which not only were the operating conditions reason-
ably well defined but the acoustic environment (outside of the boiler) was
tape recorded for further frequency/amplitude analysis. The major char-
acteristics of this instability were as follows:
a. The most violent instability exhibited maximum amplitude
at about 12. 5 Hz. Mild vibrations were also observed at
about 43 Hz.
b. Maximum instability appeared to occur with 25 to 30 percent
of the burners operating air-only. Larger or smaller frac-
tions of burners operated air-only appeared to result in
lower amplitude oscillations.
c. All of the instabilities occurred when a special set of gas
injection spuds were being used. These spuds were spe-
cifically designed to increase the rate of mixing of the gas
with the combustion air, very likely resulting in greater
combustion within the burners. These gas spuds were sub-
sequently removed from the burners and the instabilities
essentially disappeared.
Any other observations are clouded by the fact that the boiler
was being operated in a start-up mode at the time of the instabilities with no
data taken on total air flow and with little steady operation. A large number
of burner configuration, load and other operating condition changes were
made simultaneously in an attempt to get the boiler started without damage.
Figure 8 shows plots of the gain curve (magnitude of the open
loop response versus frequency) calculated using the subject analysis for all
of the nominal hardware and operating conditions in Table 2 except that three
values, nominal and ±10 msec, of the combustion time delay are used. These
three values are used because the true value appropriate to these combustion
conditions is not known, as well as to show that the assumed value of this
delay, not at all related to the furnace cavity resonances, has a strong and
varied effect on the maxima in the three frequency ranges of interest.
39
-------
NOMINAL OPERATING CONDITIONS
® MEASURED FREQUENCIES ON TEST WITH
PARTICULARLY VIOLENT VIBRATIONS
4)1800 PHASE SHIFT
AT THIS FREQUENCY
COMBUSTION TIME
DELAY, milliseconds
M I LD
Figure 8. Comparison of analytical predictions with
experimental observations.
30
40
50
EST IMATED
RANGE OF
SEVERE
u-J
C
LI,
uJ
C
C
-J
C
LU
=
0
LU
0
z
VIBRATIONS
5
4
3
2
1
0
0
EST
IMATED
I
RANGE OF
VIBRATIONS
10 20 30 40 50
FREQUENCY, Hz
40
-------
In the low frequency range, an instability is indeed analytically
predicted, in the frequency range around ii Hz. Frequencies for maximum
amplitude were estimated by the engineers in the 10-12.5 Hz range, with the
one instability measured in this frequency range at 12. 5 Hz. The frequency
and stability agreement is considered good, particularly when compared to-
the 7 Hz instability predicted by the simpler analysis (Figure 7).
In the highest frequency range an instability is not predicted,
(the gains are less than one at the frequencies where the phase shifts are
180 degrees) but if the gain were high enough the instabilities would be in the
43-45 Hz range. Instabilities were, in fact, identified by the engineers in
the 40-50 Hz range, with the one instability measured in this frequency range
at 43 Hz. The frequency agreement is considered excellent. The gain cal-
culation appears to be in error, since it is less than one (stable) at the fre-
quency for 180 degrees phase shift. The calculation does show, however,
that a gain greater than one is possible in this frequency range (at 46 Hz)
and that the maximum gain in this range is very sensitive to the combustion
time delay assumed. It is possible that just slightly different operating con-
ditions could yield a gain greater than one at the appropriate frequency.
Figure 7 indicates that the simpler analysis would predict soundly stable
operation over this whole frequency range.
Observations from the test where the frequencies were mea-
sured indicate that the 12. 5 Hz vibrations were very violent while those at
43 Hz were relatively mild. The relative calculated gain in those two fre-
quency ranges are in agreement.
Agreement in the intermediate frequency range (27-36 Hz)
is more questionable. No experimental estimates or measurements indi-
cated any instabilities in this frequency range. The calculations, although
they do not actually indicate an instability, do indicate a potential for insta-
bilities intermediate in both magnitude and frequency between the ranges
discussed above. Figure 7 shows that in this intermediate frequency range
relatively high gain results from the first and second harmonics of the verti-
cal (L 1 + L 2 ) mode while the gain maxima in both the lower and higher fre-
quency ranges result from fundamental resonances.
The calculations shown in Figure 8 represent the response of
a burner very low in the furnace cavity (see L 1 and L 2 in Table 2), where
the pressure oscillations in all of the vertical resonant modes of the furnace
cavity, and the pressure coupling with burner flow, are maximum. In the
actual case the burners are spread vertically from about 10 to 30 percent of
the vertical height of the furnace cavity. Figure 9 shows that the loop gain
is maximum in both the low and intermediate frequency ranges when the
burner is located at the lowest level. Inthelowerfrequency range (15-22.5 Hz
41
-------
—
C J
L i
.0
IC-)
I
0
• 0.2 0.25
OF THE VERTICAL HEIGHT
U
0
C-)
0
1<
27.5 35 Hz
• 0.15
FRACTIQN
Figure 9. Variation of the maximum gain within the low and
intermediate frequency ranges for burners located
at various vertical positions
0.3
2.0
LU.
(P1
0
0 -1.
(#1
LU
0
0
0,
•1
I L.
L i i
0
0
LU
08
0
LU
I-
15-22.5 Hz.
POTENTIALLY UNSTABLE _____ - _______
STABLE
0.4—
0—
0. 1
42
-------
between the fundamental and the first harmonic of the vertical resonance)
the response decreases with increasing vertical height but is always rela-
tively high and always potentially unstable. In the intermediate frequency
range (27. 5-35 Hz), between the first and second harmonics, the loop gain
decreases rapidly in the higher levels and burner locations above about 0. 2
of the vertical height are stable in this frequency range. As a result, the
appropriate total response of all of the active burners in a real boiler may
always be less than one (and stable) in this intermediate frequency range.
This interpretation would be in agreement with the limited experimental
observations with this boiler.
The variation in loop gain for burners at various vertical
positions can be explained simply from the pressure coupling. Since it is
the furnace pressure oscillations at the location of a burner exit that cause
the variations in air flow through the burner, pressure coupling is greatest
at the pressure anti-nodes of the resonant mode and is zero at the nodes.
Table 3 lists the locations of the anti-nodes and nodes in the first three of
the vertical resonant modes in the furnace cavity.
Table 3. Anti-Nodes and Nodes in the Vertical Resonant Modes of
the Furnace Cavity
Fraction of the
Vertical Height
Pressure Anti
-Node (A) or Node
(N)
Fundamental
1st
Harmonic
2nd
Harmonic
0
A
A
A
1/6
-
-
N
1/4
-
N
-
1/3
-
-
A
1/2
N
A
N
2/3
-
-
A
3/4
-
N
-
5/6
-
-
N
1
A
A
A
43
-------
The lowest pressure nodes in the first and second harmonics
of the vertical mode are also shown in Figure 9. Th node between these
two harmonics, then, should fall at about 21 percent of the vertical height,
approximately at the location, in Figure 9, of the minimum in the gain in
that frequency range. The average .node between the fundamental and the
first harmonic, howev r, is off scale in Figure 9, at about 37; 5 percent of
the vertical height. Nevertheless, the fig ire shows the gain in the lower
frequency range decreasing toward that node.
As a result, it seems reasonable to interpret tk ë abovécä.l
culations to indicate that the net total response of an array of burners spread
over a vertical distance large compared to the wavelength of the 2nd harmonic
of the ertical resonance would tend to be stable in the frequency range be-
tween the first and Second harmonicé. It is also reasonable to iinply that .this
vertical spreading of the burners wth ld have little effëct oh stability In the
frequency range between the fundamental and fi- st harmonic resonances
(15-225Hz) and no ‘effect on stability in the frequency range of the trans..
verse (horizontal) r sonances (45 - 49 Hz). Experimental data confii nis
all of these observations.
- - - As a further checkS that the gain maxima discussed here and
shown in Figure 9 are indeed related to the vertic’ai acoustic resonances,
the acoustic reflection coefficients at the top and bottom of the furnace,
affecting only the vertical mode, were analytically set to zero. These re-
sults are shown in Figure 10. Clear,ly the absence of acoustic reflections
severely reduces the low and intermediate iTlaxima.
-. It ma r well be that these ‘obsèr tiôi s on the relative stability
of this boiler in these three frequency ranges represent the most fundamen-
tal verification of the subject analysis. The analysis not only reasonably
accurately predicts the frequency ranges and relative sta i1ity of the low and
high frequency instabilities but also correctly predicts the relative’stability
iri the intermediate frequency range. The only alternativ explanation known
to the writer for the absence of instability in the intermediate frequency
range might be that related to the Helmholtz acoustic absorber represented
by the ash pit of this boiler (1). Total (analytical) elimination of reflections
from the bottom of the furnace, however, still leaves a fairly high and poten-
tially unstable gain ih the intermediate frequency range.
The second source of experimental verification of the subject
analysis is the previously discussed observation that the maximum instability
appeared to occur with 25 to 30 percent of the burners operating air-only.
The fractio i of the burners which are operated on air-only has a strong effect
Oi l NO emissions. Therefore, that parameter will be discussed in the next
section (5. 2. 2), where the effects of major combustion modifications on
stability are evaluated. The reason that the boiler appeared to operate
44
-------
5
w
Cr,
0
0
Cr,
u - I
a-
0
0
-J
u - I
0
u - i
=
I-
U-
0
u - I
ACOUSTIC REFLECTIONS
AT THE TOP AND BOTFOM
OF THE FURNACE CAVITY
9
4
3
2
1
00
10 20 30 40
FREQUENCY, Hz
Figure 10.
50
Effects of acoustic wave reflection efficiency.
45
-------
somewhat more stably when more than 25 to 30 percent of the burners were
operated air-only is thought to result from the more stable position of the
flame, deep within the burner.
This same variable, the fraction of combustion completed
within the burner, is also thought to be at the root of the violent instability
in this boiler in the first place. All of the natural gas burner spuds were
changed just prior to the observed instabilities to incorporate a new design
specifically intended to promote rapid mixing of the fuel with the air within
the burner. It seems reasonable that this more rapid mixing would cause
more of the combustion to be completed within the burners. No other ,change
or effect was apparent. After two attempts to start the boiler and to achieve
rated load, during which the violent instabilities occurred, the original gas
spuds were returned to the burners and stable operation was restored.
In the subject, as in the previous (1) analysis, the fraction of
combustion occurring within the burner was. described by a function of-the
type:
C=K 3 i (1)
where, with the standard gas spuds, K 3 a nd n were estimated at 0.777 and
0.5, respectively. Equation (1), with these values for the constants, .was
used for nearly all of the parametric stability calculations. It is not known
exactly how the form of Eq. (1) might change-with more rapid fuel-air, mixing
within the burner. It seems reasonable, however, ‘that a flame which’ is long
compared to the burner length (from the gas spuds to the exit) would not move
substantially in and out of the burner with’ burner flow velocity changes. Con-
versely, a flame which is short compared to this burner length might easily
move completely inside the burner with normal burner flow velocity vária-
tions. Thus Eq. (1) for the slow-mixir g flame might involve a nearly zero
exponent (-n) while,a rapid mixing flame might involve much more negative
slopes (l3rger values of n). For very rapid mixing the flame might be com-
pletely, contained within the burner at all.flow velocities and the exponent (n)
might again be zero.
A series of parametric calculations were made to investigate
the effect of the negative slope of Eq. (1) on stability. For these calculations
the total burner resistance at a given burner air flow rate was held constant
while the exponent n was varied from 0 to 2. The value of the constant, K 3 ,
then, also had to be varied:’ ‘ - ‘
Figure ii shows plots of the low frequency open loop response
for values of the exponent ranging zero to two. At very low frequency (near
zero) the effect of the exponent variation is to increase the response by more
46
-------
FREQUENCY, Hz
Figure 11.
Effects of the fraction of combustion
completed within an active burner on
combustion stability.
n=2
w
‘I ,
0
u - I
0
0
0
-J
0
0
UJ
U-
0
u -I
z
10
8
6
4
2
00
4 8 12 16
20
47
-------
than a factor of two as the value of the exponent was increased from zero to
two. At the measured unstable frequency (12. 5 Hz) the increase in response
is only about 25 percent. With an exponent of zero, which implies that air
flow velocity changes would have no effect on burner resistance or stability,
the figure shows that the open loop is still potentially quite unstable (gain
greater than one).
In general the subject analysis tends to support the observa-
tions regarding the effect of combustion within a burner. First, in cases
where combustion is partially completed within the burner, the burner boiler-
loop tends to be more unstable. Second, in cases where combustion is al-
most totally completed outside or downstream of a burner (oil- and coal-
fired boilers) or inside of a burner (very rapid mixing with very low air flow
velocity) the boiler tends to be more stable. The relatively high gain, rela-
tively unstable system predicted even with the zero exponent is somewhat
surprising since instabilities are often observed with the fast-mixing and
burning gaseous fuels but almost never with the oil or coal fuels. This
writer, however, is aware of at least one oil-fired boiler which not only
exhibits repeatable instabilities but the instabilities appear to exist only
when the flame is visually observed to be partially within the burner (rather
than distinctly separated from, and downstream of, the burner). T-here
appears to be more to learn about the effect of this parameter on stability.:.
of utility boilers.
5. 2. 2 Further Parametric Calculations
Some of the parametric calculations made in this study have
been discussed in the previous section. The purpose there was to compare
results with experimental observations to confirm the validity of the analysis.
In this section it will be assumed that the comparisons of the previous sec-
tion have reasonably validated the analysis and the results of some of the
parametric calculations will be reviewed for the purpose of evaluating and
explainingthe effects of certain ir dependent variables on stability.
The primary purpose of the overall EPA grant, of which this
analysis of combustion instability in utility boilers is a part, is to evaluate
practical techniques for the control of NO emissions by combustion modi-
fication. Stable combustion is a necessary requirement over the full range
of modified combustion conditions. In previous studies (1) and (2), it was
determined that the singlemost effective combustion modification technique
for NOx control with natural gas, oil or coal-fuels is to concentrate the fuel
flow into a fraction of the burners,, letting the remainder of the burners (and,
perhaps, special NO ports) operate with air-only. The active burners, then,
would operate very fuel-rich. These previous studies have also indicated that
the fuel-rich active burners should be located low in the burner array for
maximum NO reduction. There are also indications from other studies that
48
-------
fuel-air mixing within the burners can have a significant effect on NOR.
For these reasons, the effects of three major combustion modifications on
combustion stability are discussed in this section: (a) the fraction of burners
operated air-only; (b) the vertical location of air-only burners, and (c) the
fraction of combustion completed within a burner. For convenience, NO
ports will be considered an equivalent air-only burner.
s. 2. 2. 1 The Fraction of Air-Only Burners
The main effect of shutting off the fuel to some of the burners
and diverting it to others is to provide an initial fuel-rich region in the boiler
f n which the initial hydrocarbon-air reactions can take place. This minimizes
both the conversion of fuel-bound nitrogen to NO as well as the formation of
NO by thermal mechanisms, at least in this initial combustion stage. Indi-
cations from the studies summarized in Ref. 3 are that it is desirable to
operate the active burners as fuel-rich as possible, consistent with the
flammable limit and with stable combustion.
The main effect of the burner air-fuel ratio on combustion
stability is in the amplification of heat release and mole change perturba-
tions resulting from air-fuel ratio perturbations at mean burner air-fuel
ratios well away from stoichiometric. This amplification is discussed in
Appendix B and is represented by the function F(r), defined by Eq. (B-55).
Equation (B-73) shows that F(r) has no significance at very low frequencies
(approaching zero) but can represent a significant gain term, or destabiliz-
ing influence, at the frequencies of interest here. Figure i2 shows a plot
of the approximation analytical expressions used in this study to represent
Fr). The data for these approximations were derived from equilibrium
combustion calculations performed at The Aerospace Corp.
The figure shows that values of F(r) apparently become
strongly positive at air-fuel ratios lower than about 90 percent of stoichi-
ometric (fuel-rich). It seems likely that the value of F(r) could even become
infinite as the flammable limit is approached, although the equilibrium com-
bustion calculations could not predict this. Depending on mixing within the
burner and the recirculation necessary to anchor (provide continuous ignition
for) the flame, the flame could repeatedly blow off the burner (lift off) and
flash back or re-attach. In such a case, the heat release at the burner exit
would alternate between zero and some significant, finite level. Equations (B-76)
through (B-87) in Appendix B show that for very large values of the function
F(r), the magnitude of the open loop response is directly proportional to F(r)
and the open loop response would in turn be very large.
Thus, the onset of instability and the incidence of flame lift-
off could both be rather simply related to a burner air-fuel ratio, near to the
flammable limit. For natural gas, the fuel-rich flammable limit is an air-
fuel ratio about 62 percent of stoichiometric. With a boiler operating at
49
-------
+3
.F( r) = 0. 0911 - 0.00631 : rb
,, /STOICHIOMETRIC
C J
+
)<
U-
+2
+1
0
—1
-2 —
-3
1(r) .= -0.0361 + 0. 000517 rb.
I — I
10 14 18 22 26. 30
rb
I
Figure 12. Analytical expressions used to approximate
the function F(r).
50
-------
about 15 percent excess air, and taking into account the increased resistance
to air flow as a result of partial combustion within the active burners, the
rich flammable limit could be approached in the active burners with about
one-third of the burners operated air-only. Vibration data from two boilers
firing natural gas fuel appear to indicate an onset of undesirably high vibra-
tions with about 25 to 30 percent of the burners air-only. Some subjective
observations of the flames at the burner exit operating under these conditions,
however, tend to indicate that combustion is in fact occurring at air-fuel
ratios very near to the rich flammable limit (1). Figure 13 shows an esti-
mate of the actual burner air-fuel ratio, as a function of the fraction of the
total burners that are active, for the boiler used as the nominal in this study.
Figure 12 also shows that fairly large negative values of F(r)
are calculated at air-fuel ratios typical of the overall boiler (or of the burners
when all burners are active and no NO port air flow exists). Operation at
about three percent oxygen in the flue gases represents about 15 percent
excess air. At low frequencies the negative gain would tend to have a stabi-
lizing effect but at higher frequencies the effect would depend on other phe-
nomena such as the acoustic time delays and could be destabilizing. At very
high (lean) air-fuel ratios, regardless of whether the flameholding is adequate,
the values of F(r) begin to approach zero and additional gain from this source
tends to become negligible. The large dilution of any heat release and mole
change variations by the excess air very likely damps any subsequent destabi-
lizing effects.
If one neglects the extreme cases of F(r) that might be possible
at air-fuel ratios near the rich flammable limit, the effects of the fraction of
the burners operated air-only do not appear large. Figure 14 shows that the
open loop gain at the unstable frequency (10-11 Hz) increases as the fraction
of the burners operated air-only increases but over the range from zero to
about 38 percent of the burners air-only this unstable response increases by
only about 15 percent. Similarly, Figure 15 shows that although increasing
the fraction of air-only burners from zero to half increases the maximum
gain in the 30-3 5 Hz range by more than 80 percent, it actually decreases
the maximum gain in the 15-20 and 45-50 Hz ranges.
Thus, it appears that one of the major combustion modifica-
tion techniques for the control of NO can be accomplished without signifi-
cantly increasing the potential for combustion instability, as long as the
burner air-fuel ratio is maintained above the rich flammable limit. It is
probable that the burner air-fuel ratio, on the average, could be safely re-
duced even below the rich flammable limit if care were exercised to main-
tain a solidly anchored flame, avoiding any flame liftoff.
5. 2. 2. 2 Vertical Location of Air-Only Burners
Some of the parametric calculations regarding the effect of
this parameter have been shown and discussed in Section 5. 2. 1. Figure 9
51
-------
NO PORTS
CLOSED
OPEN
HAYNES 516, FULL LOAD
NATURAL GAS, 3% 02
C l) ORIGINAL
CONFIGURATION
®MODIFIED FOR LOW NO
ACTIVE BURNERS, FRACTION OF TOTAL
Ffgure 13. Burner air/fuel equivalence ratio with/without
flames in burners.
1.0
1.2
1.0
I-
w
C-,
08
>
= ,
Li J
—I U.
w.
0.4
L7I
NO FLAME
IN BURNER
APPROX RICH FLAMMABLE
LIMIT, PRE-MIXED GASES
2
0.5 0.6 0.7 0.8 0.9
-------
2.
i.
a-
0
0
-J
L&J
a-
0
=
I-
0
L&J
I-
1.
1.
1.
Figure 14.
Effects of the degree of burners-out-of-service on
instability at the unstable frequency.
‘3
0.1 - 0.3 0.4
FRACTION OF BURNERS AIR-ONLY
0.5
53
-------
FREQUENCY, Hz
Figure 15.
Effects of the degree of burners-out-of-service
on instability at various frequencies.
5
4
‘I
‘S
ALL BURNERS
ACTIVE
0
cL
0
0
0
LLi
0
0
LU
=
U-
0
UJ
0
HALF OF THE.
BURNERS AIR-ONLY
I ’
0
10
20
30
40
50
54
-------
for example shows that the maximum gain of the open loop occurs when the
active burner is located as low as possible in the burner array. As dis-
cussed previously [ (1) and (2)] this is exactly where the active burners should
be to minimize NO emissions, at least in the case of oil and coal fuels.
Dykema (1) discusses the possibility of minimizing NO emissions when
burning gaseous fuels (containing no fuel-bound nitrogen) by locating all of
the air-only burners as low as possible in the burner array or by providing
a configuration with all air-rich (fuel-lean) active burners in conjunction with
fuel-rich NO ports. Although both of these latter configurations have only
been suggested, and have not been adequately demonstrated, both represent
more stable configurations than those with fuel-rich active burners in the bot-
tom of the burner array. In cases where the latter configuration is necessary,
care will have to be taken to maintain combustion stability.
5. 2. 2. 3 Flame Within the Burner
Operation of a boiler with all or part of the combustion taking
place within the burners is not a consciously recognized NO reduction tech-
nique. From study of both the steady state and the vibration data of Ref. 1 it
appears that combustion within a burner has a significant effect on both NO
emissions and on combustion stability.
At least some initial combustion in the burner or near the
burner exit is necessary to provide an anchor, or a continuous ignition
source, when the fuel is gaseous. It may also be desirable in some oil-
fired configurations, particularly when atomization is fine and vaporization
and burning is rapid.
The discussion in Section 5. 2. 2. 1 and in the previous study (1),
as well as the calculations plotted in Figure 13, show why it may sometimes
be thought desirable to maintain a significant flame within the burners. The
primary effect of combustion within a burner is to increase the resistance to
air flow through the burner which, in turn, reduces that air flow rate and
decreases the burner air-fuel ratio. In general, a lower burner air-fuel
ratio reduces NO emissions. As a result, low NO emissions are observed
with the flame in the burner; high NO emissions are observed with the flame
outside of the burner and the general conclusion is (sometimes) drawn that a
flame is necessary within the burner to achieve low NO emissions.
On the other hand, the incidence of high combustion vibration
is much higher when a partial flame is anchored within a burner. Nearly
all cases of excessive vibrations of the type described herein, at least within
the experience of the writer, have occurred in natural gas-fired boilers,
where the flame almost must be anchored in the burner. In the one (undocu-
mented) case of such vibrations in an oil-fired unit of which the writer is
aware, the vibrations reportedly were present when the flame was said to be
55
-------
in the burner exit but absent when the flame was somewhat downstream from
the exit. From the standpoint of combustion vibrations, then, it might be
concluded that it is necessary to avoid the presence of a flame within a
burner to avoid excessive combustion vibrations.
The calculations shown in Figure 13, however, show that a
flame need not be present within a burner to achieve a given active burner
air-fuel ratio and the.related NO reduction., The effect of flames within the
active burners is to increase the resistance to air flow through the active
burners relative, to the air-only, burners and: to divert combustion air flow
away from the active burners and through the air-only burners. This also -
increases the overall resistance of the total burner array, requiring more
fan energy to drive the combustion air into the furnace. Ofteri,-attempts to•
reduce NO emissions by concentrating the fuel flow in a-fraction of the
burners result in “running out of fan t - rated load cannot now be achieved
because the combustion air fans cannot drive air across the burner array.
at that rate.
With no flames within the active burners, the air flow resis-
tances of active arid air-only burners and the overall resistance of the total
burner array are essentially independent of the fuel flow rates in the indi-
vidual burners. The fuel flow must now be ‘concentrated in a smaller frac-
tion of the burners to achieve the active burner air-fuel ratio, but once this
is done the NO reductiqn should be the same.(dis regarding the slightly dif-
ferent mixing and burning pattern in the furnace with the fuel concentrated
into a smaller number of active burners).
Most of the above steady-state effects of partial combustion
within a burner were first derived and discussed in the studies reported by
Dykema (1). The subject study has added to this some underst’anding of the
dynamic .effects., Gene rally, the presence of a flame within an active burner
represents a destabilizing influence on the boiler combustion. This is par-
ticularly true if the active burner air-fuel ratio is reduced to near the rich
flammable limit of the fuel. The -studies reported here, however, show that
the presence of the flame within the burner does not guarantee unstable or
vibratory, corn bus tion.
A particularly important result of this study is the observ’a-
tion that the absence of any influence of the flame on’ air flow through the
active burners does not in itself assure stable combustion. Note that the
open loop response shown in Figure 11 for,this case (n = 0) still shows sig-
nificant potential for instability (response greater than one). Care must be
taken to design for and, avoid combustion instability in all cases, even if the
flame is fully outside.of the burner and/or solidly anchored. This also im-
plies that similar ins,tabilities can occur in oil- and coal-fired boilers as
well.
56
-------
Generally the presence of a flame within the active burners
has two negative effects: (a) it has a destabilizing effect, particularly as the
active burner air-fuel ratio is reduced near to the level of the fuel-rich
flammable limit for the fuel; and (b) it unnecessarily increases the overall
resistance to air flow through the total burner array. Since there appear tobe
no c irect compensating positive effects, (steady-state air flow resistance can
be more reliably controlled with the inlet resistance) it is probably best to
design for minimum or no flame within the active burners. As in all combus-
tion systems (8), there is no dynamic substitute for high, constant (inde-
pendent of fluid flow) resistance across the burners to maintain stable
combustion, but this is probably better achieved mechanically than as a result
of partial combustion within the active burners.
57
-------
REFERENCES
1. 0. W. Dykema, Analysis of Test Data for NO Control ih Gas- and
Oil-Fired Utility Boilers , EPA-65072-75-012 (NTIS PB 241918/AS),
U.S. Environmental Protection Agency, Research Triangle Park,
N.C., January 1975.
2. 0. W. Dykema, Analysis of Test Data for NO Control in Coal-Fired
Utility Boilers , EPA-600/2-76-274 (NTIS PB 261066/AS), U.S. Envi-
ronmental Protection Agency, Research Triangle Park, N. C.,
October 1976.
3. 0. W. Dykema and R. E. Hall, “Analysis of Gas-, Oil- and Coal-
Fired Utility Boiler Test Data,” proceedings of the EPA Symposium
on Stationary Source Combustion , EPA-600/2-76-152c,
(NTIS PB 257146/AS) June 1976.
4. E. H. Manny, W. Bartok, A. R. Crawford, R. E. Hall, and J. Vatsky,
“Studies of Waterwall Corrosion with Staged Combustion of Coal,”
Presented at the International Conference on Corrosion and Deposits
from Impurities in Combustion Gases, New England College, Henniker,
N.H., June 1977.
5. 0. W. Dykema, “Feed System Coupled Instability in Gas/Gas Corn-
bustors,” proceedings 11th JANNAF Combustion Meeting - VII , CPIA
Pub. 261, p 51, September 1974.
6. M. Summerfield, “A Theory of Unstable Combustion in Liquid Pro-
pellant Rocket Systems,” J. Amer. Rocket Soc. , September 1951.
7. L. M. Wenzel and J. R. Szuch, Analysis of Chugging in Liquid-
Bipropellant Rocket Engines with Different Vaporization Rates ,
NASA TN D-3080, October 1965.
8. JANNAF Working Group on Combustion, Design and Development
Procedures for Combustion Stability in Liquid Rocket Engines,
0. W. Dykema, Committee Chairman, CPIA Pub. 256, September
1974.
58
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APPENDIX A
EFFICIENCY LOSSES
Five areas of efficiency losses are addressed here,
representing energy losses resulting from:
a. Incomplete combustion
b. Uncondensed water vapor in the flue gases
c. The high temperature of the flue gases entering the stack
d. Steam turbine expansion processes
e. Electrical generating and control equipment
The reference input energy rate is taken as the high heat of combustion. By
definition, the high heat of combustion is released and available to do work
in a boiler only if all the carbon and hydrogen in the fuel are oxidized to car-
bon dioxide and water and the resulting combustion products are cooled to
their initial ambient temperature. The first three of the above losses repre-
sents departures (heat losses) from this ideal. When these heat losses are
subtracted from the high heat input rate, the remainder is essentially the
rate at which heat enters the boiler steam cycle.
The electrical equipment is also not 100 percent efficient.
After subtracting these electrical losses from the measured electrical load,
the remainder is essentially the energy output rate of the steam turbine.
The ratio of this output rate to the rate at which heat enters the boiler steam
cycle represents a rough measure of the steam, or Rankine, cycle efficiency
of the boiler. The Rankine cycle of a modern utility boiler is usually compli-
cated by various reheat and regenerative partial cycles and other means
designed to maximize performance. Since the prime purpose of this study
was and is NOx control, not enough data were gathered to allow detailed
direct evaluation of losses, and changes in losses with combustion modifi-
cations, throughout the steam cycle.
The largest single steam cycle efficiency loss is that inherent
in the thermodynamic process (ideal cycle efficiency). We have no interest
59
-------
in this loss in this study except to note that it will be different for boilers
designed for different steam pressures and temperatures. The next largest
loss, however, is usually a result of nonadiabatic expansion through the
steam turbine. In most cases, the steam turbine is designed for maximum
efficiency in the expansion process under the most common operating condi-
tion and is less efficient under otheç conditions. For a utility boiler, it is
reasonable to expect that maximum expansion efficiency may occur at about
80 percent of rated load. Efficiency losses at other loads largely result
from the need to control the power output of the fixed geometry, constant
speed turbine by controlling the thermodynamics of the steam. It is also
reasonable, then, to assume that a large part of the variation in efficiency
of the turbine with load would result from the steam power control technique
(for example, throttling), essentially ind ependent of the profile of the heat
flux into the steam. These losses would result from the need to vary the
boiler load (electrical output) independent of any combustion modifications
introduced for the purpose of controlling NON.
Thus, by calculating efficiency I osses in four of the five areas
mentioned above, the major known losses can be accounted for, leaving a
residual efficiency which can contain all the remaining variations in efficiency
which cannot easily be analyzed. The efficiency loss calculations are devel-
oped in Sections A. 2 through A. 5 of this appendix.
A. 1 STOICHIOMETRY
Before many of the efficiency losses for which we want to
account can be calculated, various quantities rnust be derived from stoichi-
ometry. These quantities are derived in this section and cast in a form for
calculation by computer.
Writing a generalized hydrocarbon fuel molecule as
CaHbOcNdSe and considering only the major combustion products or those’
of special interest to efficiency or air pollution, stoichiometry canbe
written
(CaHbOcNdSe) +,X(0 0 42 N 1 . 57 ) —A(C) + B(CO) + C(C0 2 )
+ D(H 2 0) + E(0 2 ) + F(N 2 ) + G(NO) + H(S0 2 ) (A-i)
It would probably be desirable to include free hydrogen in the products, but
no data are available on hydrogen concentrations in the flue gases.
Samples of the combustion products are usually analyzedafter
all the water vapor has been condensed and the remaining gases thoroughly
60
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dried (dry basis), and concentrations are normally cited as volume or mole
fractions of the total sample. The product specie concentrations then can be
defined by (e.g., for C0 2 )
[ co 2 ] = SMTD (A-2)
where the square brackets [ ] denote a concentration, SMTD is defined as
the sum of the moles of combustion products, excluding the water (dry) per
mole of fuel, and C is the coefficient of CO 2 in Eq. (A-i). Equation (A-2)
defines that the concentration represents a volume or mole fraction, dry.
There are three parameters of interest to efficiency and air
pollution studies which can be derived from flue gas analyses using Eqs. (A-I)
and (A-2):
a. SMTD
b. Overall air-fuel ratio
c. Heat of combustion
The theoretical stoichiometric relation of CO 2 to 02 is also useful. From
material balances in Eq. (A-I), expressions for these parameters are
MTD- NUM 3
S - DENOM (A-
NUM = 0. 9406b - 1. 8811c + 0. 5d + 4. 7623e (A-4)
DEMON = 1 - 2.8811 [ CO] - 4.7623 [ CO 2 ] - 4.7623 [ Oz] - 2.388 [ NO] (A-5)
X = 2.382i ( .b - c + Ze) + ( [ Ca] + 2 [ C0 2 ] + 2 [ 0 ] + [ NO])SMTD (A-6)
AHh = - (AH + 34.ZOb + 69.3e) - (26.42 [ co] + 94.38 [ C0 2 ]
- 21.5 [ NO])SMTD (A-7)
1 a(0.2099 -
°2Jth = a + 0. i975b - 0. 395c + 0. 105d + e (A-8)
61
-------
The constants in Eq. (A-7) were obtained using the following heats of
formation:
K-cal
Heats of Formation,
Specie g-mole
Fuel AH
CO -26.42
CO 2 -94.38
NO +21.5
SO 2 -69.3
H 2 0 (gas) -57.83
H 2 O (liquid) -68. 39
In the data (1) the measured levels of CO were generally less
than about 200 ppm and of NO were less than about 1000 ppm. For efficiency
calculations, then, both product concentrations can be neglected in Eqs. (A -5)
through (A-7). For more ready use ih’thiá study, Eqs. (A-3)through (A- 7)’
were converted to th forms
SMTD = 0 . 2099 - CO 2 J - 102] (A-9)
AFR = AK2 + ( [ Go 2 ] + [ b 2 J) SMTD (A-jo)
= AK3 + [ co 2 ] SMTD (A-il)
where
AK1 = 0.Z099(0.9406b - i.881ic + 0.5d + 4.7623e) (A-i2)
AKZ = 0 .25b - 0.5c + e (A-13)
62
-------
a
b
C
d
e
MWf
K-cal
ff’ g-mole
AK!
AK2
A K 3H *
AK3L*
Natural Gas
1.1103
4. 118
0.0427
0.0174
0. 0
18.41
-19.71
0.7980
1. 0081
1.2832
1.0528
Low Sulfur Oil
0. 5647
0. 8622
0.0034
0.0014
0. 0007
7.742
0. 1697
0.2146
0.2997
0.2514
} J 0 - H° + 6 9.3e
LI 2 Lw
- 94.38
and = the heat of formation of water (gaseous or liquid state).
For the natural gas and low sulfur oil fuels used in the
boilers (1)
(A-14)
-1.25
*
Also, the theoretical relations of CO 2 to 02 (under the assump-
tion that all carbon in the fuel goes to CO 2 for the two fue [ s are
Natural Gas
[ C0 2 l = 0.1221 - 0. 5818 [ O ] (A -15)
L Jth 2
The AK3H represents the case where the water in the combustion products
is condensed (the high heat of combustion). The AK3L represents the low
heat of combustion.
63
-------
Low Sulfur Oil
[ CO 2 ] = 0. 1614 - 0. 7688 [ Ozl (A ’16)
If the stoichiometry developed here is used, some of the heat
losses can now be calculated.
A. 2 INCOMPLETE COMBUSTION
The heat input rate usually taken as a reference in combustion
efficiency calculations is the high heat ,of combustion times the fuel flow rate.
Both the high and low heats of combustion are somewhat unrealistic combus-
tion references in that they result from simple stoichiometry and the assump-
tion that all the carbon and hydrogen in the fuel are completely oxidized to
carbon dioxide and water. Such a case is reasonably approximated in com-
bustion systemswhere there are: -
a. Plenty of excess air
b. Good uniform mixing
c. : Sufficient time at high temperature to assure a close
approach to equilibrium combustion in the initial reactions
d. Sufficiently slow cooling to assure that equilibrium is
maintained throughout the cooling cycle
In the natural gas and oil diffusion flames in a utility boiler, where the reac-
tants are simultaneously mixing, reacting and cooling, it is quite conceivable
that none of these conditions are adequately satisfied.
The question of complete combustion is particularly pertinent
with regard to the staged-combustion technique for NO control. In this tech-
nique, the initial combustion reactions take place in a fuel-rich environment.
There is not enough oxygen available in the first. stage to oxidize all the C to
CO 2 ‘and the H 2 to H 2 0, so higi i ,concentratjons of CO, free H 2 and a variety
of hydrocarbon species must result. If the combustion products in this stage
are cooled too much before the remaining air is mixed in (the’ second stage),
the composition could be frozen at the first stage concentrations. The total
heat then released would e much less than the high or low heat of combustion.
Figure 16 shows plots of the concentration of CO 2 in the flue
gases over a range of excess oxygen levels with natural gas and oil fuels. The
curves shown in the figure result from equilibrium combustion calculations
64
-------
NOX
FUEL _________ PORTS
o NATURAL 0 ALL CLOSED
GAS 0 ALL OPEN
• LOW 02/3 CLSD/OPEN
SULFUR
OIL
.
Figure i6. Experimental C0 2 /0 2 data for all firing
configurations in one boiler type.
FRACTION
BURNERS
ACTIVE
0
>-
c .J
0
C-,
15
14
13
12
11
10
9
.
.
1 2
3 4
02% DRY
5
6
65
-------
using a complex computer program. The straight lines are from Eqs. (A-15)
and (A- 16). The equilibrium combustion calculations simulate all the com-
bustion conditions necessary to release the low heat of combustion except that
the combustion products have not been cooled. The high temperature tends to
favor the more energetic products, largely through the water-gas reactions.
As a result, there are less CO 2 and H 2 0 and more CO and H 2 (and excess
at high temperatures. As the products are cooled, the CO and H 2 oxidize
further to form more CO 2 and H 2 0, approaching the concentrations calcu-
lated from stoichiometry.
Figure 16 also shows some of the measured C0 2 /0 2 data for
the tests (1). The data lie about midway between the high temperature equi-
librium combustion and the stoichiometrjc calculations. Measured values of
CO corresponding to the data are all very low. The few attempts to measure
hydrocarbons also indicated very small, almost undetectable concentrations
of this pollutant. Boiler operation was always adjusted (excess air) to assure
negligible smoke. No other chemical species were measured in the flue
gases.
Thus, it might be reasonable to assume that all the carbon in.
the fuel is either oxidized to CO 2 or ends up as soot or some other form of
carbon, with a heat of formation of approximately zero. These are the
assumptions involved in the development of Eq. (A-il).
When the actual measured CO 2 levels (expressed as mole
fractions) are used, the actual heat released and available for transfer tâ
the working fluid can be calculated from Eq. (A-il). Since the water vapor
in the flue gases is clearly not condensed before the gases leave the boiler,
the appropriate heat of formation of the water in Eq. (A-i4) (Mi ) is that
of the gaseous state. The heat loss due to incomplete combustion, however,
is independent of whether the water vapor condenses or not, and either the
high or low heat of combustion can be used. For purposes of this calculation,
the high heat of combustion is used.
If the maximum high heat of combustion (QHHCM) is defined as
that released when all the carbon is oxidized to C0 2 , then QHHCM can be cal-
culated from Eq. (A-li) using measured values of 02 and values of CO 2 con-
centrations calculated from Eqs. (A-i5) or (A-16). The actual high heat of
combustion (QHHC) can then be defined as that calculated from Eq. (A-li)
using the measured concentration of CO 2 . If these definitions and calcula-
tions are used, the heat loss due to incomplete combustion (OLIC) is
QLIC QHHCM - QHHC (A- 17)
66
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The measured levels of CO 2 cannot exceed the levels shown
by the straight lines in Figure 16 [ Eqs. (A-is) and (A-16)] for any measured
level of 02. Regardless of any possible error in the 02 measurement, the
CO 2 level certainly cannot exceed the levels defined by the zero 0 intercept
(12.21 and 16.14 percent for natural gas and oil fuels, respectively). Flue
gas analyses for CO 2 were made either by grab samples subsequently ana- .
lyzed by the chemical laboratory of the utility or by continuous sampling and
analysis in the Air Quality Control Mobile Laboratory of the utility. The
data shown in Figure 16 are only those from analysis by the chemical labora-
tory. Four data points are shown where the measured CO 2 concentrations
were larger than the maximum considered theoretically possible.
The data resulting from flue gas analyses conducted in the
mobile, on-site laboratory, however, showed about 50 percent of the mea-
sured CO 2 levels above the maximum theoretical for both gas and oil fuels.
As a result, these data were omitted from consideration in this part of the
analysis, where the measured CO 2 level is critical to an estimate of the
degree of incomplete combustion. For all other calculations depending on
the CO 2 measurement all data were used, but if the recorded CO 2 level was
above the maximum theoretical or below the equilibrium flame calculation,
then the CO 2 level was taken as the average of these two calculated values.
Most of the data shown in Figure 16, particularly with the natural gas fuel,
are near this average. Measured 2 levels had to be assumed accurate, but
this is more likely because 02 levels are an important boiler control
parameter.
The data shown in Figure 16 represent:
a. The nominal or reference configuration with all burners
active and NO ports (if any) closed
b. The reference case except with NO ports open
c. All data where one-third of the burners were operated
air-only, and NO ports were open or closed
The latter case is represented by four boilers with four of 12 burners oper-
ated air-only and two boilers operated with eight of 24 burners air-only.
The data show that there is no apparent trend for the mea-
sured CO 2 levels to either increase or decrease as a result of these two
combustion modifications. Since the heat loss due to incomplete combus-
tion [ Eqs. (A-9), (A-li), and (A-17)] is a function primarily of CO 2 . this
heat loss should also show no effect of these combustion modifications.
Calculations of the effects of NO ports and one-third of the burners out of
service (air-only) on efficiency losses resulting from incomplete combustion
for all the data of Figure 16 show changes of about ±0. 2 percent, well within
the uncertainty of the data.
67
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A. 3 UNCONDENSED WATER VAPOR
Heat losses due to water in the flue gases which leaves the
boiler system as uncondensed vapor are represented simply by the differ-
ence between the high and low heats of combustion. Note that in
Eq. (A- Il) represents the high or low heat of combustion, depending on the
value of the constant AK3 (AK3H or AK3L); this heat loss is a constant per
unit weight of fuel flow for a given fuel, independent of the values of [ co 2 }
or SMTD. The calculation used is simply -
QLH2O = 169, 800(AK3H - AK3L)/MWf (A-18)
A.4 SENSIBLE HEAT
Since the flue gases leaving the boiler stack are relatively
hot, a significant amount of the heat relea ed by èombustion leaves the boiler
system in this manner, without contributing to the-development of useful work.
This heat loss can generally be expressed by
•s = tCpfg(Ts - Tamb) (A-19)
The ambient temperature Tamb can be taken as 300 K (80°F),
but the temperature of the gases leaving the air pre-heater (and entering the
stack) was not recorded in many of the tests available to this study. Review
of such data as were available indicated that the data could reasonably be fit
with a function where the temperat4re difference (T 5 - Tamb) is expressed
as a linear function of boiler load (FLOAD)
T - T = 100 + D’(FLOAD) (A-20)
s amb
Equation (A-20) represented a good fit for ll the data (at least above 50 per-
cent o rated load) available from the Haynes boilers (1) whenthe constant D
was equal to 71, but a value of 128 was required to fit the Scattergood data.
Similarly, data are not normally available on the specific heat
of the flue gases (Cpfg). Equilibrium combustion calculations made at Aero-
space, employing the Aerospace n-element chemistry program, were used
to derive theoretical values of Cpfg. It was found that these values could be
reasonably approximated over the range of interest by expressing Cpfg as a
linear function of the overall boiler air-fuel ratio (AFR)
68
-------
C = A’ - B’(AFR) (A-Zi)
pfg
where the values of the constants A’ and B’ were found to be:
Constant Natural Gas Oil
A’ 0.409 0.374
B’ 0.00322 0.0025
Finally, the total weight flow rate of combustion products
going up the stack can be expressed as
W = W (1 + AFR) (A-2Z)
The OLSEN is defined as the sensible heat loss per pound of fuel burned
OLSEN =Q/ f (A-23)
Substitution of Eqs. (A-20) through (A-22) into Eq. (A-23) yields the expres-
sion used in this study to estimate sensible heat losses
OLSEN = (1 + AFR) * (CES - CES! * AFR)
* (100 + CES2 * FLOAD) (A-24)
where * is the FORTRAN symbol for multiply.
A.5 ELECTRICAL LOSSES
Almost nothing was available to this study to evaluate possible
variations in efficiency losses in the electric generating and control equip-
ment with operating conditions. Such losses do appear to be relatively inde-
pendent of geometry and operating conditions (except load). For this study,
simply to keep the overall efficiency values in reasonable perspective, a
constant three percent load loss was assumed for all conditions.
69
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A. 6 EFFICIENCY CORRELATIONS
After accounting for the above four efficiency losses, the
remaining efficiency represents as nearly as possible the ratio of the energy
of the steam turbine shaft (output) to the heat entering the boiler steam (input).
Expressed in heat units (Btu’s) per second, the shaft output was given by
0 out = 977. 1 * LOAD, Btu/sec (A-25)
where LOAD is in megawatts, and the numerical constant includes the three
percent electrical efficiency loss described in Section A. 5.
The heat entering the steam per unit of fuel flow was calculated
from the initial maximum heat release (QHHCM), reduced by the three heat
losses described in Sections A. 2 through A. 4. Thus the input heat was
given by
in (QHHCM - QLIC - QLHZO - QLSEN) * WFT (A-26)
where WFT is the fuel flow rate (and * is the FORTRAN symbol for multiply).
The so-called “steam cycle” efficiency studied here is the
ratio of Eqs. (A-25) and (A-26) - -
ESC out” 0 in (A-27)
In general, it would appear that this efficiency could only be significantly
affected by the steam turbine performance (involving losses such as those
due to control valve throttling and off -design expansion conditions in the
turbine) and by rather large shifts in the heat transfer profile through the
boiler. The former effects should result largely from load variations (an
operational requirement), while the latter could result from variations in the
heat release rate, which in turn might result from th use of NO ports and
the burners-out-of-service technique for NO reduction. Initial efforts to
analyze variations in the efficiency (ESC) were directed toward identifying-
and correcting for the load effect in order to examine the remaining varia-
tion for a significant effect of combustion modifications made for the pur. -
pose of NO control.,
x
70
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APPENDIX B
ANALYSIS OF A COMBUSTION-AIR FEED
SYSTEM COUPLED MODE OF INSTABILITY IN
A UTILITY BOILER
The basic assumptions and approximations necessary to this
derivation as well as schematics of the models assumed are presented in
Section 3. The details of the derivation of the analytical expression for a
combustion-air feed system coupled mode of instability in a utility boiler are
presented in this appendix. The derivation is divided into three phases:
a. Burner air flow response to perturbations in furnace pressure
at the burner exit
b. The response of furnace pressure at the burner exit to per-
turbations in the mass flow and air-fuel ratio issuing from
the burner
c. Coupling of a. and b. into a complete combustion-air feed
system coupled mode of instability
B. 1 BURNER AIR FLOW RESPONSE
Referring to Figure 2, the following equations for air flow
through a burner can be written:
Inlet Region
P -P =R.w 2 (B- I)
wb 1 ii
Air-only me rtanc e
1 L d 1 BZ
Pi 2ZA dt (-)
71
-------
1 Lbd 1b B3
p 2 3 iA g dt -)
Air - only capacitance
AbLb dP 2
= g a 2 dt (B-4)
Exit region (including flame )
P 3 - Pf = R 3 i (B-5)
To avoid the complications of nonlinear stability analysis,
Eqs. (B-i) through (B-5) must be linearized and transformed ii ito the LaPlace
domain. This is a standard procedure in stability analysis except for the
observation, first made in the ‘previous study (1), that R 3 in Eq. (B-5) is not
necessarily a constant [ as is R 1 in Eq. (B-1)], but may be a function of the
degree of reaction which is taking place within the burner. The degree of
reaction, as discussed previously (1), is in turn a function of the flow
velocity, or weight flow rate, of air through’ the burner.
As discussed in Ref. 1, the dynamic events which take place
in an active burner probably follow the sequence:
a. A small increase in furnace pressure at a burner exit causes
an initial decrease in air flow velocity through the burner.-
b. This decrease in flow velocity allows more of the reaction to
be completed within the burner (the flame moves deeper into
the burner).
c. The greater reaction within the burner increases the resist-
ance to air flow.
d. The air flow velocity decreases even further, continuing the
sequence of events.
It is not known how rapidly the flame can move in and out of the burner in
response to furnace perturbations [ i.e., the time delays involved in resis-
tance (R 3 ) changes at constant air flow rates ( ) relative to the time delays
for changes in the air flow rates at constant resistance (R 3 )]. For pur-
poses of this linear stability analysis, it is simply assumed that the burner
72
-------
exit resistance, like all other variables, can be expressed as the sum of a
constant plus a perturbation (small) variable. Thus, to linearize Eqs. (B-i)
through (B-.5), the following substitutions were used:
P 1 = P 1 + OP 1 (B-6)
P 2 = P 2 + OP 2 (B-7)
P 3 = + OP 3 (B-8)
Pf=Pf+OPf (B-9)
= + (B-b)
Wb Wb+OWb (B-il)
R 3 R 3 + 0R 3 (B-i2)
where OR 3 is taken as
/dR \
OR I . 3 16W (B- 13)
3 \dwb/ b
Substituting Eqs. (B-6) through (B-i 2) into Eqs. (B-i) through (B-5), linear-
izing and taking the LaPlace transform results in the following
- ( o 1 ) = 2R.’5v 1 (Ov 1 ) (B-14)
(o 1 ) - ( o 2 ) = 2 bg (B- 15)
73
-------
(o 2 ) - (o 3 ) - , .s(o v ,) (B-16)
(6 vb) = (6 vj) - g (B-17)
(o 3 ) _(oPf) = z 3 b(6 vb) + (ort 3 ) B-18
To simplify these equations, drop the (6) designation [ keeping in mind that
the variables are of perturbation (small) magnitudes], substitute Eq. (B-13)
into Eq. (B-18), note that
= ‘ ‘b (B-19)
and let
L = (B-ZO)
A g
c = g kb’ b (B-Z1)
P.. = ZR. (B-22)
ii i b
- 2 dR
Rf = 3Wb + “b d’ Svb (B - 23)
Then Eqs. (B-14) through (B-18) can be written
-P 1 = R.? 1j (B-24)
74
-------
P 1 - P 2 = - LS’c i 1 (B-25)
p - p =. LS’ ,b (B-26)
= - CSP 2 (B-27)
P 3 - Pf = Rü rb (B-28)
If simultaneous Eqs. (B-24) through (B-28) are solved and
R.R C+L
T = (B-29)
ii fi
1/2
T 2 =( - .LC) (B -30)
11/3
T 3 = (B-3 1)
T 4 = R. 1 C (B-32)
the linearized burner response is
/ , \ 1 +T S+T 2 S 2
- R. 1 +Rf 1 1+T 1 S+T S 2 +T S 3 -
75
-------
When there is no flame within a burner, the linearized exit resistance Rf is
taken as zero, and
T 1 = T I = (B-34)
- 1- 2 \1/3
T 3 = T 3 = ( LC ) (B-35)
and the response of a burner becomes
1* \ 1 +T S+T 2 S 2
22 33 -
‘ nf I+T 1 ,S+T 2 S +T 3 aS
To evaluate the frequency response of the burners, the substitution
S = jui (B-37)
is made, and Eq. (B-33) becomeg
fWb\ 1 [ 1 - (T w)2] + j(T w )
- - R 1 + Rf 1 [ i - (T 2 w) 2 ] + [ (Tjw) - (T 3 w) 3 ] -
To evaluate the significance of each of the Tw terms in Eq. (B-38), a 350 MW
horizontally opposed boiler was selected, represented by the following data:
Lb = 1 . 5 in (5 ft)
Ab = 0.5 m 2 (5.24 ft)
R. = 0.083 sec 2 /N - m 2 (0.0343 sec 2 /lbf - ft 2 )
76
-------
a = 479 rn/sec (1570 ft/sec)
wb = 123 N/sec (27.7 lbf/sec)(rated load)
or:
wb = 61.6 N/sec (13.85 lbf/sec)(half load)
If the maximum frequency of interest in large boilers of approximately 100 Hz
(628 sec 1 ) is considered, the maximum possible values of the (Tw) terms
in Eq. (B-38) are
(T w) = 1.34
1 max
(T w) = 1.42
2 max
(T w) = 2.70
3 max
(T w) = 0.411
4 max
Thus, all of the (Tw) terms are significant (compared to 1. 0) and none can be
neglected. Equation (B-38), however, can be written in the form
- R I+Rf 1 (B-39)
and Fb can be written either in the complex form
Fb = Re(Fb) + jIm(F ) (B-40)
or in the phase-amplitude form
j eb
Fb = Mag(F )e (B-41)
77
-------
where
F
Re(Fb) =
3
F
Im(F )=—
-b F 3
Mag(F ) =
tan( 8 b) =
F 1 = 1 + (T 1 T 4 -
/2
T T 4 )w 4
(B-42)
(B-43)
(B-44)
(B-45)
(B-46)
(B-47)
(B-)
F 2 = [ T 4 - T 1 + (T 1 T - T T 4 + T )W2 - T T W4]W
F 3 = 1 + (T - 2T )w2 +(T - 2TjT )w 4 + T w 6
At low frequencies (w — 0), F 1 and F 3 approach 1. 0 while F 2 approaches
zero. Therefore, the phase lag 0 b approaches zero, the magnitude Mag(F )
approaches 1.0, and Eq. (B-39) becomes
I ’b
kPf.
fa
1
- R 1 +Rf
(B-49)
F 2
F 1
78
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There are three types of burners for which the low frequency response is of
interest:
First, all burners when no flame (nf) is within any of the burners
(?) = - ZR.
f ianf
nf
then an air-only burner (a) when there is some combustion within the active
burners
( ) = - 2R. (B-51)
and finally, an active, fuel-plus-air (fa) burner when there is some combus-
tion within the active burners
( i) fa [ Ri +R 3 + Wf(::3)] (B52)
B.Z FURNACE PRESSURE
The basic assumptions necessary to the derivation described
in this section are presented and discussed in Section 5. 1. 2. Figure 3 in that
section shows a schematic of the separation of mass flow and combustion
driving forces in both time and space which is fundamental to this analysis.
The basic assumptions necessary to develop a simple method to account for
three-dimensional acoustics in the furnace cavity, including damping and non-
perfect reflections from the solid boundaries of the furnace cavity, are listed
in Section 5. 1. 2 as assumptions c. through f.
To reiterate briefly, these major modeling assumptions are:
a. Perturbations in mass flow and combustion rates at the exit
of a burner initially generate acoustic waves of negligible
pressure but finite velocity amplitudes (implies zero specific
acoustic impedance).
79
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b. Significant pressure perturbations are generatçd at the burner
exit only by the sum of reflected acoustic waves (finite acoustic
impedance).
c. The acoustic waves generated at the burner exit propagate
away in equal intensity plane waves in all six directions along
the three Cartesian coordinates.
d. Interactions between bui ners are neglected.
e. The small acoustic time delay between the burner exit and the
point further out in the furnace where the concentrated com-
bustion occurs is also neglected.
Assumptions a. and b. above are reasonable approximations
of the true case. If the furnace cavity were very small, such that wave travel
times throughout the cavity were very small compared to the period of the
oscillations of interest (“acoustically small”),- then the cavity pressure would
respond almost instantly to flow and combustion variations and there would
be little room for velocity oscillations (high acoustic impedance cavity). The
cavity coul4 be treated as a single lump, with pressures uniform everywhere
in the cavity. This is the case most commonly treated in the literature. On.
the other hand, if the burner flow and combustion rate variations were occur-
ring in an infinite (or acoustically very large) reservoir, it is quite reason-
able to expect that pressure amplitudes which might be developed at the
burner exit would be sufficiently small to be of little intei est in stability
analyses. ‘ -
The accuracy of assumption c. above is not easy to evaluate.
We do not know and cannot describe the complex fluid mechanical processes
which generate acoustic waves from the subsonic swirling and turbulent
flow issuing from a real burner and/or from the related combustion. The
local expansion of the furnace gases as a result of local combustion is
undoubtedly a scalar quantity, but the burner flow rate is clearly a vector
quantity of some sort which could bias the inte.nsity of acoustiQ waves propa-
gating in the.various directions. Assumption c., however, not only allows a
simple approximation ,of the coupling between the burner flow, and combustion -
rates and the furnace cavity acoustics, but also provides a simple model’ of.
three-dimensional cavity acoustics which allows for damping during wave
travel and non-perfect acoustic reflections at cavity boundaries. Considering
the accuracy with which other important parameters (such as the combustion
time delay and the acoustic damping and reflection coefficients) are known,
assumption c. is considered a sufficiently accurate description to be useful
in this analysis.
Assumptions d. and e. are simplifying approximations which
should have little effect on the final results. With these five assumptions,
the response of the furnace pressure at a burner exit to perturbations in air
flow through an active burner can be developed.
80
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The perfect gas law, written for a control volume at the burner
exit, is
Pf = (Ro)(wT) (B-53)
The total time derivative of Eq. (B-53) is
dPf dw 1 dT 1dM
Pf dt - at + f dt a (B 54)
The last two terms in Eq. (B-54) are almost totally functions of the air-fuel
ratio r during combustion. In this analysis, fuel flow rates are considered
constant; therefore, air-fuel ratio variations are due to air flow rate vari-
ations alone.
Defining
1 dT 1 dM
F(r) = = - .=- (B-55)
allows Eq. (B-54) to be written as
dP dw dw
i f_j_ s F(r) b B56)
Pf dt at dt -
Since there is a time delay between the time when the unreacted gas from the
burner enters the furnace and the average time when it reacts (the combustion
time delay the effect of air-fuel ratio variations on the furnace pressure
at time t is the result of air flow rate variations occurring at the burner exit
at an earlier time t - r . Thus, Eq. (B-56) should be written
dP dw F d*(t-T)
1 f_i s (r) b c 7
dt_• _ at at -
f S lb
81
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From continuity, the rate of change of the mass of gas stored
in the control volume is the result of the difference in the mass flow into and
out of the volume
dw
=w. - w (B-58)
dt in out
In the simple case of an acoustically small furnace cavity, the control volume,
is essentially the furnace cavity, and the only flow into the cavity is that from
the burner in Eq. (B-39)]. The only flow out of the cavity is that which
flows out of the radiant section (called here the furnace cavity), through the
back-pass and stack to atmosphere.
This flow rate can be written as in Eq. (B- i) and linearized
into the form of Eq. B-14)
2 fe = (B-59)
In Eq. (B-59), a given change in furnace pressure at the exit
Pf will act over the entire flow area, resulting in a relatively large change
in low out of the furnace. The rest of the analysis described herein concerns
a single burner and the associated flow rates through that burner wb. For
simplicity in the analysis, it will be assumed that all burners are acting in
concert and that w of Eq. (B-59) represents an equal flow out of the control
volumes at the exit of each burner. Then
‘ vt
w — (B-60)
out
Defining:
R =ZR *n (B-61)
e bptbt
the flow out of the furnace cavity can be derived from Eqs. (B-59) through
(B-61)
p
= fe (B-62)
out R
e
82
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where Rb in Eq. (B-61) can be calculated for a given boiler from data
according to
p p
R. = fe amb (B-63)
bp
wt
Substituting the burner flow and Eq. (B-62) into Eq. (B-58) and the result
into Eq. (B-57)
dP , / P \ , d*(t-,-)
1 £ ,. fet F r) b c
a = b - + dt (B-o4)
Pf W 5 \ el wth
As long as the furnace cavity is acoustically small, variations in the burner
flow and combustion rates will result immediately in cavity pressure, uni-
formly everywhere within the cavity, according to Eq. (B-64). The fe will
always be equal to Pf. If the furnace cavity is acoustically large, however,
the furnace pressure variations generated by each of the three terms on the
right side of Eq. (B-64) will not occur instantly or even at the same time.
If assumptions a., b., and c. above are followed, it is
assumed that variations in ‘b [ the first and third terms in Eq. (B-64)] gen-
erate acoustic waves, of negligible pressure but finite velocity amplitudes,
which propagate spherically and uniformly away from the source. These
spherical waves are here resolved into acoustic plane waves which propagate
away along the positive and negative directions on the three Cartesian
coordinates.
It is not necessary to write an acoustic expression for these
waves. If Eq. (B-64) is written as
(dP\ /dP\ __
pf dt)pfkdt)jpfkdt)Z -
then each of the plane waves propagating away from the source at time t can
be represented by
/dP d v (t-T
I f 1 1 b F(r) b c
6Pf\dt/l 6 w 5 dt
83
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The wave represented by Eq. (B-66) will travel through the hot gases in the
furnace cavity, reflect off some solid boundary and return to the source after
some time delay T.. During that acoustic travel it can suffer both acoustic
damping and non-p èrfect reflections. These are accounted for here by
defining the ratio of the amplitude of the returning wave to its initial ampli-
tude as a constant K., where -
- .a
K. = K .e 1 (B-67)
1 ri
The constant Krj represents a reflection coefficient and can range from zero
to one. The acoustic damping coefficient in Eq. (B-67), the in the argument
of the exponent, can range from zero to small positive values.
According to assum tionb. above, the pressure variations at
time t in Eq. (B-66) can only be the result of burner flow and combustion rate
variations which occurred at times t - T. earlier. Thus, (B-66) should be
written
• (dP = ! K \ rb(t _ T. ) + F(r)’ d vb(t - - T. ) - (B-68)-
6Pf dt 6 w 5 wm dt
and the total contribution of waves returning from all six directions is
(dPi). = [ 1 b(t - T. ) - d (t - T.)]} (B-69)
The second term in Eq. (B-65) is
Ip\ p
1 1 f) = — fe (B-70)
•Pf dt 12 wsRe
The pressure fe actually exists at the furnace exit some distance from the
burner exit. It results from the pressure which existed at the burner exit at
some time t - r /2 earlier, and the effect of this exit pressure on the flow
out of the furna& exit will only be felt at the burner exit after a further time
delay equal to Te/Z• In this case, Te represents the acoustic wave travel in
84
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one of the six directions discussed above (usually the long vertical distance
from the burner level to the top of the furnace cavity). Equation (B-70)
should be written
/dP \ Pf(t -
= — (B-71)
Pf \ Jz WsRe
Substituting Eqs. (B-69) and (B-71) into (B-65)
dPf — Kr. 1\ b(t - •i- ) F(r) d /b(t - T Pf(t - Te )
dt L 6 [ s + dt - sRe
i=i
(B-72)
Equation (B-72) is the expression for the variation of the furnace pressure
at the burner exit in an acoustically large furnace cavity. In a smaller
furnace cavity, the time delays rj and Te would decrease toward zero and
Eq. (B-72) would approach Eq. (B-70). If the furnace cavity were infinitely
large, Eq. (B-72) indicates that burner flow and combustion rates would not
generate any pressure variations at the burner exit. These are all reason-
able limits and approximations for this analysis.
Taking the LaPlace transform of Eq. (B-72) and rearranging
terms
, \ [ . . (K.eTi5)] [ l +F(r) 5 e Tc5 ]
I_ L1 = R i=1 Wfb (B-73)
W I e -TS R
e e
Pf
Equation (B-73) is the linearized furnace response, analogous to the burner
response in Eq. (B-33).
Equation (B-73) is a rather complicated function of the complex
variable S. To evaluate frequency response, the substitution of Eq. (B-37)
85
-------
can again be made. After much complex algebra, the furnace response again
can be written either in the real and imaginary form
( ) = FF - jFG (B-74)
or the phase-amplitude form
Mag(F )e 1’ (B-75)
where
Mag(F ) = (Ft. + F J) Z (B-76)
and
tan(e) FG/FF (B-77)
These are simpler forms of Eq. (B-73) except that now
FF = -(FAFc - FBFD) (B-78)
FG =..2 (FAFD + FBFC) (B-79)
FA = F + Frr [ F cos(T w) + Fc sin( .rw)] (B-80)
FB = F - Frr [ Fc cos(T w) - F sin(rw)] (B-81)
86
-------
Fc = cos(1-w) (B-82)
FD = Ff - Slfl(TW) (B-83)
FE = 1 - ZFf Sin(TW) + F (B-84)
6
F K.cos(T.cD) (B-85)
6
F = - K. sin(T.w) (B-86)
1=1
F = F(r) w (B-87)
rr
F = S e (B-88)
fe Pf
Following Eqs. (B-78) through (B-88) at very low frequencies (w 0)
F =K. R (B -89)
F iavg e
FG = 0 (B-90)
therefore
Mag(F ) = K. R (B-91)
p iavg e
and
B = 0 (B-92)
87
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B. 3 OPEN LOOP RESPONSE
The overall frequency response of the open loop, as shown in
Figure 4, can be obtained simply by multiplying the burner response by the
furnace response
= ( )( ) (B. ..93)
If the burner response is given by Eq. (B-39), with Fb in the
form given in Eq. (B-41)
j
Wb\ Mag(F )e
(-p;) = - R 11 + R (B-94)
and if the furnace response is given by Eq. (B-74)
( L) = Mag(F )e 1 ’ (B-74R)
then the overall open loop frequency response is given by
i(eb+e )
P 0 - Mag(F )Ma (F )e (B-95)
P. -- R. +R
1 ii fI
To be unstable, the following conditions must be met
Mag(F )Ma (F )> 1 (B-96)
R. +R
il fi
and
+ B = (Zm - 1)ir (i.e., 180 deg) (B-97)
where M = 0, 1, Z,
88
-------
At very low frequencies (w O), Eq. (B-95) becomes
P K. R
o_ iavg e B- 8
i ii fl
In a typical analysis of a low-frequency, feed system coupled
mode of instability in a liquid-propellant rocket engine, the following assump-
tions, different from those in this analysis, are usually made:
a. There is no flame in the injector orifice (burner) (Rf 1 = 0).
b. The orifice (burner) capacitance is negligible (C = 0).
c. The acoustic wave travel time from the region of concentrated
combustion to the combustor exit is negligible (Te 0).
d. There is no effect of mixture ratio variations on heat release
rates and subsequent pressure variations [ F(r) = 0].
e. Wave travel times from the injector face (burner exit) to all
reflecting surfaces Ti are negligible.
f. All acoustic damping and losses due to inefficient reflections
are negligible (K = i).
g. Perturbations in mass flow from an orifice (burner) do not
immediately appear as mass addition into the combustor
volume (liquids) but are delayed until the concentrated reaction
occurs, at a time T later.
C
Assumptions e., f., and g. result in the modification
6 -T.S -‘r S
1 = e C (B-99)
A further result of assumptions a. and b. is that
T 1 = (B-i00)
89
-------
and
T 2 = T 3 = T 4 = 0 (B-b!)
Finally, Eq. (B-33) becomes
( ,) - R 1 ÷ r 1 s (B 102)
As a result of assumptions c. through g., Eq. (B-73) becomes
-1• S
c
= R e — (B-103)
tw , e wR
\ b, s e
Pf
If Eqs. (B-1ÔZ) and (B-!03) are multiplied, the open loop response becomes’-
-T S-
P R c
0 - - e e (B-!04)
- R. 1 (i + L + sRe s)
Equation (B-104) represents a common solution for a low-
frequency, feed system coupled mode of instability in a liquid rocket engine,
with Only one propellant feed system involved see, for example (Ref. 5
through Ref. 7) . Thus, the solution derived for this case represents an
extension of previous analyses and yields solutions typical of those previous
analyses when appropriate simplifying assumptions are made.
90
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TECHNICAL REPORT DATA
(Please read Inwuct ions on the reverse before completing)
1. REPORT NO
EPA— 600/2 - 77-190 I
3. RECIPIENT S ACCESSION- NO.
4. TITLE AND SUBTITLE a
Effects Oi Combustion Modifications for
Ox Control on Utility Boiler Efficiency and Combustion
5. REPORT DATE
September 1977
6. PERFORMING ORGANIZATION CODE
Stability
7. AUTHOR(S)
8. PERFORMING ORGANIZATION REPORT NO.
)wen W. Dykema
—
9. PERFORMING ORGANIZATION NAME AND ADDRESS
10. PROGRAM ELEMENT NO.
The Aerospace Corporation .
Environment and Energy Conservation Division
El Segundo, California 90245
IABO14; ROAP 2IADG-089
11. CONTRACT/GRANT NO.
Grant R803283-02
12. SPONSORING AGENCY NAME AND ADDRESS
EPA, Office of Research and Development
Industrial Environmental Research Laboratory
13. TYPE OF REPORT AND PERIOD COVERED
Final; 7/75-7/76
14. SPONSORING AGENCY CODE
Research Triangle Park, NC 27711
EPAI600/l3
15. SUPPLEMENTARY NOTESIERL_RTP project officer for this report is Robert E. Hall, Mail
Drop 65, 919/541-2477.
16. The report gives results of an evaluation of the possibility that plant effic-
iency losses or combustion instability might limit NOx reduction by combustion modi-
fication. Data from natural-gas- and oil-fired boilers were used in the analyses. The
study of effects on plant efficiency of combustion modifications for NOx reduction
showed that the effects were negligible, at least within the scatter of available data.
Nearly all plant efficiency variations (losses of up to 6%) resulted from plant load
variation, which is not considered a combustion modification for NOx control. Corn-
)ustion instability, however, appeared to be a possible limitation, if not properly
understood and accounted for by hardware modifications. Fuel-rich burner operation,
in the staged combustion technique, can create an unstable air-side feed system
coupled mode of combustion instability. A method of analyzing such instability modes
was developed for use in providing stable operating conditions, even with very fuel-rich
urners.
17. KEY WORDS AND DOCUMENT ANALYSIS
a. DESCRIPTORS
b.tDENTIFIERS/OPEN ENDED TERMS
C. COSATI Field/Group
ir Pollution Boilers
ombustion Stability Utilities
Htrogen Oxides
Tatural Gas
1’uel Oil
fficiency
Air Pollution Control
Stationary Sources
Combustion Modification
Fuel-Rich Burners
13B 13B
‘ 1B
07B
‘ lB
18. DISTRIBUTION STATEMENT
Unlimited
19. SECURITY CLASS (This Reporrj
Unclassified
21. NO OF PAGES
104
20. SECURITY CLASS (This pagej
Unclassified
22 PRICE
EPA Form 2220.1 (9-73)
91
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