EPA-600/2-76-038
February 1976
Environmental Protection Technology Series
RESIDENTIAL OIL FURNACE
SYSTEM OPTIMIZATION
Phase I
Industrial Environmental 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 traditional grouping was consciously
planned to foster technology transfer and a maximum interface in related fields.
The five series are:
1. Environmental Health Effects Research
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 instrumentation, equipment, and methodology to repair or prevent
environmental degradation from point and non-point sources of pollution. This
work provides the new or improved technology required for the control and
treatment of pollution sources to meet environmental quality standards.
EPA REVIEW NOTICE
This report has been reviewed by the U.S. Environmental
Protection Agency, and approved for publication. Approval
does not signify that the contents necessarily reflect the
views and policy of the Agency, nor does mention of trade
names or commercial products constitute endorsement or
recommendation for use.
This document is available to the public through the National Technical Informa-
tion Service, Springfield, Virginia 22161.
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EPA-600/2-76-038
February 1976
RESIDENTIAL OIL FURNACE SYSTEM
OPTIMIZATIONPHASE I
by
L. P. Combs and A. S. Okuda
Rocketdyne Division
Rockwell International
6633 Canoga Avenue
Canoga Park, California 91304
Contract No. 68-02-1819
ROAP No. 21BCC-27
Program Element No. 1AB014
EPA Project Officer: 6. Blair Martin
Industrial Environmental Research Laboratory
Office of Energy, Minerals, and Industry
Research Triangle Park, NC 27711
Prepared for
U.S. ENVIRONMENTAL PROTECTION AGENCY
Office of Research and Development
Washington, DC 20460
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CONTENTS
Section I: Conclusions 1
Thermal Efficiency 1
Pollutant Emissions 3
Section II: Recommendations 5
Section III: Introduction 7
Objectives 8
Approach 8
Section IV: Analytical Investigation 11
Current Commercial Practices 11
Furnace Systems Performance Analysis 19
Section V: Experimental Investigation 55
Optimized Conventional Burner Study 55
Combustion Gas Recirculation (CGR) Burner Study 89
Flue Gas Recirculation (FGR) Burner Study 97
Section VI: Prototype System Design Investigation 109
Preliminary Designs: Areas of Commonality 109
Conceptual Design Study: Warm-Air Furnace 118
Conceptual Design Study: Hydronic Boiler 131
Section VII: References 141
Appendix A
Warm Air Oil Furnace Computer Model A-l
Appendix B
Flue Gas Compositional Analysis B-l
Appendix C
Data Tabulation: Optimum Burner Experiments C-l
Appendix D
Data Tabulation: CGR Burner Experiments D-l
Appendix E
Data Tabulation: FGR Burner Experiments E-l
Appendix F
Pre-Prototype Air-Cooled Finned Combustor Tests F-l
ill
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ILLUSTRATIONS
1. Furnace Floor Area and Volume vs Output Capacity for High-Boy
Installations 17
2. Occupied Floor Space and Volume vs Output Capacity for Commer-
cially Available, Warm-Air, Low-Boy Oil Furnaces 18
3. Calculated Steady-State Thermal Efficiency Superimposed on
Experimental Una*Spray Furnace Efficiency Curves 24
4. Calculated Effects of Flue Gas Temperature and Stoichiometric
Ratio on Oil Furnace Steady-State Thermal Efficiency 25
5. Calculated Effects of Furnace Cabinet Heat Losses to the
Surroundings and of Stoichiometric Ratio on Oil Furnace
Steady-State Thermal Efficiency 26
6. Calculated Effects of Ambient-Temperature Changes on Oil
Furnace Steady-State Thermal Efficiency 27
7. Calculated Effect of Air Humidity on Oil Furnace Steady-State
Thermal Efficiency 28
8. Calculated Effect of Emulsifying Oil With Water on Oil Furnace
Steady-State Thermal Efficiency 29
9. Calculated Cycle Thermal Efficiency Compared With Measured
Overall Heating Efficiencies of Combustion Improving Devices ... 34
10. Calculated Effect on Oil Furnace Cycle Thermal Efficiency of
Continuing to Circulate Warm Air Furnace Coolant After
Burner Cutoff 36
11. Calculated Effect of Ambient-Temperature Changes on Oil
Furnace Cycle Thermal Efficiency 37
12. Calculated Effects on Oil Furnace Cycle Thermal Efficiency of
Varying Cycle Timing at Constant Thermal Demand 39
13. WAFURN-Calculated Overall Cycle Thermal Efficiency Compared
With Measured Efficiencies 42
14. Calculated Effects on Oil Furnace Cycle Thermal Efficiency
of Varying Cycle Timing at Constant Thermal Demand 45
15. Calculated Thermal Efficiencies for Lennox OF7 Furnace for
Some Dual-Firing-Level Situations 48
iv
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16. Calculated Effects of External Flue Gas Recirculation (FGR) on
Warm-Air Oil-Fired Furnace Efficiency and Flue Gas Temperature . . 51
17. Optimum 1 ml/s (gph) Oil Burner 56
18. Experimental Combustion Chamber and Heat Exchanger
Arrangement 59
19. Heat Exchanger for Variable Combustion Chamber Configuration
Tests 62
20. Schematic of Oil Burner and Research Combustion Chamber Test
Installation 64
21. Effect of Combustion Chamber Length Upon Flue Gas Emissions as
Functions of Stoichiometric Ratio in a 0.222 m (8.75 inches)
Diameter, Air-Cooled, Tunnel-Fired Combustor 68
22. Cycle-Averaged Flue Gas Emission, 1.0 ml/s Optimum Oil
Burner in 0.279 m Diameter, Air-Cooled, Tunnel-Fired Combustor
at Various Heat Exchanger Positions 70
23. Cycle-Averaged Flue Gas Emission, 1.0 ml/s Optimum Oil Burner
in a 0.22 m Diameter, Partially Insulated, Tunnel-Fired
Combustor at a 0.40 m Heat Exchanger Position 72
24. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s (gph) Optimum Oil
Burner in a 0.175 m (6.89 inches) Diameter, Insulated, Tunnel-
Fired Combustor at Various Heat Exchanger Positions 74
25. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s Optimum Oil Burner
in a 0.222 m Diameter, Insulated, Tunnel-Fired Combustor at
Various Heat Exchanger Positions 75
26. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s (gph) Optimum Oil
Burner in a 0.222 m (8.75 inches) Diameter, Water-Cooled,
Tunnel-Fired Combustor at Various Heat Exchanger Positions ... 77
27. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s (gph) Optimum Oil
Burner in a 0.279 m (11 inches) Diameter, Air-Cooled, Side-
Fired Combustor at Various Heat Exchanger Positions 79
28. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s (gph) Optimum Oil
Burner in a 0.222 m (8.75 inches) Diameter, Air-Cooled, Side-
Fired Corabustor at Various Heat Exchanger Positions 80
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29. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s Optimum Oil Burner
in a 0.222 m (8.75 inches) Diameter, Partially Insulated, Side-
Fired Combustor at a 0.40 m (16 inches) Heat Exchanger Position . . 82
30. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s Optimum Oil Burner
in a 0.222 m Diameter, Insulated, Side-Fired Combustor at
Various Heat Exchanger Positions 83
31. Burner/Chamber Configuration for Combustion Gas Recirculation
(CGR) Tests 91
32. Photograph of Combustion Gas Recirculation Burner 92
33. Photograph of the Flue Gas Recirculation Research Burner .... 99
34. Schematic of the Residential-Size Optimized Burner With
Positive-Pressure Air Flow/Draft Loss Control Device Ill
35. Conventional and Sealed-Air Furnace Installations 115
36. Preliminary Layout Drawing of the Air-Cooled-Combustor,
Warm-Air Furnace 119
37. Layout Drawing of the Finned-Combustion, Prototype Warm-Air-
Furnace Unit With a Sealed Air System 123
38. Layout Drawing of the Horizontal-Pass, Prototype Hydronic-
Furnace With a Sealed Air System 132
VI
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TABLES
Hydronic Heat
Support Systems
Operating Mode,
1. Summary Assessment of Commercial Practices: Oil Burners
2. Summary Assessment of Commercial Practices: Fireboxes .
3. Summary Assessment of Commercial Practices: Forced Air Heat
Exchange
4. Summary Assessment of Commercial Practices:
Exchange
5. Summary Assessment of Commercial Practices:
6. Summary Assessment of Commercial Practices:
Safety, and Control
7. Optimum Burner/Chamber Matching Test Matrix
8. Summary of Results From 1 ml/s (GPH) Optimum Burner/Combustion
Chamber Matching Experiments
9. Combustion Gas Recirculation Burner Test Matrix
10. Summary of Combustion Gas Recirculation Burner, Hot Fire
Experiments
11. Composition and Properties of Fuel Oils Used in Oil Burner
Experiments
12. Summary of Finned Combustor/Heat Exchanger Operating
Conditions Heat Transfer Analysis
13
14
14
15
15
16
65
86
94
95
106
121
vii
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ABBREVIATIONS AND SYMBOLS
A Area, m
A. Area, burner combustion air inlet, m
B = Exponential decay coefficient
c = Heat capacity, constant pressure, J/kg-K
CGR = Combustion gas recirculation
D = Diameter, m
FGR - Flue gas recirculation
H,. Effective flue height, m
AH = Heat of combustion, J/kg
NOFI - National Oil Fuel Institute
Q = Heat exchange rate, J/s
3
Qf Volumetric output of fan, m /s
SR = Stoichiometric ratio, (actual air/fuel ratio)/(stoichiometric
air/fuel ratio)
T Temperature, K
t = Time, s or min
UHC = Unburned hydrocarbons
V = Volumetric flowrate, m /s
W = Mass, kg
w = Mass flowrate, kg/s
x = Mass fraction
n = Efficiency
O = Average value of a parameter
viii
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Subscripts
air
amb
b
eye
elec
emul
f
fan
fb
fg
fuel
furn
he
hum
H2OU)
H20(v)
losses
net
n.d. air
off
on
react
surr
ss
th
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
Pertains
to air supply
to ambient atmospheric conditions
to burner
to furnace firing cycle
to furnace electrical power consumption
to water emulsified in fuel oil
to fuel
to warm air furnace coolant fan
to firebox
to flue gases
to fuel oil
to furnace
to furnace heat exchanger
to humidity in combustion air
to liquid water
to water vapor
to furnace thermal losses
to lower heating value of fuel
to natural draft air flow through furnace
to burner cut-off time or off-period
to burner start-up time or on-period
to reactants (fuel and air)
to furnace surroundings
to steady state
to thermal conditions
ix
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SECTION I
CONCLUSIONS
Conclusions from the research reported herein may be separated conveniently into
those related to fuel utilization efficiency and those related to emission of air
pollutants.
THERMAL EFFICIENCY
Residential space heating equipment - warm-air furnaces and hydronic boilers -
typically have net thermal efficiencies (based on the fuel's lower heating value)
on the order of 70 to 80 percent during steady-state operation, with season aver-
ages of perhaps 60 to 75 percent. Gross efficiencies (based on the fuel's higher
heating value) are about 5 to 7 percent lower than net efficiencies. Potentially,
then, fuel utilization efficiencies might be increased by as much as 35 to 40
percent. Achieving such large gains, however, would entail very significant de-
partures from current design concepts, and manufacturing, marketing, and utiliza-
tion practices. To maximize efficiency, considerably larger heat exchangers would
be needed to cool flue gases to near room temperature and condense their combustion-
generated water vapor. A number of problems would arise immediately, e.g., inade-
quate firebox draft, need for corrosion-resistant furnace and flue construction
materials, condensate disposal, and noncompetitive initial costs. For these rea-
sons, it was concluded that the current research should be addressed to defining
efficiency gains achievable by less drastic modifications of existing commercial
technology, i.e., improvements which can be achieved in a relatively near-term
period of time and remain cost-competitive in the market place.
The greatest source of thermal inefficiency in residential space heating equip-
ment is the convection of heat up the flue. When the burner is being fired,
exhausted product gases carry off substantial sensible heat and the heat of
vaporization of combustion-generated water vapor. Flue gas sensible heat losses
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may be reduced by lowering either excess air or flue gas temperatures. Steady-
state net thermal efficiencies might thus be increased by 5 to 10 percent before
furnace operation is constrained by pollutant formation, low draft, and condensa-
tion problems. During this same time, most installations use heated, humidified
living-space air for burner combustion air and for barometric draft control air.
Although the resultant heat losses are not charged against furnace thermal effi-
ciency, fuel utilization efficiency can be raised by 10 percent or more by using a
sealed air system to bring in outdoor ambient air for these uses.
When the burner is not being fired, a natural draft flow of air through the burner,
firebox, etc., cools furnace components and continues to convect heat up the flue.
This loss can reduce net thermal efficiency by as much as 5 percent.and is a major
fraction of the transient heat losses which cause cycle-averaged efficiencies to
be lower than steady state. The lower the draft air loss is, the less sensitive
is cycle-averaged efficiency to variations of cycle duration and of fractional
burner-on time. Draft air loss can be eliminated by providing a positive shutbff
device in the combustion air supply. It can also be reduced or eliminated by
having the burner fire more nearly continuous, e.g., by using modulated flow or
high/low/off burner control. No real thermal efficiency advantage can be realized
with such control schemes, however, because the flue gas temperature varies enough
to offset the reduction or elimination of draft air heat losses.
Heat conducted to the furnace's exterior cabinet is radiated and convected to the
surroundings. Although treated as a furnace loss, this heat in some installations
may contribute directly to heating the residence and not be a true heat loss.
These furnace setting losses average about 1-1/2 to 2 percent for warm-air furn-
aces. They are higher (3 to 3-1/2 percent) for hydronic boilers because compo-
nent temperatures are more nearly constant during standby than are those in warm-
air furnaces. Reduction of these losses nominally depends upon a straightforward
economic tradeoff of fuel economy versus more expensive insulation, although manu-
facturers must also be concerned about first-cost competition.
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The electrical energy expended in operating furnaces amounts to no more than 2 to
3 percent of the fuel's net heat of combustion. Nevertheless, minimizing electric
power consumption is a worthwhile goal because of its greater cost to the homeowner
than fuel oil and because of the factor of 2-1/2 to 3 on total energy savings when
the electric generating plant's inefficiency is considered.
POLLUTANT EMISSIONS
Goals were established for flue gas pollutant emissions as follows: CO < 1.0 g/kg
fuel burned, UHC 1 0.1 g/kg fuel burned, NO < 0.5 g/kg fuel burned, and smoke <
Bacharach No. 1. It was found experimentally that these goals could be satisfied
by an optimized low-emission conventional burner fired into a properly matched
firebox. At a 1-ml/s (gph)* firing rate, the optimized firebox was found to
require: (1) A somewhat larger diameter (>0.28 m (11 inches)) than is conventional
practice, (2) uniform extraction of 20 to 25 percent of the heat through the fire-
box walls over approximately 0.5 m (20 inches) of firebox length, and (3) retention
of heat in the firebox during the burner-off or standby time. It was concluded
that the burner/firebox combination could be either water-cooled or air-cooled,
either side-fired or tunnel-fired, and designed for nominal operation at a rela-
tively-efficient 15-percent excess air (13-percent C02 in dry flue gases).
A similar experimental effort to determine firebox matching criteria for a 1-ml/s
(gph) combustion gas recirculation (CGR) burner was frustrated by an inability to
simultaneously satisfy the NO, CO, and UHC emission goals. Because of these prob-
lems and the inherent potential problems of passing moderately high-temperature
mixed gases through the combustion air fan, it was concluded that further work
with this burner concept was not warranted.
*1.00 ml/s = 0.951 gph. The 5-percent difference is neglected in this
report when referring to burner firing rates.
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Tests of a 1-ml/s (gph) flue gas recirculation (FGR) burner were more successful,
and showed good potential for satisfying the emission goals under efficient op-
erating conditions. However, actually achieving all of the emission goals prob-
ably would be contingent upon using a more complicated burner start sequence to
eliminate start-spike emissions of CO, smoke, and UHC experienced under conditions
with acceptably low NO. Again, it was concluded that the optimized conventional
burner in cooled combustion chambers is a preferred choice over the FGR burner.
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SECTION II
RECOMMENDATIONS
Recommendations derived from the Phase I research are concerned with incorporating
the results into a prototype furnace capable of satisfying the pollutant emission
and performance goals, while remaining cost-competitive, and evaluating its be-
havior experimentally. In essence, this constitutes a go-ahead recommendation for
the Phase II activities as planned.
5/6
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SECTION III
INTRODUCTION
The Environmental Protection Agency has sponsored studies over the past few years
to document the emission of air pollutants from existing residential and commer-
cial oil-fired space heating equipment (Ref. 1 and 2). Concurrently, the EPA has
also supported applied research programs to determine the effects of design and
operating parameters on exhaust gas emission levels and to devise strategies for
minimizing pollutant emissions (e.g., Ref. 3 through 5). These and related studies
have shown that substantial reductions in total emissions from furnaces can be
realized by applying combustion control technology, such as flue gas recirculation,
two-stage combustion, and advanced burner designs. As. a portion of those EPA
activities, Rocketdyne performed an intensive investigation of residential and
commercial oil burners (Ref. 5) which led to criteria for optimizing conventional
burner designs with respect to emissions. The optimized burners were shown to be
capable of reducing NO emissions to below 1 g NO/kg fuel, in contrast to typical
Jv
manufactured residential burners which operate at about 1-1/2 to 3 g NO/kg fuel or
higher. During that same program, some variations in the combustion chamber con-
struction (adiabatic refractory walls versus metal heat-sink or water-cooled walls)
and design (relative orientation of the burner and chamber axes and relative burner
and chamber diameters) were studied. The results clearly demonstrated that the
total emissions are sensitive to the design of each component comprising a resi-
dential heating system and to design interactions among the components. It became
obvious, therefore, that minimum emissions could be achieved only by systematically
optimizing the burner in conjunction with the combustion chamber and furnace as
well as operating mode. The research described and discussed in this report was
undertaken to investigate such optimization and to delineate furnace design re-
quirements for commercializing the optimum furnace technology.
In view of uncertainties in fossil fuel supplies and prices and of the national
emphasis on striving toward energy independence, an optimized furnace should, in
addition to providing low levels of pollutant emissions, also provide some sub-
stantial increase in overall thermal efficiency. Further, for the developed
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technology to be adopted commercially, the optimized design must be cost competi-
tive. In this regard, significantly increased fuel economy will be important in
allowing consumer recovery, within a reasonable number of heating seasons, of any
increased capital costs which might result from modifications required for system
optimization. Establishing the technological requirements for optimizing residen-
tial heat systems is important, therefore, because it addresses both improvement
of air quality and better utilization of our dwindling energy resources.
OBJECTIVES
The principal objective of this research program is to establish the technology
required for engineering optimization of residential heating furnace systems com-
bining minimum pollutant emissions with maximum system efficiency. Primary em-
phasis will be given to systems fueled with No. 2 distillate fuel oil; however,
the impact on performance and emissions of substituting gaseous fuels for oil also
will be assessed. General overall goals are to reduce emissions of oxides of
nitrogen to less than 0.5 g NO/kg fuel burned, while maintaining minimum emissions
of CO, UHC, and smoke, and to increase overall season-averaged furnace energy ef-
ficiencies 10 percent or more above those achieved by current conventional systems.
APPROACH
The research program undertaken to accomplish these objectives is divided into
two phases, each of which is further subdivided into several tasks. The first
phase, Systems Analysis and Experimental Research, has been performed in the first
year of the program and is the subject of this interim report. It comprised tasks
as follows:
1. Systems Analysis, in which current commercial designs were reviewed
and analyzed to identify potential areas for improvement
2. Conventional Burner/Combustor Matching Experiments, in which the
1.0 ml/s (gph) optimum burner was tested in research combustors
having a variety of sizes, configurations, and constructions to
further the optimization
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3. Experimental Definition of Requirements for an Integrated Furnace
System, in which burner/combustor matching tests were conducted with
burners employing forced combustion gas recirculation (CGR) and flue
gas recirculation (FGR)
4. Data Evaulation and Systems Analysis, in which the results of prior
tasks were synthesized to support preliminary conceptual designs for
two prototype, low-emission, improved-efficiency furnace units. One
preliminary design is for a warm-air furnace and the other is for a
hydronic boiler. Each design was submitted to a leading manufacturer
of its particular type of residential heating system for review and
assessment by knowledgeable engineering personnel. Cognizance was
taken of their comments and recommendations in selecting design fea-
tures for candidate prototype residential furnace systems. These were
then discussed with the EPA Project Officer, and the warm-air system
was selected as the more promising of the two; therefore a prototype
warm-air unit will be built for further study in the continuing work.
The second phase of the program will involve the construction and experimental
evaluation of a prototype optimized warm-air furnace unit. The data and experi-
ences gained will be applied to the final design of a cost-competitive, commer-
cially producible warm-air furnace which embodies the derived low-emission, im-
proved-efficiency technology. This phase will be documented in a final report
upon completion of the research.
9/10
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SECTION IV
ANALYTICAL INVESTIGATION
An analytical investigation of conventional residential oil heating equipment was
conducted to identify areas where improvements might be made and to quantify the
potential gains. Although both pollutant emission reduction and thermal effi-
ciency enhancement were of interest, the emissions are more amenable to experi-
mental than to analytical characterization, so efficiency aspects were emphasized
in this effort.
As a preface to numerical analysis, a brief survey was conducted of current and
alternate practices in the design of components for residential heating systems.
Thereafter, two thermal analysis models for furnaces were formulated and pro-
grammed for computer solution. The first model (FURNAC) treated furnace design
and operational aspects strictly parametrically (and largely empirically) to
screen and isolate those aspects requiring more vigorous analytical examination.
The second model (WAFURN) accounted in a simplified manner for transient firebox
and heat exchanger heating and cooling phenomena in a rather generalized warm-air
furnace.
CURRENT COMMERCIAL PRACTICES
A brief informal survey was conducted to gather and systematize information re-
garding predominant current commercial practices, alternate approaches, trends
foreseen for the near future, and areas for potential improvement in residential
oil furnace systems. Sources of information included:
1. Previous work performed at Rocketdyne
a. EPA Contract 68-02-0017 (Ref. 5)
b. NOFI contract for development of Una*Spray Integrated Furnace
2. EPA-R2-73-084a, "Field Investigation of Emissions from Combustion Equip-
ment for Space Heating" (Battelle)
11
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3. Manufacturers' brochures on currently available burners and furnaces
in the 0.5- to 1.5-ml/s (1/2 to 1-1/2 gph) size range
4. Published results from NAPCA/EPA in-house studies
5. Applicable codes and standards
6. Proceedings of API and NOFI Workshops
7. Personal contacts with furnace manufacturers, NOFI, Battelle, etc.
Qualitative information obtained was systematized into a number of summary assess-
ment tables, categorized by subsystems, components, and functions. These assess-
ments are given in Tables 1 through 6, which are essentially self-explanatory.
Current catalogs describing warm-air and hydronic residential oil furnaces were
requested from approximately 30 U.S. manufacturers. They were studied to dis-
cern design trends and practices and to collate data on burner types, furnace
weights and volumes, floor space requirements, etc., as functions of system type
and structural configuration (high-boy, low-boy, horizontal). While there are
substantial differences among the various manufacturers' products, there are
also some distinct similarities. For example, the floor areas and volumes of
high-boy oil furnaces increase approximately linearly with firing rate, and ex-
hibit only about a ±25 percent scatter about mean values (Fig. 1). The moderately
small scatter probably reflects the constraint imposed by closet-type installa-
tion. Low-boy furnaces are more often installed in basements, and are not so
spatially confined as are high-boys; similar floor area and volume plots for
low-boys (Fig. 2) show a substantially greater scatter (about ±35 percent).
Comparing current data with results from earlier surveys revealed that oil furnaces
have experienced little change since the Rocketdyne survey conducted for NOFI
(Ref. 6) and the Battelle survey for the U.S. EPA (Ref. 4). Since those surveys
were both broader and more detailed than the informal investigation conducted
here, the interested reader is referred to those documents.
12
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TABLE 1. SUMMARY ASSESSMENT OF COMMERCIAL
PRACTICES: OIL BURNERS
SUBSYSTEM OR
COMPONENT
Burner head
Oil no?:!**
1 gn \ t ft r
Fuel pump
Conbust ion air
fan
Motor
Fuel cut-off
PRr.DOMINANT
CURRENT
PRACTICES
Conventional head
Snail peripheral swirl
Sonc choke
Plane retention head
Low-v:l turbulent-
wake devices
Moderate-high swirl
Shell head
Inner nixing
ch.in.ltcr (Shell)
'6 CPU
ALTERNATE
CURRENT
PRACTICES
Low-press air-atomizing
Rotary atomizing
Pre-vaporizing burners
Old techniques
Avg ages -15-18 yrs
Usage declining
As above
Interrupted
spark
500 psig output
press, for
larger burners
Squirrel cage at
1725 RPH
Radial-flow fans
Higher output press.
Rotary burnera
Solenoid valve
Centrifugal clutch
Combine cut-off with
press, regulator In
/ual ptaap
FUTURE DIRECTIONS:
DESIGN TRFNDS
Conventional head
still numerically
dominant
Increasing usage of
flame-retention heads
Adaptable to a
variety of furnaces
Smaller dia blast
* tubes
Lower smoke
Decreasing ns.ige of
shell heads
Produce finer sprays
Higher oil press.
Multiple noztles
Smaller oriflcei
Lower power
consumption
Solid-state devices
Increasing uiage of
100 psig pump*
5450 RPM
Smaller fan
Steadier flow
Decreasing applica-
tion to small
burners
POTr.NTIAl ISTTROVE^IENT
ARFA
Lower NOj
eclssioni
Retention-head
designs for
lower NOX
Shell head
designs for
lower NOX
Match fine spray
distribution to
coeibustlon ni r
Reduce power
consumption
Rtduce power
consumption
Reduce fan noise
MT1 MODS TO USE
head
Higher swirl
Greater choke
Further opiini:e
burner/ chirsber
I up rove stabl I ity
Non-adiabatic combustion
techniques
Plane cooling
Design favorable for
pseudo-staged-cocbustion
Needs hiat removal
from flame ahead of
by-pass air
Interrupted spark
Smaller, pul*c-duty
transformer
Piezoeltctrlc/soltnoid
capacltlvt discharge
Reduce by-pass/flring
rate ratio
Use a smaller punp
Multi-speed drives for
throttling applications
CO'.STPAINTS
Cost.
plugging
Codes fi
Sl'-i*
Cost
Cost.
power
Cost,
ratchlng
itoich-
ratio
Cost.
safetv
13
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TABLE 2. SUMMARY ASSESSMENT OF COMMERCIAL
PRACTICES: FIREBOXES
SUBSYSTEM OR
cc:iro-.rNT
Construction
materials
Firebox/burner
orientation
PREDOMINANT
CURRI.NT
pRAaias
jg. t circu y in
Volume matched to
nominal firing rate
-1.8 x 10° J/S-mJ
(1.74xl05 Btu/hr-ft3)
All dinensions { 0.2 m
(8-in.)
Refractories
Uncoolcd. adiabatic
Firebrick
High hrat capacity
Slow thermal resp.
Insulating brick
Low heat capacity
Fo*t thermal resp.
Side-fired (Perpendicular port
Vertical firebox
ALTERNATE
CURRENT
PRACTICES
Ceramic fiber
Similar to
insulating brick
Shape flexibility
Steel
Temp limitations
Rad'n or convection
cooling
Tunnel-fired (Coaxial)
Horizontal axes
FUTURE DIRECTIONS:
DESIGN TRENDS
cyl. combustors
Increasing use of
shorter chambers
Increasing uses of
ceramic fiber I steel
Prefabricate
Design flexibility
Lower weight
Static
K
AREA
Chamber/burner
matching
Develop vertical
coaxial
Minimize floor
space
rTEHTIAL IMPROVEMENT
METHODS TO USE
Vertical burner
axis
CONSTRAINTS
Smoke
prod'n
TABLE 3. SUMMARY ASSESSMENT OF COMMERCIAL
PRACTICES: FORCED AIR HEAT EXCHANGE
SUBSYSTEM OR
COMPONENT
Air supply
Air blower
Heat exchanger
Distribution
PRELCMINANT
CURRENT
PRACTICES
Return duct froa
hcafd r.poc*-
Low velocity
Low iP
Uninsulated
riberglAss filter
SO in.2/gph
Centrifugal squirrel-
cage fan
7 or 3 speed
1200-1400 fm/gph
LOK pressure
(1-2" HjO)
Formed, welded carbon
steel plate
10-12 ffVlOO.OOO Btu/hr
~SO% thermal efflc.
*»iOO F gas temp . discharge
Insulated, galvanized
sheet steel ducts
Flov. control dampers
ALTERNATE
CURRENT
PRACTICES
Higher capacity if
unit Includes
air conditioning
Fiber duets
FUTURE DIRECTIONS:
DESIGN TRENDS
Static
Static
AREA
None
Reduce noise
More efficient
rotors
Compact design
concepts
Higher efflc.
Integral
humidifier
POTENTIAL IMPROVEMENT
METHODS TO USE
CONSTRAINTS
Cost
Coat,
cleaning
14
-------
TABLE 4. SUMMARY ASSESSMENT OF COMMERCIAL
PRACTICES: HYDRONIC HEAT EXCHANGE
SUBSYSTEM OR
COMPOS! '.'i
Systpci type
Heat exchanger
Pimp
PREDOMINANT
CURRENT
PRACTICES
Closed, forced circulation
Low-tenp. < 250 F
working press t 30 psig
Low pimping heads
Low water temp, changes
(~lllt) (-20*7)
Catt-iron boiler type
ALTERNATE
CURRENT
PRACTICES
Gravity ci rcula-
tion tialcr
circuit
Steam heating
system
Formed, welded carbon
steel plat* « tubas
ASME find-boiler coda
Fire-tub* or water tuba
Pinnad watar tuba tvpa
Positive displacement
FUTURE DIRECTIONS:
DFSIGN TREWS
systems supplanted
by forced-circ. water
Dual -temperature
water systems
Hot in winter
Cold in summer
Higher water AT's
Better control
Smaller equipment
Lower flows, costs
POTENTIAL IMPROVEMENT
AREA
METHODS TO USE
CONSTRAINTS
TABLE 5. SUMMARY ASSESSMENT OF COMMERCIAL
PRACTICES: SUPPORT SYSTEMS
SUBS'SfTM OR
COMPONFNT
system
Atom) -.it inn
Filters
Furnace
Furl Nn:;le
Fuel line
Burner air
Flue System
PREDOMINANT
CURRENT
PRACTICES
Single pipe system
Zcro-pre*s
« gravity head
10- SO psi|
Glass wool
Sintered netal
In-line cur
(metal or paper)
Nonr
Draft (neg. press.)
ALTERNATE
CURRENT
PRACTICES
Two-pipe system
(large by-pass)
Self-atoml ring
Screen or
permanent
cleonable filter
Fine mesh screen
No filter
Sealed, pressurlied
FUTURE DIRECTIONS
DESIGN TRENDS
Single pipe
Self -atomizing
Glass wool
Sintered metal
Draft
POTENTIAL IMPROVEMENT
AREA
Small, low cost
air pump
O^Jlet air pump
Draft system
requires low
AP system
METHODS TO USE
Draft - no leak
hazard
CONSTRAINTS
Fumace-to-ol 1
distinct
safety code?
Coat (fob.)
15
-------
TABLE 6. SUMMARY ASSESSMENT OF COMMERCIAL
PRACTICES: OPERATING MODE, SAFETY, AND CONTROL
suBsvvrrH OR
COMPONENT
Operating node
Safety/Control
Ignition
FU.c
detect Ion
Electrlcnl
fill-lift
Tojlclty
Operating
ode
pRr.DOMiNANT
CURRENT
PRACTICES
On-Off
Continuous on
CaJ-vilfide cell
No vonltorlng
Fire box pre-is
6 GPU - add fuel cut-off
>10 CPH - add lo-flre
start
- Add pre-lgn
purge
FUTURE DIRECTIONS:
DESIGN TRENDS
Detect/shut -down
tlMe varies with
firing rate
Solid atate
electronic
control circuit
POTENTIAl IMPROVIDENT
AREA
Throttling
Reduce site 1
cost of
circuitry (IC's)
METHODS TO USE
Lo-fire continuous.
Hi-fire On-Off
or
Modulated flov
COXSTR,>P:TS
Cost
(control
circuit e
equip- wear)
U.L.
U.L.
U.I.
Standardize
firing
cycle
16
-------
50
40
ro
- 30
20
10
50,000
8
nS
Q)
i.
2 4
50,000
A 011 Furnace
Gas Furnace
V Electric
100,000
150,000
Figure 1
100,000
Output Capacity, Btuh
Furnace Floor Area and Volume vs Output
Capacity for High-Boy Installations
150,000
17
-------
50,000
Furnace Output Capacity, Btu/hr
Figure 2. Occupied Floor Space and Volume vs Output Capacity
for Commercially Available, Warm-Air, Low-Boy Oil
Furnaces
18
-------
In the process of examining oil furnace manufacturers' specification sheets,
three companies were selected as being approximately representative of the in-
dustry, and arrangements were made for technical visits to their plants. Two of
the three companies make warm-air furnaces; the other builds hydronic boilers.
One-day visits were made to each of the three plants by Rocketdyne personnel.
The principal objective of the visits was to engage in informal technical dis-
cussions to enhance our understanding of current design practices, how they are
constrained by fabrication, installation, and cost considerations, and trends
foreseen for the near future.
FURNACE SYSTEMS PERFORMANCE ANALYSIS
Numerical analyses of furnaces' thermal performance was undertaken to help quanti-
fy the importance of various design and operating parameters on overall effi-
ciency. Two separate computer models were developed to perform the numerical
analysis. Consider a usual definition of thermal efficiency:
useful heat delivered f.,
nth ~ heat input l J
The two models differed in their methods of calculating the "useful heat delivered'
in the numerator of Eq. 1. The first, and simpler, model attempted to quantify
the known thermal losses and subtracted them from the heat input to obtain an
estimate of the useful heat output. This model was used, principally, to examine
the potential and relative impacts on thermal efficiency of individually varying
a number of parameters. It was also instrumental in determining which parameters
should be modeled accurately, which could be treated with simplified approxima-
tions and which could be omitted in formulating the second furnace thermal analy-
sis model.
The second model's approach to obtaining a value for the useful heat delivered
was to solve for the heat transferred to the furnace cooling medium. With this
approach, loss quantification is necessary only for those losses which interact
with the heat transmission rates (and, secondarily, so that they can be under-
stood well enough to design ways for reducing them).
19
-------
Thermal Loss Deduction Method
The "heat input" denominator of Eq. 1 should include all of the energy supplied
to the oil furnace. This is sometimes assumed to consist solely of the fuel
oil's heat of combustion. While that is unquestionably the largest energy source,
there are other sources, e.g., the sensible heat (which may be negative) asso-
ciated with reactant supply at nonstandard conditions, and electricity to drive
the fuel pump and air blowers and to power the ignition circuit and basic furnace
controls. For simplicity, the energy inputs may be categorized as reactant
thermal energy, C^^, and electric power, Qelec-
Heat transferred to the furnace cooling medium constitutes the useful heat output
and is designated as Q, . Its value may be calculated by subtracting thermal
losses from the total energy inputs:
^he ~ rreact + ^elec ^losses * *
The last term must include all energy losses from the furnace, viz, the sensible
heat of combustion gases convected up the flue, the latent heat of water vapor
in the flue gases formed by fuel combustion, fresh air induced into the flue for
furnace barometric pressure control, and all heat dissipated from the furnace
to the surroundings by radiation and convection. Several assumptions were made
concerning the losses in formulating a model for computer solution. First, it
was assumed that none of the water formed by fuel combustion would condense within
the flue. As a result, the input reactant thermal energy was based upon the fuel's
lower or net heating value, (AH ) obtained by adjusting the measured value at
C TIG t
21 C (70 F) for water emulsified in the fuel, humidity in the air, and reactant
supply temperature deviations from that reference temperature:
(AH ) = (AH )2* C - (21-T- ,) (1-x ,) c - . + x . c _,.. -
v c net * c'net v fuel' I ^ emul' p,fuel emul p,H20(Jl)J
(21-T . ) (1-x, ) c . + x, c -r J (3)
v air' [v hum7 p,air Tium p,H20(V)J v '
20
-------
Second, it was assumed that the heat dissipated from the outside of the furnace
is completely lost, even though it does actually contribute to satisfying the
residential demand heat load in many installations. Similarly, the furnace's
electrical power consumption was assumed to be entirely lost, even though much
of it is deposited in the reactants and combustion gases as kinetic and thermal
energy, which may be recovered in the heat exchanger, and the remainder is dis-
sipated in the furnace surroundings. With these assumptions, Q can be
dropped from Eq. 2 and the remaining losses categorized as flue gas losses, Q- ,
and losses to the surroundings, Q . Both of these are time-varying quantities
«
which take place over an entire firing cycle. The terms Q, and Q ..in Eq. 2,
ne x*eact
however, may be applicable only during some discreet portion of a cycle, such as
the burner firing time, At . To perform energy balances and efficiency calcula-
tions, mean effective rate terms may be defined by
(4)
cycle
with all of the various heat rates normalized to the burner firing time. Then
Eq. 2 becomes
The overall cycle-averaged furnace thermal efficiency is
nfurn = ? C6)
T
In this form, it is possible to separate the electrical power consumption from
the fuel usage by defining a cycle-averaged reactant thermal efficiency
s =AL
react
i i
eact
21
-------
Combining Eq. 6 and 7 gives
Qreact
turn react
_
A + A
Teact elec
from which an approach was derived first for studying reactant thermal efficiency
and, subsequently, for those situations still of interest, studying overall effi-
ciency by assessing their electrical power requirements.
The foregoing concepts were formulated as a system of equations requiring input
values for a number of reactant and furnace .operational parameters, and pro-
grammed for solution on a Honeywell Timeshare computer. Entitled FURNAC, the
program is designed to calculate n t and r\f iteratively for several values
of stoichiometric ratio* (SR) .
Although the FURNAC program is not restricted basically to a particular cooling
medium, it has been used primarily to evaluate warm-air furnace behavior. Except
where specifically denoted as applying to a hydronic unit, the following discus-
sion concerns its application to warm-air systems.
Steady-State Operation. Quantitative analysis of cyclical heat loss effects re-
* *
quires that 0 _ , 0 , and 0 , be defined as functions of time throughout the
n xfg xsurr xelec &
cycle. Before attempting that, however, they can be treated as simple numerical
parameters to gain some insight to their relative importance or, alternatively,
to examine steady- state operation. This approach was used in running a number
of exploratory cases to aid in selecting appropriate nominal parametric values
for subsequent use in sensitivity analyses. Computed thermal efficiencies were
compared with some available steady-state furnace efficiency data as indicated
*The definition of stoichiometric ratio permits ready identification of the ex-
cess air level. For example, SR = 1.50 is equivalent to 150 percent of stoichi-
ometric air, which is 50 percent excess air. The relationships between SR and
flue gas concentrations of C02 and 02 are illustrated for No. 2 fuel oil in
Fig. B-2, page B-ll.
22
-------
in Fig. 3. Thereupon, the following nominal values were chosen for steady-state
input parameters:
-4
wfuel = 0.885 x 10 kg/s (corresponds to 1.00 gph dry oil)
xemul = °-0025 k8 H20(£)/kg oil (tank moisture)
x, = 30 percent of saturation, kg H20(v)/kg dry air
T- . T . = 10 C (50 F)
fuel air v '
T£ = 325 C (617 F)
Qsurr = 1580 J/s (5400 Btu/hr)
Qelec = 500 J/s (watts)
1 <_ SR £ 3
The FURNAC computer program was structured to compute points across the stoichio-
metric ratio range for each selected set of parametric values. Computed overall
steady-state thermal efficiencies from computer runs using the nominal values and
using one-parameter deviations from nominal are plotted in Fig. 4 through 8.
Figure 4 shows that flue gas exhaust temperature has a strong effect on achiev-
able furnace efficiency. The effect is accentuated by increasing the stoichio-
metric ratio because, if flue gas temperature remains constant, increasing the
excess air simply increases the sensible heat swept up the flue.
The effects on thermal efficiency of varying furnace cabinet heat losses (other
than flue losses) are shown in Fig. 5. Efficiency decreases linearly with in-
creasing losses; the nominal value of 1580 J/s (1.5 Btu/sec) corresponds to 4.2
percent of No. 2 fuel oil's lower hi
Ibm) at the designated firing rate.
percent of No. 2 fuel oil's lower heating value of 4.273 x 10 J/kg (18,370 Btu/
Ambient-temperature variations were found to exert very little influence on fur-
nace thermal efficiency if other parameters remain fixed (Fig. 6). (This result
is seen later not to apply to cyclical furnace operations.)
23 --
-------
CALCULATED BY FURNAC
100
-4->
C
0)
o
fc 90.
0.
A
O
-------
100
-------
100
VQll = 1.065 ml/s (1 gph)
Tamb = 10 C
Tfg = 325 C (617 F)
0.25% H20 in Oil
30% Humidity in Air
90
C
o>
u
(U
d.
Q)
80
70
60
50 -
40
1.0
2.0
Stoich. Ratio
3.0
Figure 5. Calculated Effects of Furnace Cabinet Heat Losses
to the Surroundings and of Stoichiometric Ratio
on Oil Furnace Steady-State Thermal Efficiency
26
-------
QL
0}
i.
-------
TOO,
90
c:
O)
o
S-
-------
100,
V , = Variable, such that
.emu I *
Wdry oil = 8'87 x 10
Tf = 325 C (617 F)
Tamb ' 10 C
Qsurr = 1580 J/s (1.5 Btu/s)
30% Humidity in Air
90 _
0)
o
J_
O)
Q.
OJ
o
LU
cO
-------
Data plotted in Fig. 7 show that variations in air humidity have an almost negli-
gible effect on furnace thermal efficiency when the ambient temperature is 10 C
(50 F) . An even less significant effect would be expected at lower ambient tem-
peratures, because absolute saturation humidity falls sharply with decreasing
ambient-air temperature.
Emulsifying fuel oil with increasing quantities of water is seen to degrade
overall steady-state efficiency by about 1 percent An per 10 percent water
increase, Fig. 8.
At a nominal oil flowrate of 1 ml/s (gph) , the oil heat production rate is 38,000
J/s (129,700 Btu/hr) . The relationship between overall efficiency and reactant
efficiency (Eq. 8) becomes
- 38,000 -
furn react + *
' xelec
which for various values of Q , , gives
0 , (watts) = 0; 250; 500; 750; 1000
elec
= 1'000; °'993; °'987; °'981; °'974
Typical steady-state power consumption for warm-air furnaces in this size range
is on the order of 880 W (500 W blower motor, 250 W burner motor, 100 W ignition
transformer, and 30 W control circuit). Thus, the overall steady-state efficiency
is typically on the order of 97 to 98 percent of reactant thermal efficiency.
This result immediately suggests that efforts to reduce power consumption can
yield only relatively small gains in overall system efficiency. However, because
electrical power costs are substantially higher than fuel oil prices (for an
equivalent energy level), a 1-percent increase in system efficiency achieved by
reducing power consumption may represent a significant decrease in annual oper-
ating cost. Therefore, continuing attention should be given to finding reasonable
ways of reducing power consumption consistent with safety, reliability, and ini-
tial cost considerations.
30
-------
Cyclical Operation. There are two types of flue gas heat losses during cyclic
furnace operation, one associated with the combustion gases during the burner
firing time and another associated with a natural draft flow of air through the
burner, firebox, etc., during the standby period when the burner is off. These
two components were treated separately in FURNAC. The steady-state calculation
of 6~ described above was applied directly during At to evaluate heat con-
vected up the flue with the combustion gases. Natural draft heat losses during
At ,-c were evaluated as follows.
orr
All of the hot furnace components were lumped into one mass having a given mean
heat capacity and initial mean temperature'at the burner cutoff time. This mass
was cooled by transfer of heat to the furnace coolant air flow, as long as it
continued, and to the natural draft air flow. The latter flowrate was calcu-
lated as a function of the mean furnace temperature, ambient outdoor temperature,
flue height, and effective combustion air inlet area* of the burner's adjustable
air louvers:
1/2
/U£..J. rw «...,"" - ' ~ ' ~"
D stoich
n.d.air
amb
r.- - . -i
fb amb
T j. T
L fb 'amb J
wherein an effective area flow coefficient of 0.60 was included and it was assumed
that this air flow has negligible effect on the flue's draft effect. The draft
air flow was assumed to be heated to the mean temperature of the lumped "firebox"
mass.
Rather than try to calculate the external losses to the surroundings, they were
represented heuristically as a time-varying fraction of a steady-state loss rate
(which represents the maximum loss rate during a cycle). The assumed behavior
of loss rate versus time is sketched below for one cycle of furnace operation.
*The read-in value of A^, the effective combustion inlet area, corresponded to a
burner air setting yielding SR = 1. That value was multiplied by SR to account
approximately for the changing air louvers' setting as stoichiometric ratio was
increased.
31
-------
The burner is started at t , to start this cycle, turned off at t _,., and re-
started at t - to end the cycle. The warm-air (furnace coolant) fan is assumed
to run for a few minutes after burner cutoff, and is stopped at t~ . As a re-
sult of furnace cooling during the standby period, the losses will be at their
lowest value at the end of the cycle (and the beginning of the next cycle). In
the computer program, they were taken to increase linearly from that low magni-
tude to (0 ) after 1/3 of At , to be equal to 0 for the remaining
xsurr,ss' on n xsurr,ss 6
2/3 of At and to drop linearly over the time, t ,-e < t < t-
on v J > Off < _ fan-
Qsurr(t)
*surr,ss
'
Time
eye
The level to which the losses fall during this latter time period was determined
by the cooling effect of that continuing air flow. The heat content of the fur-
nace's firebox and heat exchanger was estimated at tQff and again at t£an; the
fractional value of the loss at time t,. was assumed to be proportional to the
ratio of those two estimated furnace heat contents.
32
-------
During the period between t_ and t , a Gaussian decay of the heat losses was
assumed:
(Here, the value of B controls the rate of decay of the furnace heat losses. B =
0.004 gives 0 (t )/Q (t_ ) = 0.407 for t - t_ =15 minutes, and gives
6 xsurr^ on' xsurrv fan on2 fan 5
0.202 for 20 minutes.) Total furnace setting heat loss is the integral over the
cycle time of the time-varying losses.
Nominal values of the several adjustable parameters in the computer model were
selected by running exploratory cases and comparing calculated efficiencies with
cycle efficiency data obtained by Howekamp and Hooper (Ref . 7) . A graph of their
results is reproduced from Ref. 2 in Fig. 9. Those data were obtained by cyclical
testing (10 minutes on, 20 minutes off) of a reference ABC burner fired in a
Williamson Low- Boy furnace with several alternate burner heads. Superimposed on
their data is a bold-face dashed line calculated by running the FURNAC computer
model with the following nominal parametric values:
Vf
Tamb
Tf
x,
hum
At
Qn
At
off
sun? y ss
1.0 ml/s (gph)
"7 c (20 F)
325 C (617 F)
0
0
10 minutes
20 minutes
527 J/s (0.5 Btu/sec)
ft,
_
fan
53.1 kg (117 pounds)
= 50° J/kg~K C°-12 Btu/lb-R)
* 35S C
Hf
Qfan
B
= t .,. + 5 minutes
off
= 1.12 x 10"3 m2(1.2xlO"2ft2)
= 6.10 m (20 feet)
= 0.472 m /s (1000 cfm)
= 0.004
This set of values was neither unique nor particularly thought to be the "best"
set for matching the Howekamp and Hooper data, but represented a reasonable com-
promise among the efficiency level and the slope with stoichiometric ratio as
33
-------
u
o>
100
90.
80.
70.
60
50.
OJ
.c
tu 401
Calculated Using Nominal Values
for all Parameters
Device
Standard ABC
Monarch Combustion Head
Delavan Flamecone
Shell Head
Gulf Econojet
Union (Pure) Flame Retention Head
1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 2.
Stoich. Air/Fuel Ratio
Figure 9. Calculated Cycle Thermal Efficiency Compared with Measured
Overall Heating Efficiencies of Combustion Improving Devices
(Ref. 7)
34
-------
influenced by flue gas temperature, steady-state heat loss level, warm-air cool-
ing effect, and natural draft losses. There are probably significant deviations
from reality in the assumption of constant flue gas temperature, in the use of a
rather small value for Q and in the calculation of draft air heat losses.
o urrs o
In spite of such shortcomings, it is highly instructive to examine the variation
of cycle thermal efficiency with variations of certain parameters.
The nominal value of 0 corresponded to about 1.4 percent of the oil heat-
surr,s s r r
ing value. For the nominal 10 minutes on/20 minutes off cycle, losses to the
surroundings rose to about 2.3 percent of the oil burned, essentially irrespec-
tive of the stoichiometric ratio. Calculated draft air heat losses, however,
were a strong function of stoichiometric ratio, increasing from 2.9 percent at
R . , = 1 to 6.9 percent at R . , = 3.
stoich r stoich
Continuing to circulate the warm air furnace coolant for several minutes after
burner cutoff can contribute significantly to reducing the cycle losses in
thermal efficiency. This is illustrated graphically in Fig. 10. Five minutes
of fan operation after burner cutoff apparently can recover approximately 70
percent of the hardware-stored heat which would otherwise be lost. Even so, the
cycle efficiencies are about 5 percent lower than the steady-state values. Even
longer fan operation would help recover some of that loss, but not all of it is
available because losses to the surroundings continue while it is being recovered.
The natural draft air flow, and therefore the standby flue gas heat losses, are
dependent upon ambient outdoor temperature. As a result, cycle thermal effi-
ciency is somewhat more dependent upon ambient temperature (Fig. 11) than is the
steady state (Fig. 6).
Cyclical heat losses can be reduced, conceptually, in a number of ways. If the
firing rate, average cycle time, and ratio of on-time to cycle-time are fixed,
one must consider: better external insulation (lower Q ); lighter weight,
more compact firebox and heat exchanger components; and, possibly, a method to
eliminate natural draft air flow (i.e., mechanical solutions). If the firing
35
-------
0)
u
r-
«*-
cu
u
Nominal Values for All
Parameters Except
Stoich. Ratio
Figure 10. Calculated Effect on Oil Furnace Cycle Thermal Efficiency
of Continuing to Circulate Warm Air Furnace Coolant After
Burner Cutoff
36
-------
100
Nominal Values for All
Parameters Except
OJ
"u
-------
rate and cycle parameters can be varied, it might be possible to reduce losses
by tailoring the cycle to the heating demand load. For a fixed heating demand,
the firing rate would be invariant with cycle time per se^ but would vary inversely
with the ratio ton/tcvc- Analyses were made with values of fractional burner on
time of 1/5, 1/3, and 1/2. Under the assumption of constant thermal demand, the
corresponding firing rates were 1-2/3, 1, and 2/3 ml/s (gph) , respectively. Cal-
culated cycle thermal efficiency results are plotted in Fig. 12 for a constant
stoichiometric ratio of 1.5 and for t,. = t +5. Plotted as the zero-cycle-
ran on '
time intercepts are the calculated steady-state efficiencies for the various
firing rates. The dashed lines extend through a region where the warm-a'ir fur-
nace coolant must be circulated continuously (i.e., At <. 5 min) .
Examination of Fig. 12 shows that there is very little influence of cycle time
per se3 on cycle efficiency but that significant gains result from approaching
continuous firing. If more nearly continuous firing were achieved by throttling
the oil flow, and flue gas temperature were to remain constant as in Fig. 12, on
the order of 5 to 10 percent higher thermal efficiency would be expected. With
a fixed heat exchanger however, substantially lower flue gas temperatures should
be experienced during throttling, resulting in even higher efficiency gains.
However, if the furnace system has been designed to give high steady-state thermal
efficiencies by producing low flue gas temperatures, throttling or modulating the
oil flowrate could easily drive the flue gas temperature into the neighborhood of
the dew-point with concomitant corrosion problems.
Hydronic Boiler Performance. Several FURNAC cases were run with appropriate input
parameters modified to approximate a hydronic boiler. Specifically, the "lumped
parameter" firebox and heat exchanger weight and mean temperature were altered to
include approximately 110 .kg of water at about 75 C as a heat sink rather than as
a circulating coolant.
Calculated thermal efficiencies for hydronic cases were compared (at corresponding
cycle and firing rates, stoichiometric ratios, and ambient and flue gas tempera-
tures) with warm-air cases in which the warm-air fan ran for 5 minutes after
38
-------
100 .
Nominal.Values for All Parameters
90.
*!. 80-
, 60-
50-
4n
Except Mf, Ab, ton, tof, tf
*fan = ^ff + 5
RStoich = 1'5 00-0* C02)
^ ton/tcyc (Firing Rate, ml/s)
Vx-, , _ 0.50 (0.67)
\ " 0.33 (l.OOJ - -
N\^^^ 0.20 (1.67)
10 20 30
Cycle Time, minutes
40
Figure 12. Calculated Effects on Oil Furnace Cycle Thermal
Efficiency of Varying Cycle Timing at Constant
Thermal Demand
39
-------
burner cutoff. Generally, the hydronic efficiencies were within about 1 to 2 per-
cent of those for the corresponding warm-air cases. Although the draft air losses
were initially substantially lower for the hydronic unit than for warm-air, they
continued essentially unabated for the entire burner offtime; their persistence
appeared to effectively counterbalance their lower level.
Heat Transfer Analysis Method
The foregoing oil furnace thermal analyses were performed by treating flue gas
and furnace coolant outlet temperatures as known, constant-valued parameters.
Optimizing efficiency requires careful selection of a nominal design point and
more detailed analysis of furnace behavior under a variety of ambient and demand
conditions. Analyses of that type would best be done with a furnace performance
analysis model which includes a treatment of the dynamics of furnace heat exchange
so that furnace coolant and flue gas temperatures can vary dynamically. Such a
capability is also needed for analyzing effects on efficiency of flue gas
recirculation.
To facilitate more realistic analyses of design aspects (such as firebox and heat
exchanger size and construction), operational aspects (such as firing level,
cycle timing, and climatic conditions) and pollutant reduction concepts (such as
flue gas recirculation and water-emulsified fuel), an expanded-capability com-
puter model for a warm-air oil furnace was formulated using a rather general,
physically simplified, lumped-parameter approach. The resultant formulation
was programmed as a warm-air oil-furnace model which marches in time through an
operating cycle and calculates combustion gas, draft air, warm air, firebox (lined
or unlined), and heat exchanger cooling and heating as functions of time. Entitled
WAFURN, that computer program is described and listed in Appendix A.
Furnace Cyclical Operation Analysis. The WAFURN computer program was run with a
variety of inputs to: (1) assess briefly its ability to simulate furnace behavior
and (2) re-examine the influence on cycle thermal efficiency of some parameters
previously indicated as being important.
40
-------
For comparison with the overall cycle thermal efficiency data of Howekamp and
Hooper (Ref. 7), WAFURN input parameters corresponding to a Lennox OF7-105M
furnace were used. (The construction and size of this Lennox furnace are quite
similar to the Williamson furnace used by Howekamp and Hooper.) Calculated over-
all cycle thermal efficiency* data at four stoichiometric ratios are plotted as
black triangles superimposed upon Howekamp and Hooper's curves in Fig. 13. The
calculated results are seen to be in quite good agreement with the experimental
data. There is, however, an apparent curvature in the calculated results not
indicated by the straight line representations of the experimental data. A simple
explanation of the curvature comes from examination of the calculated flue gas
temperatures, which tend to increase with increasing stoichiometric ratio but
approach an asymptotic value at about a stoichiometric ratio of 1.8:
Stoichiometric Ratio 1.2 1.5 1.8 2.1
T (10 minutes), K = 600 638 653 653
(F) = (620) (689) (715) (715)
However, two additional factors may contribute to the differences between the
calculated and experimental results. These are that WAFURN assumes 100 percent
combustion efficiency in the firebox; whereas, there is evidence of significant
heat losses due to incomplete combustion in the Williamson furnace at low stoich-
iometric ratios (Ref. 2). The other factor is the loss of heat due to draft air
flowing through the furnace and up the flue during the burner off time. This
amounted to less than 1/2 percent in the WAFURN calculated cases, limited by hav-
ing a barometric control damper 2 m above the center of the firebox. Had these
losses been calculated as being higher, they would have reduced the curvature of
the calculated efficiency plot, because the draft air flowrate is approximately
proportional to stoichiometric ratio (Eq. 9).
"Thermal efficiency was defined by Howekamp and Hooper as the total heat trans-
ferred to the warm-air stream divided by the total net heating value of oil
burned in a cycle. This is completely consistent with the definition in WAFURN.
41
-------
100
90
80
70
2 60
o>
Q.
O 50
UJ
O
o: 40
UJ
30
20
10
DEVICE
A
B
C
1 1
I T
t = 30 min
eye
t =10 min
on
I. I I I
1 1
1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4
STOICHIOMETRIC AIR-FUEL RATIO
2.6 2.8
Figure 13. WAFURN-Calculated Overall Cycle Thermal Efficiency
Compared With Measured Efficiencies (Ref. 7)
42
-------
For further assessment of the WAFURN computations, the firing rate was reduced
from 1.00 ml/s (0.95 gph) to 0.75 ml/s (0.71 gph) and the calculated flue gas
temperatures were compared with experimental temperatures measured during steady-
state operation of the OF7 furnace with a Union AFC burner (Ref. 5). Over the
1.2 to 2.1 stoichiometric range, the measured temperatures ranged from 528 to
579 K (490 to 582 F). Tenth-minute WAFURN-calculated flue gas temperatures for
the same stoichiometric range fell between 528 and 605 K (490 to 630 F). Al-
though the calculated temperatures differ significantly from those measured, it
was believed that the correspondence was good enough, particularly in the 1.2 to
1.5 stoichiometric range, to support immediate application of the computer model
without further empirical "tuning up" of the heat exchanger film coefficient
expressions.
Calculated net cycle thermal efficiencies were approximately 5 percent higher
at the 0.75 ml/s (0.71 gph) firing rate than at the 1.00 ml/s (0.95 gph) rate,
predominantly as a result of the lower flue gas heat losses attending reduced
flue gas temperatures. Another way of reducing flue gas temperatures is to utilize
a more effective heat exchanger. Increased heat exchanger effectiveness might
be achieved in several ways; the simplest way is to increase the heat transfer
area. To illustrate the potential performance improvement, WAFURN was run with
Lennox OF7 inputs, except for a 1/3 greater heat exchanger area.* Computed 4th
minute flue gas temperatures and cycle efficiencies for a 12-minute cycle and a
4-minute burner operation at the 1-ml/s firing rate were as follows:
Stoich.
Ratio
1.2
1.5
Heat Exchanger
Area, m^
2.25
3.00
2.25
3.00
Flue Gas Temperature
K
599
455
637
495
(F)
618
359
686
432
Cycle Thermal
Efficiency, %
82.25
89.39
76.72
85.25
*Although the mass of the heat exchanger was left invariant for these computer
program runs, it was varied parametrically in other runs by ±25 percent and
found to influence flue gas temperature by less than 1 K, and cycle efficiency
by less than 0.5 percent.
43
-------
Cycle Timing Effects. In view of the calculated variability of cycle thermal
efficiency with firing rate, cases were run to examine the dependence of efficiency
on cycle timing. A situation in which thermal demand and flue gas temperature are
both constant was considered previously using the FURNAC computer model; the
results are reproduced from Fig. 12 as dotted curves in Fig. 14. WAFURN-calculated
results for the fixed heat exchanger of the Lennox OF7 are the solid curves in
this figure. It is seen that there is only a small predicted effect of cycle time
per se on furnace efficiency, but that a substantial efficiency penalty can be
expected if a given thermal load is satisfied by brief firings at elevated firing
rates. Reduced efficiencies at higher firing rates are caused primarily by in-
creases in flue gas temperature.
It is constructive at this point to examine the effects of cycle timing in satisfy-
ing a variable thermal load with a constant firing rate. For the Lennox OF7 fired
at 1.00 ml/s (0.95 gph) and 1.5 stoichiometric ratio, the following results were
calculated by WAFURN:
t cycle'
minutes
12
20
30
t /t .
on cycle
0.20
0.333
0.50
0.333
0.20
0.333
0.50
Tfg at 'off
K
647
637
638
637
636
638
638
(F)
704
686
689
686
685
689
689
\hermal' %
76.01
76.72
77.72
76.78
76.99
77.79
78.05
Comparison of these efficiency values with those plotted in Fig. 14 reveals that
flow modulation is almost wholly responsible for the wide range of efficiency
variation in Fig. 14. Apparently, at a constant firing rate, quite wide excur-
sions of cycle time and burner-on time can be experienced without incurring
44
-------
Lennox OF7 Furnace
1.5 Stoichiometric Ratio
100
90
* 80
o
O)
'o
70
60
Case:
o11:
Constant Flue Gas Temperature
Fixed Heat Exchanger
A, B, C
0.50, 0.333, 0.20
0.667, 1.00, 1.667 ml/s .
A
o ©
I
1
I
0
10 20 30
Cycle Time, minutes
40
Figure 14. Calculated Effects on Oil Furnace Cycle Thermal
Efficiency of Varying Cycle Timing at Constant
Thermal Demand
45
-------
substantial deviations from steady-state thermal efficiency.* If that constant
firing rate is the design rate, the flue gas temperature approaches its design
value closely in all but the very shortest firings. Higher firing rates lead
to reduced efficiency primarily because the flue gas temperature increases. Con-
versely, firing rates lower than the design value result in increased thermal
efficiency by decreasing the flue gas temperature. Assuming that the design
value of flue gas temperature was selected to avoid condensation in and corrosion
of the heat exchanger and flue, the lower temperatures attained by lowering the
firing rate should be unacceptable. Thus, the two constraints of maintaining
high efficiency and maintaining some minimum level of flue gas temperature pro-
vide persuasive arguments against designing for reactant flow modulation in resi-
dential furnaces having fixed heat exchangers.
High/Low/Off Burner Operation. There appears to be a modest efficiency
advantage in designing for high ratios of burner firing time to cycle time. This
suggests that high/low/off burner control might provide significant fuel savings
over an entire heating season. The furnace would be designed with a two-firing
level capability. Continuous firing at the lower level would meet or exceed the
residential thermal demand for some large fraction of the demand time. Conven-
tional on-off control would be used at this firing level, which would be selected
to give maximized efficiency and operation - on the average - much closer to steady-
state than is conventional practice. The higher firing level would be called upon
when the demand exceeds the continuous-firing capability of the lower firing level,
and on-off control would be supplanted by high-low control. The idea is to accept
brief periods of reduced efficiency at the high level to permit operating at en-
hanced efficiencies for long periods.
*This observation is applicable to well-designed and well-insulated furnaces
which have low external convective and radiant heat losses to the surroundings
and low draft air losses during the bumer-off time.
46
-------
This concept was evaluated breifly by providing a high-low optional analysis path
in WAFURN and calculating what might happen if the Lennox OF7 furnace were modified
for low fire at 0.75 ml/s (0.71 gph) and high fire at 1.125 ml/s (1.07 gph).
WAFURN-calculated cycle efficiencies for 12 minute cycles are plotted in Fig. 15.
High/low/off operation at two stoichiometric ratios is shown as solid curves.
The 100 percent point on the abscissa represents steady-state low-fire, and the
150 percent point represents steady-state high fire. Slightly decreasing effi-
ciency with decreasing demand is caused by starting transient losses and by draft
air losses during the burner off times. The more abrupt efficiency decline as
demand exceeds 100 percent is the result of increased flue gas temperatures during
high-fire times.
To assess the impact of such operation on overall season-averaged efficiency, it
was assumed, for simplicity, that the time distribution of firing levels during
some hypothetical heating season has a normal distribution about the 75 percent
firing level point. The calculated firing-level-weighted and firing-time-weighted
overall seasonal average efficiency was 87.05 percent at 1.2 stoichiometric ratio.
If the high-fire capability were deactivated, the demand would exceed the low-fire
heat supply about 15 percent of the time and the average efficiency would rise to
87.47 percent. Thus, an almost negligible efficiency penalty might be paid for
the extra high-fire capability.
There are, of course, other ways to provide the thermal demand. One way is to
simply use on-off control at the high-fire level. Peak (steady-state) efficiency
is represented by the 150 percent point, and that is seen to be about 5 percent
below the on-off/high-low seasonal average. A more meaningful comparison is to
increase the heat exchanger area so that steady-state high-fire flue gas temper-
atures would be comparable to those for steady-state low fire. Calculated results
for on-off control of the high-fire level using a 20.4 percent greater heat ex-
change area are plotted as dashed curves in Fig. 15. To match the seasonal heat-
ing demands, shorter burner on times/cycle are needed for this situation. The
seasonal average efficiency was calculated to be 86.75 percent or only 0.3 per-
cent below that for the more complicatedand more costly--on-off/high-low operation.
47
-------
100
90
c
OJ
I
O
r~
«*-
(0
O)
-C
0)
O
80
70
Hiqh/Low/Off Burner Control
On-Off High-Fire with 20.4% Greater Heat
Exchanger Area
S.R.=1.2
20
On-Off
Control
High-Low
Control
50 100
Relative Thermal Demand, % of Heat Input at a
Steady 0.75 ml/s Oil Flowrate
150
Figure 15. Calculated Thermal Efficiencies for Lennox OF7
Furnace for Some Dual-Firing-Level Situations
48
-------
An efficiency difference of 0.3 percent is equivalent to that which would be
gained by providing a 1 to 2 percent greater heat exchange area or by operating
at a stoichiometric ratio that's smaller by about 0.01 to 0.02. Further, the
furnace's electrical components would be required to run nearly 45 percent longer
for the high-low operation than for the simpler on-off at high fire. The addi-
tional electric power consumption undoubtedly would cost considerably more than
the small quantity of fuel saved and, in addition, more frequent maintenance and
earlier equipment replacement should be anticipated. Thus, this kind of modulated-
flow furnace control does not appear to be worthy of further consideration.
Water-Emulsified Fuel Oil. Emulsifying"fuel oil with water, under the FURNAC
assumption of invariant flue gas temperature, was found to effect an appreciable
decrease in thermal efficiency which depended only upon the mass fraction of water
in the fuel (Fig. 8). For example, 40 percent water in No. 2 fuel oil lowered
the calculated thermal efficiency by about 5 percent over the stiochiometric range
from 1.0 to 3.0. However, because the composition, temperature and mass flux of
combustion product gases all change with varying fuel water content, a fixed heat
exchanger should not be expected to produce a constant flue gas temperature. To
get a more realistic preliminary indication of this effect, WAFURN was run with
a 40 percent water emulsion supplied to the Lennox OF7 furnace at a dry oil flow-
rate of 1.00 ml/s, several stoichiometric ratios and a 4-minute firing in a 12-
minute cycle. The following tabulation characterizes the results.
Stoichiometric
Ratio
1.2
1.5
1.8
% H20
in Oil
0
40
0
40
0
40
Combustion
Temperature ,
K
2081
1929
1810
1690
1605
1509
F
3285
3011
2729
2583
2430
2255
Fourth-Minute
Flue Gas Temperature,
K
599
613
637
640
651
648
F
618
643
686
693
712
707
Cycle-Averaged
Thermal
Efficiency, %
82.25
77.41
76.72
72.09
71.96
67.47
It is seen that, although the combustion zone temperature is lowered by 100 C or
more by diluting the fuel with water, the flue gas temperature is changed only
slightly. Thus, the emulsion's impact on furnace cycle-averaged efficiency is
quite comparable to that shown in Fig. 8.
49
-------
Three ways in which emulsified oil could conceivably be beneficial (Ref. 8) were
not included in the WAFURN calculations. First, if the oil spray nozzle were
producing a coarse spray, leading to substantial combustion inefficiencies, emul-
sified fuel would tend to correct this poor operation by steam generation within
heating droplets, causing them to disintegrate and burn more rapidly. It might
then be possible to operate smoke-free with significantly less excess air and thus
gain slightly in efficiency, rather than losing. Second, if the emulsified fuel
flowrate were held constant (rather than dry oil flowrate, as above), the frac-
tional burner-on-time would be increased and standby losses reduced proportionately.
As indicated by the slopes of the curves on Fig.15 at demands below 100 percent,
the potential gain from this effect would normally be only a small fraction of
the loss caused by adding water to the fuel. Finally, if the heat exchange sur-
faces were sooted up, the extra water vapor in the combustion gases could con-
ceivably aid in its oxidation and removal. However, combustion gases from burn-
ing dry oil contain about 63 percent as much water vapor as those produced by
burning a 40 percent water-oil emulsion. Thus, dramatic efficiency increases, as
a result of emulsifying the fuel with water, appear to be highly unlikely.
Flue Gas Recirculation. The WAFURN computer program was used to analyze
some effects of recirculating flue gases and mixing them with the combustion air
supplied to the burner. Calculations were made for the Lennox OF7 furnace fired
for 1/3 of a 12-minute cycle. Flue gas recirculation (FGR) effects on cycle
thermal efficiency and flue gas temperature are shown graphically in Fig. 16. At
near-stoichiometric conditions, imposition of FGR causes the calculated flue gas
temperature to increase substantially. As the excess air is increased, however,
this effect becomes decidedly less pronounced until, at a stoichiometric ratio
of about 2, the flue gas temperature and thermal efficiency apparently are little
influenced by the presence or level of FGR*.
*Essentially identical trends were observed in comparable calculations for a
10-minute firing, 30-minute cycle condition.
50
-------
90
80
I
(U
T-
U
70
60
1.0
Lennox OF7 Furnace, 4 min. Firing in 12. min Cycle
700
No FGR
20% FGR
40% FGR
l
1.5
Stoich. Ratio
2.0
650
>o
-------
The adiabatic flame temperatures for conditions represented on Fig. 16 decrease
monotonically with both increasing stoichiometric ratio and increasing FGR:
% FGR
0
20
40
Ac
RStoich = l'2
2081 (3285)
1747 (2685)
1549 (2528)
iabatic Flame Temperatures, K (F'
1.5
1810 (2797)
1546 (2322)
1345 (1961)
1.8
1605 (2430)
1340 (2043)
1242 (1776)
2.1
1444 (2139)
1265 (1817)
1118 (1552)
Therefore, the calculated reduced effects of FGR on flue gas temperature as
stoichiometric ratio is increased cannot be explained by reactant/diluent
thermochemical phenomena. Rather, they must result from the dynamic interaction
of flowrate and supply temperature in altering the heat transfer behavior of the
furnace's heat exchanger.
Refractory-lined versus Unlined Fireboxes. As a final example of the
WAFURN computer model's analysis capabilities, the 0.019 m (0.75-inch) thick
refractory-lining in the Lennox OF7's firebox was assumed to be made 1/3 as
thick in one case and removed altogether in another. Calculated results for a
4-minute firing at 1.00 ml/s (0.95 gph) oil in a 12-minute cycle were as follows:
Stoich.
Ratio
1.2
1.5
1.8
Refractory
Thickness ,
m (inch)
0.019 (0.75)
0.0063 (0.375)
0.0
0.019 (0.75)
0.0063 (0.375)
0.0
0.019 (0.75)
0.0063 (0.375)
0.0
Refractory
Surface
Temperature ,
K (F)
1796 (2773)
1541 (2314)
(527)* ( 489)
1532 (2298)
1296 (1873^
(475) *( 395)
1335 (1943)
1117 (1551)
(440) *( 332)
Gas Temperature, K (F)
Firebox Exit
1984 (3111)
1878 (2920)
1700 (2600)
1741 (2674)
1669 (2544)
1562 (2352)
1556 (2340)
1504 (2247)
1434 (2121)
Flue
599 (6i8)
599 (618)
596 (613)
637 (687)
634 (681)
626 (667)
651 (712)
647 (70S)
638 (688)
Thermal
Efficiency,
%
82.25
83.31
83.75
76.72
77.80
78.44
71.96
72.92
73.80
( ) represents uninsulated metal temperatures
52
-------
It is seen that the computer model predicts that approximately 1-1/2 to 2 percent
higher thermal efficiencies can be achieved by designing an air-cooled firebox
without a refractory lining. The temperatures at burner cutoff tabulated above
do not particularly indicate such an appreciable difference, but examination of
the complete cycle printouts revealed that the flue gas temperatures early in
the burner-on times were substantially lower for the unlined case, which accounts
for the higher efficiency.
53/54
-------
SECTION V
EXPERIMENTAL INVESTIGATION
An experimental investigation was conducted in which pollutant emissions and com-
bustion performance were characterized as functions of burner and combustion cham-
ber design and operating conditions. The objective was to define, in conjunction
with the analytical investigation, the design requirements for an integrated fur-
nace system capable of meeting the emission and efficiency goals stated in the
Introduction.
The experimental investigation, which utilized research combustion chambers rather
than typical furnaces, consisted of three separate, but related parts. One part
dealt exclusively with the 1 ml/s (gph) conventional burner, optimized to produce
low emissions, as described in Ref. 5. This effort sought to optimize the match-
ing of the firebox (combustion chamber) to that fixed-burner design. The other
two parts extended to burners having, in one case, forced flue gas recirculation
(FGR) to the burner air intake and, in the other case, forced combustion gas re-
circulation (CGR) from the combustion chamber to the burner air intake. These
efforts were directed toward optimizing the burner/firebox combination for low
emissions. The three segments of the experimental investigation are described,
and the results obtained are presented and discussed in the following subsections.
OPTIMIZED CONVENTIONAL BURNER STUDY
A previous intensive Rocketdyne investigation of residential and commercial oil
burners (Ref. 5) led to criteria for optimizing conventional burner designs with
respect to pollutant emissons. For high-pressure, atomizing, luminous-flame
burners fired into refractory-lined combustion chambers, minimum pollutant emis-
sions 'were obtained with burners having (1) no flame-retention device, (2) choke
diameter related quantitatively to the firing rate, and (3) large peripheral swirl
vanes oriented at 25 degrees relative to the blast tube axis. This swirl vane
angle gave the best comprimise between smoke emissions and nitric oxide emissions.
An optimum burner having a 1 ml/s (gph) firing rate is illustrated in Fig. 17.
55
-------
(a) External View
5DZ21-8/6/73-S1
(t>) Optimum 1 ml/s
Burner Head
5DZ21-8/6/73-S1A
Figure 17. Optimum 1 ml/s (gph) Oil Burner
-------
In addition to minimizing nitric oxide formation, the optimum burners were capable
of operating with substantially less excess air than is used in conventional prac-
tice, without producing unacceptable levels of carbonaceous pollutants (CO, UHC,
and smoke). As discussed in Section IV, such operation is beneficial in that it
enhances overall furnace thermal efficiency.
Nitric oxide emissions were lower when the optimum burners were tunnel-fired into
refractory-lined fireboxes than when the burners were side-fired. The explanation
offered in Ref. 5 for this result was the same as that for obtaining lower NO emis-
sions with nonretention burners than with those having flame retention heads. That
is, at low excess air levels, formation of NO is minimized by the absence in the
combustion zone of recirculation eddies, steep concentration gradients, and strong
mixing. A uniform dispersion of fuel spray in an air stream that flows and reacts
smoothly along the combustion chamber (i.e., in the manner of a plug flow reactor)
produces the least nitric oxide, while vigorous intermixing of incoming fresh air
with combustion products (as a result of adiabatic internal recirculation or steep
gradients in the combustion field) tends to promote production of NO.
Combustion chamber design features other than burner orientation were varied in
the Ref. 5 studies, but not sufficiently to delineate clear trends or to derive
optimization criteria. However, there appeared to be significant effects on pol-
lutant formation and emission concentrations of varying chamber diameter and length,
at a given firing rate, and of providing nonadiabatic combustion zone conditions
by cooling the combustion chamber walls. It was noted that recirculation of com-
bustion gases within cooled (or partially cooled) combustors, was not necessarily
detrimental and could even be beneficial with respect to reducing smoke and NO
emissions.
This current experimental effort was undertaken to examine more thoroughly the
matching of fireboxes to the optimum burner. The objective was to derive design
criteria for further minimization of pollutant emissions and for acceptable oper-
ability (e.g., ignition, flame stability, and component temperature).
57
-------
Experimental Apparatus
Optimum Burner. The 1 ml/s (gph) optimum oil burner illustrated in Fig. 17 was
used. This burner consisted of a conventional burner (Beckett Model AF) body
fitted with an optimized nonretention burner head at the discharge end of its
0.10 m (4-inches) diameter blast tube. The optimum head conformed to the design
criteria given in Ref. 5. In the 1 ml/s (gph) size, the choke diameter was 0.042
m (1.65 inches), and the swirl vanes were relatively large, being approximately
0.05 m (2 inches) long and extending from about 0.03 m (1.2-inches) diameter out
to the inside diameter (ID) of the blast tube. As noted earlier, the vanes were
canted at 25 degrees from the blast tube centerline, cut from 0.0016 m (1/16-inch)
Type 321 stainless sheet, and welded to the machined choke plate of the same
material.
Most other burner components, viz, the housing, blast tube, fuel pump, combustion
air fan, drive motor, ignition transformer, CdS flame-sensing cell, and burner
control circuitry, were stock equipment supplied with the burner. The stock
ignition electrodes were replaced by a smaller-diameter pair compatible with the
more restricted space imposed by the swirl vanes. Use of a 1.0-60°-A Delavan
spray nozzle provided good mixing between the fuel spray and air while minimizing
spray impingement on combustion chamber walls.
As the testing progressed, some alterations of the optimum burner were made in
order to satisfy specific test needs that arose. These alterations are later de-
scribed and discussed in context with the experimental reasons for making them and
the results achieved. Pertinent information appears on page 69, et seq., and page
71, et seq.
Research Combustion Chambers. Three combustion chambers were built to provide a
matched set with considerable dimensional, structural and coolant variabilities.
The basic approach is illustrated in Fig. 18 for one of the chambers. Each cham-
ber was a 1.52 m (5-foot) long, flanged section of steel pipe with a stubby, flanged
side-arm section of the same size of pipe attached near one end. The tunnel-fired
58
-------
REFRACTORY INSERT
0.013 m PYROFLEX LINER
(DOUBLE THICKNESS
in
0.25 . °- °0
OD '
BLANK
FLANGE
SIDE-FIRED
BURNER PORT
V MOVABLE HEAT
EXCHANGER (SPIRAL
WOUND FINNED-TUBE)
HEAT EXCHANGER
WATER
0.25 m DIAMETER
COMBUSTION CHAMBER
TUNNEL-FIRED BURNER
PORT END FLANGE
Figure 18. Experimental Combustion Chamber and Heat Exchanger Arrangement
-------
orientation is illustrated in Fig. 18, in which the oil burner fits into the annular
flange depicted at the left end of the chamber. In this orientation, the side-arm
was redundant and so was positioned at the opposite end from the burner and simply
blanked off. To achieve the side-fired configuration, the combustion chamber was
simply turned end-for-end, with the blank and burner port flanges relocated as
appropriate. The steel chamber could either be lined with a refractory fiber in-
sert to form an adiabatic combustion zone or remain unlined and in some way be
cooled. A movable, spiral-wound, finned-tube heat exchanger, shown inserted in the
opposite end of the combustion chamber from the burner, provided effective varia-
tion of the combustion zone length.
Three chamber diameters were selected such that addition of refractory linings to
the larger two would produce lined chambers having inside diameters comparable
with the smaller two unlined chambers, viz.:
Steel Pipe Nominal Combustion Nominal Thickness
ID, m (inches) Chamber ID, m (inches) of Pyroflex, m (inches)
0.162 (6.36) 0.162 (6.36)
0.222 (8.75) 0.222 (8.75)
0.222 (lined) 0.175 (6.89) 0.024 (0.93)
0.279 (11.0) 0.279 (11.0)
0.279 (lined) 0.22 (8.7) 0.030 (1.18)
In use, the major axes of the chambers were vertical, with the burner firing ver-
tically upward when tunnel-fired and horizontally when side-fired. There were
several reasons for selecting this arrangement: (1) it provided simulation of
firebox draft and combustion-gas flow patterns present in a majority of furnace
designs, (2) the spiral-wound heat exchanger could be simply suspended in the com-
bustion chamber without having to provide lateral support, (3) natural convection
air cooling would be more uniform than it would be with a horizontal axis, and
(4) water-cooling could be effected with a fairly simple cylindrical housing
around the chamber, open to the atmosphere at the top of the housing.
60
-------
Combustion-Gas Heat Exchanger. A water-cooled heat exchanger was used to accomp-
lish rapid cooling of the combustion gases as they flowed out of the primary "fire-
box" portion of any of the combustion chambers. The intention was for the gas
temperature to be quenched rather rapidly so that changes in heat exchanger posi-
tion (e.g., firebox length) could be readily correlated with variations in pollut-
ant emissions.
As noted earlier, the heat exchanger was a spiral-wound, finned-tube assembly de-
signed to fit inside all three combustors and to be positioned anywhere along the
length of a chamber. Detailed dimensional data for the coil are given in Fig. 19;
this coil was designed to cool the oil burner exhaust gases down to about 200 C
-4 3
to 350 C (390 F to 660 F) at a maximum water coolant flowrate of 3.8 x 10 m /s
(6 gal/min). The 0.025 m (1-inch) outside diameter (OD) finned tubing used, has
a 0.013 m OD by 0.0012 m wall (1/2-by 0.049-inch) stainless steel tubing with
helically wound 0.00051 m (0.020-inch) thick carbon steel fins with nickel-chrome
clad plating, 394 fins/m (10 fins/inch). The carbon steel fin material was selected
over stainless steel because such material has better thermal conductivity proper-
ties, and the nickel-chrome clad plating provides resistance to corrosive exhaust
products. Approximately 4.1 m (13-1/2 feet) of finned tubing was required in the
construction of the heat exchanger, resulting in a total heat transfer surface
area of about 1.4 m2 (15 ft2).
Several semicircular baffle plates were cut from a 21-gage stainless steel sheet,
slipped between coils of the heat exchanger and wired in place. Cut so their
outside diameters would just fit comfortably in a chamber's inside diameter, the
baffles ensured that the gases passed repetitively over the heat exchanger coils
and prevented them from bypassing around the outside of the coils.
Midway through the experimental studies, the cooling water was inadvertently shut
off and, after a few minutes of burner firing in an insulated, adiabatic combustion
chamber, the heat exchanger was destroyed by massive melting of both the fins and
coiled core tubing. Upon finding that a replacement finned-coil could not be de-
livered in fewer than 6 weeks, a substitute heat exchanger was fabricated by coil-
ing about 15 m (49 feet) of 0.001 m (3/8-inch) copper tubing into two nested,
61
-------
Approx. 1.25m
of Bare Tubina
Colls Req'd
shown)
V
.15m O.D.
(X.06m
^X Coll -Spacing
Typ.
0.019m R
Typ. 3 Places
0.025 m (1.0-inch) OD Helically-Wound Finned Tube (Approximately 4.1 m (13,5 feet)
Finned Length)
Core Tube: 0.013 by 0.0012 m (1/2-inch by 0.049-inch) Stainless Steel '
Fins: 0.006 by 0.0005 m (1/4-inch by 0.020-inch) Carbon Steel - Nickel/
Chrome Clad 297 Fins/Meter (10 Fins/Inch)
Figure 19. Heat Exchanger for Variable Combustion
Chamber Configuration Tests
62
-------
parallel-plumbed coils. This device was checked out and found to be entirely sat-
isfactory, and so was used for the remainder of the Phase I experimental work.
Test Facility. The optimum burner was fired in the various research combustion
chambers at an outdoor test facility illustrated schematically in Fig. 20. The
principal components were attached to a waist-high steel table as shown. Not shown
is a Unistrut superstructure at the right-hand end of that table to support the
vertically mounted combustion chamber and to suspend the spiral-wound heat exchanger
within it. The facility was organized for rapid and easy changing of combustion
chambers, burner orientation, and heat exchanger position. Minimum protection from
inclement weather was provided by a simple sheet metal roof over the test apparatus.
Experimental data requirements were primarily concerned with flue gas pollutant
concentrations. Concentrations of most pollutant species were measured by conduct-
ing a continuous flue gas sample to a train of analysis instruments located indoors
in a nearby laboratory. Flue gas smoke content was measured intermittently at the
flue with a manual smoke meter. The instruments used, analyses performed, and
types of data obtained are described and discussed in Appendix B. In addition,
the firing rate was monitored regularly by measuring the fuel oil flowrate, the
flue gas temperature was indicated by immersing a thermocouple downstream of the
heat exchanger, and the temperature rise of the heat exchanger coolant water was
measured. Miscellaneous data taken less regularly, were firebox draft conditions,
firebox shell metal temperatuares, and combustion air fan characteristics.
Experimental Test Matrices
Two matrices of 80 tests each were planned for experimentally matching combustor
designs to the optimum oil burner. One matrix was for a tunnel-fired burner or-
ientation and the second was for a side-fired burner arrangement. The planned
distribution of tests was the same for both matrices and is denoted in Table 7.
Variations of chamber wall construction of cooling medium, chamber inside diameter
and length preceding the heat exchanger, and excess air level (stoichiometric ratio)
were included. All tests were planned as transient (cyclical) tests, with the
burner fired for 10 minutes of a 30-minute cycle.
63
-------
HEAT EXCHANGER
WATER COOLANT
COMBUSTOR
RESERVOIR
FUEL SYSTEM
SHUTOFF VALVE
OIL BURNER
ICE BATH
WATER VAPOR TRAP
BURNER ON-OFF
SWITCH
TO GAS ANALYSIS
INSTRUMENTS
Figure 20. Schematic of Oil Burner and Research
Combustion Chamber Test Installation
64
-------
TABLE 7. OPTIMUM BURNER/CHAMBER MATCHING TEST MATRIX
Chamber Wall
Air-cooled
i
Fractional
Air-cooled
Refractory- lined
i
Water-cooled
i
<
Chamber Diameter,
metre (inch)
0.162 (6.36)
1
0.222 (8.75)
i
*
0.279 (11.0)
1/4 0.222 (8.75)
1/2
3/4
i
r
0.221 (8.70)
i
0.175 (6.89)
i
0.222 (8.75)
i
Chamber Length,
metre (inch)
0.30 (12)
0.40 .'(16)
0.50 (20)
0.75 (30)
0.30 (12)
0.40 (16)
0.50 (20)
0.75 (30)
0.30 (12)
0.40 (16)
0.50 (20)
0.75 (30)
0.40 (16)
\
r
0.30 (12)
0.40 (16)
0.50 (20)
0.75
0.30
0.40
(30)
(12)
(16)
0.50 (20)
0.75
0.30
0.40
0.50
0.75
(30)
(12)
(16)
(20)
(30)
No. of Excess
Air Levels
3
r
3
'
3
i
3
i
3
i
3
\
3
1
1
3
65
-------
"Air-cooled" chambers consisted of the uninsulated steel combustion chambers simply
mounted vertically in the open air. As a result, the cooling was a combination of
radiant heat transfer to the surroundings and natural convection heat transfer to
ambient air. For the "fractional air-cooled" chamber configuration, the 0.222 m
(8.75-inches) diameter chamber was blanketed externally with an approximately
0.013 m (1/2-inch) thick layer of Pyroflex refractory insulation. With the heat
exchanger set to form a 0.40 m (16-inches) long firebox, that blanket was positioned
initially such that the first 0.1 m (4 inches) of the firebox, near the burner,
was uninsulated and, thus, was 1/4 air-cooled. It was then moved further downstream
in two successive 0.1 m (4-inches) steps to provide 1/2 and 3/4 air-cooled condi-
tions. Attaching the insulation to the outside allowed the steel shell of the
chamber to be overheated, but the risk of damaging it was accepted rather than
having step diameter changes placed inside the combustor by the use of partial-
length internal liners.
Pyroflex linings were placed inside the "refractory-lined" combustors, as described
earlier.
"Water-cooled" chambers were formed by placing an open-topped sheet metal "bucket"
around the firebox end of the combustion chamber and filling the box with water.
Openings in the water-jacket were provided as required for mounting the burner to
the chamber, and a RTV silicone-type, low-pressure, moderate-temperature sealant
was used to prevent water leaks. A flow of fresh water was admitted to the bottom
of the jacket, and a hot-water overflow near the top controlled the water level.
The chamber inside diameters in Table 7 conform to those listed earlier. Variation
of combustion chamber length, achieved by moving the water-cooled heat exchanger
coil as previously described, was intended to cover the range of practicality ex-
hibited by current commercial practice.
Testing at only three excess air levels was recognized as being quite limited. To
avoid operating at leaner conditions, and because operation as close to stoichio-
metric conditions as possible promotes the desired higher thermal efficiency, test-
ing was planned at stoichiometric ratios of approximately 1.10, 1.15, and 1.20.
66
-------
Experimental Results
The methodology used in the experimental testing was to proceed, more or less
sequentially, through the foregoing test matrices, first with the tunnel-fired,
then with the side-fired orientations. Data were recorded for 160 runs (sets of
different apparatus or different test conditions), although exploratory firings
were made at a substantial number of other conditions for which test data were
not recorded for one reason or another. A complete tabulation of the data obtained
is given in Appendix C, listed sequentially by run number.
Tunnel-Fired Orientation Results. The tunnel-fired (coaxial) burner orientation
was tested first (Run 1 through 81).
Air-Cooled Combustion Chambers. The recorded tests were preceded by experi-
ments with the 0.162 m (6.38-inches) diameter, air-cooled combustor, which was
found to be impractical for use with the 1 ml/s (gph) optimum burner. No operat-
ing conditions could be found to eliminate very rough combustion, characterized
by loud, low-frequency rumbling and very high smoke levels. No valid reportable
data were obtained.
With the 0.222 m (8.75-inches) diameter air-cooled chamber, combustion was signifi-
cantly better, although some audible roughness was present in all tests. Start-up
problems were experienced on cycle testing at stoichiometric ratios much removed
from unity. The flue gas pollutant species data are plotted in Fig. 21. The
various line segments on these plots represent the entire operable stoichiometric
range; these segments are generally quite constricted between excessive smoke on
the left and unreliable ignition on the right.
The relative carbon monoxide and unburned hydrocarbon levels suggest that chamber
lengths of about 0.5 m (20 inches) or greater may be needed in order to avoid pre-
mature quenching of the combustion process. Even for the longer chambers, the
quantities of CO and UHC produced were substantially higher than those typically
experienced in conventional residential furnaces (Ref. 4). It was thought that
the rough combustion may have caused these pollutant levels to be high.
67 i
-------
Bacharach Smoke < 1
Bacharach Smoke > 1
1.10
StolchltxnetHc Ratio
1.20
0 ca
L 30 em
(Heat Exchanger
Position)
1.10 1.20
StolcMometHc Ratio, (VF)/14.49
00
1 3
-L * 30 ca (Heat Exchanger Position)
75 en
i no i .20
Stolchlonetrtc Ratio. (A/F)/14.49
10.1
o
u
I 30 ca
75 c«
40 en
1 00
1.10
StolcMomtrlc Ratio
l.ZO
Figure 21. Effect of Combustion Chamber Length Upon Flue Gas Emissions as Functions of Stoichiometric
Ratio in a 0.222 m (8.75 inches) Diameter, Air-Cooled, Tunnel-Fired Corabustor
-------
The NO concentrations measured in the flue gas were substantially lower than the
steady-state levels reported in Ref. 5 for this burner. While this may be attri-
butable to the difference between cycle-averaged and steady-state values, it may
also reflect effects of air-cooling the combustor walls and of radiation from the
flame zone to the water-cooled heat exchanger, both of which could produce moder-
ately lower flame zone temperatures.
Emissions from the 0.279 m (11-inches) diameter air-cooled combustor are plotted
in Fig. 22. The upper stoichiometric ratio limit for these tests was also limited
by conditions where ignition of the air/oil mixture in the cooled combustor became
marginal. This larger combustor produced" slightly higher emissions of CO and UHC
than did the 0.222 m (8.75-inches) chamber, and slightly lower NO emissions. In-
terestingly, among the four tested heat exchanger positions, emission data taken
at 0.4 m showed distinctly different characteristics in both the 0.222 (8.75-inch*«)
and 0.279 m (11-inches) chambers. The reason for this peculiarity was not explained.
(It will be seen later that neither the water-cooled nor insulated chamber data
exhibited a similar peculiarity.)
At this point in the testing, in response to a question as to whether the optimum
oil burner head could be used effectively with burners having smaller blast tubes,
a smaller 0.075 m (3-inches) diameter version of the 1 ml/s optimum head was made
and fired on the same burner body fitted with a 0.075 m (3-inches) blast tube.
Cyclic firings in the 0.222 m (8.75-inches) diameter insulated combustor gave re-
sults directly comparable with those reported in Ref. 5, showing that the more com-
pact optimum head may be applied satisfactorily. That smaller head was retained
thereafter for the rest of the 1 ml/s optimum burner testing.
The fractional air-cooled combustor test sereis was conducted with the 0.222 m
(8.75-inches) diameter combustor wrapped externally with a movable 0.30 m (12-
inches) long section of insulation and with the heat exchanger (i.e., chamber
length) set at 0.40 m (16 inches). External rather than internal insulation was
chosen to eliminate the effect of geometric changes that would be introduced by a
movable internal liner. The objective of this test series was to study the effect
69
-------
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Figure 22. Cycle-Averaged Flue Gas Emission, 1.0 ml/s Optimum Oil Burner in 0.279 m Diameter,
Air-Cooled, Tunnel-Fired Combustor at Various Heat Exchanger Positions
-------
of changes in the location of heat extraction from the combustion zone on flue gas
composition. The resulting flue gas composition data are plotted in Fig. 23. The
numerical labels on the plotted curves indicate the uninsulated (i.e., air-cooled)
fraction of the combustion chamber, measured from the burner-end.
Few differences exist among the curves for the various locations, a result that
had not been anticipated. Two quite different inferences may be drawn from this
result. One is that the combustion gas recirculation eddy, which surrounds the
primary flame zone and supplies it with partially cooled diluent gases, must be
so short that heat removed from the wall further than 0.1 m (4 inches) downstream
does not cool the eddy. The other is that the water-cooled heat exchanger was
cold enough and close enough to the main flame zone to extract sufficient radiant
heat to overshadow the contribution of convective wall cooling to the reduction of
flame zone temperatures. A decision was made not to conduct further exploratory
tests until after the insulated and partially insulated side-fired test results
were considered as contributions to the understanding of this result.
During much of the testing of the air-cooled chambers, the tunnel-fired 1 ml/s
(gph) optimum burner flame would become very rough and unstable soon after ignition,
especially at 20- to 25-percent excess air conditions. Burners with narrow ranges
of operable conditions have restricted adaptability to field conditions and so
would experience limited industry acceptance. Therefore, some peripheral tests
were run to determine a method for alleviating this problem. Much of the past
thinking about combustion roughness and its associated noise has been biased in
the direction of flame front instability and the resulting need for flame retention
devices on oil burners. The 1 ml/s (gph) optimum burner has a nonflame, retention-
type head and that fact has an important bearing on its low emission characteris-
tics. While in the rough combustion mode, increasing the air inlet to full-open
would produce no increase in stoichiometric ratio, indicating that the air fan
operation was limited by other influences.
A two-fans-in-series burner configuration was set up to create a supercharged
burner capability, which sould seriously aggravate a flame-front-detachment-type
instability. However, the supercharged burner was found to be far more stable
71
-------
30
I
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10
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i »
1.10
KRT10
120
15 r
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Figure 23. Cycle-Averaged Flue Gas Emission, 1.0 ml/s Optimum Oil Burner
in a 0.22 m Diameter, Partially Insulated, Tunnel-Fired
Combustor at a 0.40 m Heat Exchanger Position
-------
than the single-fan burner. The combustion roughness is apparently a coupling
between the combustion air fan and the combustion process where fan stall, caused
by flame perturbations, becomes a contributing factor to further perturbations.
The squirrel-cage fans used in most oil burners are characteristically low-pressure,
high-volume devices subject to stalling from slight back-pressure perturbations.
Fan stall may be aggravated further by the inlet-choke air flow control, used on
most burners, which reduces the static pressure in the fan inlet and makes condi-
tions more favorable for fan stall. Air flow control on the discharge (i.e.,
high pressure) side of the fan may alleviate a lot of flame instability problems.
The two-fan burner set-up was initially fired during the insulated chamber experi-
ments (Run 42, et seq.) and, to allow greater flexibility, this supercharged burner
configuration was maintained for the rest of the optimum burner/chamber matching
tests.
Insulated Combustion Chambers. The insulated chamber experiments involved
adding Pyroflex refractory fiber linings [l = 0.75 m (30 inches) to two 0.222 m
(8.75 inches) and 0.279 m (11 inches)] of the air-cooled chambers to produce 0.175
m (6.89 inches) and 0.222 m (8.75 inches) ID insulated chambers (inside wall
temperatures = 1500 to 1600 C). Additional baffles were added to the finned-
tube heat exchanger to maintain flue gas temperature within the 200 through 300 C
(~390 through 570 F) range. Flue gas composition data for the insulated 0.175 m
(6.89 inches) and 0.222 m (8.75 inches) diameter combustors are plotted in Fig. 24
and 25, respectively. The data plotted are very smooth and sequentially ordered
with no peculiarities. The unburned hydrocarbon emissions for these insulated
combustors are much improved over the air-cooled combustor results and do not
appear to be much of a problem for refractory-lines combustion chambers. The
emission improvement caused by adding insulation does not apply to the nitric
oxide emissions where the higher temperatures increased the flue gas concentra-
tions by a factor of 2 to 3. A quick tradeoff evaluation of best-operating-condi-
tions in the 0.40 m (16 inches) length insulated combustor suggested a 50-percent
excess air setting for the 0.175 m (6.89 inches) diameter combustor and a 25-
percent excess air setting for the larger, 0.222 m (8.75 inches) diameter combustor.
This corresponds to a a 2.5-to 3-percent furnace net efficiency advantage for the
larger-diameter chamber.
73
-------
30 r
too
2.00
1.50
STOlCHlONVtTRlC RRT1D
ISO
STOIC.HIO^AETRlC RATIO
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Figure 24, Cycle-Averaged Flue Gas Emissions, 1.0 ml/s (gph) Optimum Oil Burner in a 0.175 m (6.89
inches) Diameter, Insulated, Tunnel-Fired Combustor at Various Heat Exchanger Positions
-------
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Figure 25. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s Optimum Oil Burner in a
0.222 m Diameter, Insulated, Tunnel-Fired Combustor at Various
Heat Exchanger Positions
-------
Water-Cooled Combustion Chamber. The water-cooled combustor experiments in-
volved adding a 0.53 m (20 inches) long water jacket to the outside of the 0.222 m
(8.75 inches) diameter combustor. With the water-cooled, spiral-wound heat ex-
changer in place, the firebox region was enclosed completely in a water-cooled
environment having inside wall temperatures on the order of 150 to 200 C (320 to
392 F). Cyclical tests were conducted by initiating the burner firing when the
water in the jacket cooled down to 65 C (150 F) and maintaining burner-on for 10
minutes duration, which brought the water jacket temperature to about 90 to 95 C
(194 to 203 F). Flue gas temperatures ranged between 195 to 245 C (380 to 480 F),
a little lower than conditions obtained in the other types of chambers as a result
of rejecting about 1/5 to 1/4 of the heat released to the water-cooled chamber
walls.
In general, poor burner starting characteristics are associated with cold combustor
walls; a starting puff of high CO, smoke, and UHC concentrations can raise the
cycle-averaged emission values dramatically. Somewhat surprisingly, this problem
was not encountered in the tunnel-fired, water-cooled combustor tests at conditions
having less than 50-percent excess air.
Flue gas composition data for these water-cooled combustor tests are plotted in
Fig. 26. The shortest chamber [0.30 m (12 inches)] was the only configuration
which produced any significant amount of carbonaceous pollutant emissions; it also
produced the lowest concentration of nitric oxide. The 0.40, 0.50, and 0.75 m
(16, 20, and 30 inches) length combustors were very low in all carbonaceous pollut-
ant emissions and operated at as low as 12-percent excess air with no apparent
problems.
Side-Fired Orientation Results. The side-fired (perpendicular-port) burner orient-
ation was evaluated in Runs 82 through 154 (Appendix E) with the 1 ml/s (gph)
optimum burner fired in the same chambers and at much the same conditions that it
was tunnel-fired, so that directly comparable data would be obtained with the two
chamber types.
76
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Air-Cooled Combustion Chambers. The 0.279 m (11 inches) ID, air-cooled (unin-
sulated) combustor was tested at heat exchanger positions of 0.30, 0.40, 0.50, and
0.75 m (12, 16, 20, and 30 inches). The cycle-averaged (10 minutes on, 20 minutes
off) flue gas composition results are presented in Fig. 27. These data are "start-
limited" on the right extremity of the stoichiometric ratio range, as were tunnel-
fired, air-cooled chamber emissions data presented previously. A relatively higher
excess-air ignition capability (SR «1.40) was demonstrated over the equivalent
tunnel-fired configuration (SR « 1.2). The higher upper-excess-air limit may be
the result of an improvement in air/oil mixing effected by added turbulence gener-
ated by the crossflow entry of the side firing or to the supercharged combustion
air fan. Evidence of improved mixing can be seen in the lower levels of both car-
bon monoxide and unburned hydrocarbon concentrations (i.e., more complete combus-
tion) . The smoke emission levels remained essentially unaffected by the change in
burner/chamber orientation, probably because of the large chamber diameter that
places the opposing wall at a relatively great distance of 0.279 m (11 inches)
(approximately the length of the short chamber, L = 0.30 m (12 inches)). This
distant opposing wall and the large air-cooled surface area of the 0.279 m (11
inches) ID chamber are probably responsible for the unexpected absence of an in-
crease in nitric oxide emissions in the perpendicular port configuration.
The 0.222 m (8.75 inches) ID, air-cooled combustor was also fired, and the result-
ing flue gas composition data are shown in Fig. 28. Again, as in the 0.279 m (11
inches) ID chamber case, the side-fired configuration shows a higher excess-air
ignition capability and more complete combustion (lower CO and UHC levels) than
its tunnel-fired counterpart. Examination of smoke and nitric oxide emission
levels reveals the more typical characteristics of side-fired chambers, i.e., more
smoke and higher NO concentrations. The cycle-averaged Bacharach No. 1 smoke level
appears to be at about 20 to 25-percent excess air as compared to about 8- to 10-
percent excess air in the tunnel-fired chamber. The difference represents a 2- to
30-percent lower overall furnace thermal efficiency. The nitric oxide emission
profiles in Fig. 28 show a nominal NO level of 0.8 g/kg which is nearly twice as
high as was experienced in the air-cooled, tunnel-fired chamber. This factor of
two times higher NO emissions from side-firing then from tunnel-firing is consist-
ent with steady-state data obtained earlier with several different burners (Ref. 5).
78
-------
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STOlC-HiOMETRlC RATIO
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Figure 27 Cycle-Averaged Flue Gas Emissions, 1.0 ml/s (gph) Optimum Oil Burner in a 0.279 m
(11 inches) Diameter, Air-Cooled, Side-Fired Combustor at Various Heat Exchanger
Positions
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The fractional, air-cooled, side-fired combustor test series was conducted as was
its counterpart tunnel-fired combustor test series, in the 0.222 m (8.75 inches)
ID combustor with the heat exchanger (i.e., chamber length) set at 0.40 m (16
inches) and using a 0.30 m (12 inches) length of external insulation. The cycle-
averaged emissions data are presented in Fig. 29. The numerical label of each
curve represents the fraction uninsulated, with the insulated section being moved
away from the burner. The completely uninsulated curves are replots of the air-
cooled chamber data from Fig. 29. Comparison of Fig. 29 with Fig. 23 reveals the
same differences between the side- and tunnel-fired configurations as were previously
noted for the air-cooled chambers (i.e., side-firing produces lower CO and UHC,
higher smoke and NO, and, additionally, has better start capabilities). Further,
partial insulation displayed about the same degree of apparent ineffectiveness in
altering emission levels with the side-fired burner orientation as with tunnel-
firing. This appearance is deceiving, however, because the emission data for the
insulated chamber (zero fraction uninsulated) are not included on Fig. 23 and 29.
Comparison of the NO data on Fig. 23 and 29 with those for the insulated 0.40 m
(16 inches) chamber length in Fig. 24 and 30, respectively, shows that removing the
insulation from the first quarter of the chamber did have a large beneficial effect
even though little further benefit resulted from removing even more insulation.
Insulation Combustion Chambers. The 0.222 (8.75 inches) ID, insulated, side-
fired combustor was tested at combustion chamber lengths (heat exchanger positions)
of 0.30, 0.40, 0.50, and 0.75 m (12, 16, 20, and 30 inches). The cycle-averaged
(10 minutes on, 20 minutes off) flue gas composition results are plotted in Fig. 30.
An evaluation of the data reveals that unburned hydrocarbons (UHC) concentrations
were very low for all cases, and the carbon monoxide (CO) concentrations were also
very low, except for the 0.30 m (12 inches) configuration at low-excess-air levels.
A comparison of UHC and CO concentrations from other configurations of the same
diameter shows no differences in UHC emissions and some small differences in CO
emission at SR <1.3 in the shorter chamber lengths [<0.4 m (16 inches)]. The smoke
emission levels for the 0.50 and 0.75 m (20 and 30 inches) chamber lengths are
comparable to the levels obtained in the tunnel-fired configuration, while the
shorter 0.30 and 0.40 m (12 and 16 inches) side-fired chambers show greater than
No. 1 smoke at SR <1.30. The cyclic nitric oxide emissions of this chamber were
81
-------
30
00
N>
i
10
ETTT
6.VOHE £ 1
V4 (FRACTION UNINSULATED)
ISO
1.00
LID 1.20 120 1.4O
STOIC.HIOMETRIC RATIO
L50
15 r
1 ID
§
£
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o
oc
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1.10 1.ZO 1.30 1AO 1.5O
ST01CH10METR1C.
1.10 120 1.30 1.4O LSO
STOICHIQJAETRIC URT10
Figure 29. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s Optimum Oil Burner in a 0.222 m (8.75
inches) Diameter, Partially Insulated, Side-Fired Combustor at a 0.40 m (16 inches)
Heat Exchanger Position
-------
30 r
00
I
to
I
ui
Q
g
K
Loo
1.00
BAtHAKACU WOKE B 1
1VO l.ZO 1.30 L40
STQlCHlOMtTRtt RATIO
1.50
.30m (HEkT ncH*N*n MIITUM)
1.10 150 130 1.4O
STOTC.HIOMETR1C RATIO
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ST01CH10METR1L RATIO
1.50
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i.oo
tlO L20 1.50 1.40
STOItHHMETBlC WkTTD
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Figure 30. Cycle-Averaged Flue Gas Emissions, 1.0 ml/s Optimum Oil Burner in a
0.222m Diameter, Insulated, Side-Fired Combustor at Various Heat
Exchanger Positions
-------
a factor of two or more greater than any other configuration of the 0.222 m (8.7
inches) ID chamber, reconfirming the continuous-firing results reported in Ref. 5.
The 0.175 m (6.9 inches) ID insulated combustor was also tested at four chamber
lengths, and the resulting flue gas composition data are presented in Appendix C,
Runs 131 through 142. The UHC emission levels again do not appear to be of any
concern for any of the four chamber lengths tested. The CO levels, however, show
some noticeable concentrations (>1.0 g/kg) at chamber lengths <0.40 m (16 inches).
The smoke concentrations were relatively high in this small chamber but, surpris-
ingly, were comparable to those of the equivalent size tunnel-fired chamber. The
smoke emissions showed a noticeable sensitivity to combustion chamber length, in-
creasing with decreasing chamber length. The NO concentrations were again a factor
of two higher than other comparable diameter configurations tested.
Water-Cooled Combustion Chamber. A 0.50 m (20 inches) long water jacket was
added to the 0.222 m (8.7 inches) ID uninsulated combustor for the water-cooled
combustor experiments. The cycle-averaged flue gas composition results are pre-
sented in Appendix C, Runs 143 through 154. The side-fired configuration showed
some improvement in CO and UHC concentration at SR >1.20 over the tunnel-fired
results, probably because of the better mixing in the combustion zone provided by
the crossflow entry. The smoke emission levels were similar in the two chambers,
with the exception of the short 0.30 m (12 inches) chamber that was unacceptably
smoky in the side-fired configuration. The NO concentrations showed no sensitivity
to orientation with water-cooled walls. The nominal NO level remained at about
0.6 to 0.7 g/kg of fuel burned.
Several months later, a brief series of tests was made with the 1 ml/s (gph) optimum
burner side-fired in a water-cooled 0.279 m (11 inches) chamber. The larger-
diameter chamber was tested to confirm the hypothesis, discussed in the following
subsection, that using an unusually large combustion chamber diameter and a water-
cooled firebox section would greatly reduce NO emissions. The data are listed in
Appendix C, Runs 427 through 432. The cycle-averaged NO concentrations were very
low indeed. The other pollutant emissions were acceptabley low, with the possible
exception of CO, which indicated that the firebox might also have to be lengthened.
84
-------
Discussion: Optimized Conventional Burner/Combustor Matching
The experimental data from firing the 1 ml/s (gph) optimum burner in the variety
of chambers were evaluated to determine a preferred combustor configuration, the
size and cooling method, and whether that preferred burner/chamber combination
can meet the objectives listed in the Introduction.
Determining whether the objectives can be met necessitates defining acceptable
emission levels for CO, UHC, and smoke. Values were selected in conformity with
average values reported in the field survey of Ref. 4:
Emiss ions :
As Found:
Tuned:
co2,
percent
-7-9
~8.1
Smoke,
Bacharach Number
3.2
1.2
g/kg Fue
CO
1.05
0.65
UHC
0.10
0.06
NO
2.6
2.7
These are quite lean conditions (8.0-percent C0_ corresponds to 90-percent excess
air and net thermal efficiencies of about 70 percent), which normally produce quite
low values of carbonaceous emissions. In view of the other goal of increasing
efficiency by 10 or more percent, partially achievable by reducing excess air,
and the modest differences between "as found" and "tuned" CO and UHC, acceptable
levels were selected to be: CO <1 g/kg; UHC <0.1 g/kg; smoke
-------
TABLE 8. SUMMARY OF RESULTS FROM 1 ML/S (GPH) OPTIMUM BURNER/COMBUSTION
CHAMBER MATCHING EXPERIMENTS
00
Chs-ber Desicn Attributes
Configuration
Tunnel -Fired
(Coaxial )
Side-Fired
(Perpendicular
Port)
Cooling Method
Air-Cooled
(Startable over i
very narrow S.R.
range)
Water-Cool ed
Insulated
Air-Cooled
(Forced Draft)**
Water- Coo led
Insulated
I.D., m
0.162
0.222
0.2.79
0.222
0.175
0.22
0.162
0.222
0.279
0.250
0.222
0.279
0.175
0.22
Requirements For Acceptable (or
Minimum) Emissions
CO & UHC-
Rough-Burnlnc
Lc -0.5m
Lc S0.75m
LC >0.5m
Lc *0.5m
Lc *0.5m
Rough- Burn in
Lc*0.5m
L *0,4m &
S?R.-1.15
Lc >. 0.5m
Lc*0.5m
Lc >. 0.5m
Lc-0.4m
Lc*0.5m
Sir.oke (ilio.i)
Throughout Operati
S.R. *1.05
S.R.*1.1 or
Lc>0.5m
S.R. s 1.15
Lr-0.75m
of S.R. >1.3
Lr*0.5m &
S?R.>1.15
g Throughout Opera
Lr ?0.5m
STR. ^ 1.15
L *045m &
S.R.-1.08
L^O.Bmi S.R.>1.05
L.^0.4 &
S?R.^1.2
L^O.5 & S.R.>1.1
L ^0.5 &
S?R.^.1.25
L ^O.Sra &
SCR.*1.15
NO
ng Range
Lc=0.4m
LcsC.4-0.5m
Short LC'S Best
.Short Lc'-s Best
Short Lc's Best
ting Range
L =0.75m slight
favor
Short Lc's Best
Short LC'S Best
Short L 's Best
Short LC'S Best
Short LC'S Best
Data Mixed
IiO,g/kg fuel
(?L =0.5 m &
1.25 S.R.
(0.6)*
(0.4)
0.7
1.3
0.9
0.9
0.6
0.6
0.7
0.4
2.0
1.9
**
*Values enclosed in parentheses obtained by extrapolation outside the star table range
*Experiments described and discussed in Appendix F
-------
Combustor design criteria that can be derived from Table 8 (and the larger body
of data upon which it is based) are subdivided into categories of burner orienta-
tion, chamber size, and cooling medium in the following paragraphs.
Burner Orientation. Combustion chamber design requirements for ensuring acceptable
levels of CO, UHC, and smoke emissions were found to be quite similar for both the
side- and tunnel-fired burner orientations. That observation held for NO emissions
from water-cooled combustors as well. However, NO production in refractory-lined
and air-cooled chambers was about 1-1/2 to 2 times as high in the side-fired or-
ientation as in the tunnel-fired. Ostensibly, this latter phenomenon resulted
from substantially longer gas residence times at high temperatures in the stronger
and more complicated eddies of a side-fired chamber.
Combustion Chamber Size. Chambers smaller than about 0.20 m (8 inches) ID must
be refractory-lined to avoid operating with unacceptable combustion roughness.
Operation of all designs was acceptable in this regard when their inside chamber
diameter was 0.222 m (8.75 inches) or larger.
Larger-diameter chambers, with or without insulation, tend to require longer
chamber lengths to achieve comparable levels of carbonaceous pollutants and,
concurrently, to produce substantially lower levels of NO. Both phenomena are
undoubtedly linked to the ingestion of recirculation combustion gases into the
flame zone. For a given burner, larger and stronger recirculation eddies can be
established in bigger chambers, thereby reducing the combustion intensity some-
what and lowering the rates of burnout of carbonaceous species. Also, larger-
diameter chambers have greater wall areas for convective and radiant transmission
of heat from the flame zone, which reduces peak flame temperatures somewhat. In-
creased average gas residence times should produce opposing trends for both NO
production and carbon burnout, which could help account for some anomalies in the
data. Further, the steady-state recirculation and radiation effects may be par-
tially masked by starting transient effects, particularly with respect to smoke
and UHC data.
87
-------
Short combustion chambers generally were more faborable for low NO but, if they
are too short, the fuel may not be completely burned before combustion reactions
are quenched in the heat exchanger. A length of 0.5 m (20 inches) is a suitable
compromise for most of the chamber designs tested. Slightly shorter chambers
might be approproate if refractory linings were to be used.
Chamber Cooling Medium. Refractory-lined combustion chambers generally exhibited
less combustion roughness and less sensitivity to starting conditions but also pro-
duced higher NO concentrations than did cooled-wall combustors. The air-cooled,
side-fired configuration had better starting characteristics (but higher NO) than
did the air-cooled, tunnel-fired chamber. Water-cooling appeared to be preferable
to air-cooling, partially because of lower side-fired NO levels, but also because
of smoother starting and more consistent CO and NO emission results. The water-
cooled chambers in shorter lengths, however, were prone to produce excessive GO.
Tests with partially insulated, partially air-cooled chambers were not found to
differ markedly from the fully air-cooled configuration.
88
-------
COMBUSTION GAS RECIRCULATION (CGR) BURNER STUDY
CGR Burner
Circulation patterns are normally established within an operating combustion
chamber. The main driving force is momentum exchange between the relatively
high-velocity reactants emanating from burners and the lower-velocity gases in
the chamber. Mixing of the ingested gases into the reactants as they burn lowers
the reaction intensity and stretches the reaction zone. The impact of this inter-
nal combustion gas recirculation on NO formation covers a broad range, from nega-
tive to positive. In well-insulated, near-adiabatic combustors, the main effect
of internal recirculation is to increase the average gas residence time; usually
this results in higher NO concentrations via a closer approach to thermodynamic
equilibrium. In cooled-wall combustors, recirculation eddies embedded within the
flow downstream of a burner may similarly increase NO production, modulated only
by radiant losses to the walls. Circulation currents that flow along the combus-
tion chamber walls, however, are also convectively cooled. If sufficient heat
is lost to the walls, these partially cooled recirculation gases drawn into the
flame zone act as an inert diluent, reducing the combustion intensity and peak
temperature, and inhibiting the formation of NO enough to more than offset the
effects of increased average residence time.
Internal combustion gas recirculation patterns and their strengths are largely
determined by the combination of firebox design, burner design, and burner
placement. Some burners have been designed intentionally to promote this kind
of recirculation (see Ref. 9), but they may be rather restricted in application
to particular firebox designs and burner positions that do not inhibit the recir-
culation. Conceptually, such dependence can be reduced by forcing the recircula-
tion of partially cooled combustion gases (e.g., by withdrawing gases from the
combustor and recirculating them to and mixing them with the burner air supply).
This recirculation is difficult to accomplish because the chamber gases are too
hot for mechanical air blowers. Two viable approaches that have been tried are to
89
-------
use the momentum of the combustion air of the burner to aspirate gases from the
combustor (Ref. 10) and to dilute the hot gases with fresh combustion air before
the mixture is supplied to the combustion air fan (Ref. 5). In both cases, a
larger, more powerful fan is required in order to force the recirculation of the
extra gases.
A burner embodying the latter approach, of diluting combustion gases withdrawn
from the combustor with fresh combustion air, was studied in this program. The
burner concept is shown schematically in Fig. 31 to have an air/combustion gas-
mixing plenum surrounding the burner blast tube that was exhausted to the suction
side of the combustion air fan of the burner. The CGR burner had the capability
of drawing combustion gases from locations near its center (B), near the chamber
wall (A), or from further downstream in the combustion zone (C). The burner was
designed to fit in the air-cooled 0.22 m (8.75 inches) diameter combustor in
either the tunnel-or side-fired orientations.
The burner head used the variable swirl vane assembly from the previous Rocketdyne
burner studies (Ref. 5) and replaceable fixed-diameter choke plates so that these
parameters could be optionized as appropriate with CGR variations.
Figure 32 is a photograph of the CGR burner assembly, showing the CGR manifold
assembled around the variable swirl vane, versatile burner head assembly. The
photograph does not show a 0.10 m (4 inches) diameter insulated ducting between
the CGR manifold outlet and the blower inlet. The electric motor was remounted
on 0.015 m (0.6 inches) insulating standoffs. The safety circuitry and spark
transformer were removed from the CGR burner body to avoid anticipated "over-temp"
problems. Also, the capacity of the combustion air fan was increased about 20
percent by adding a metal lip, inside the fan housing, to reduce the squirrel-
cage impeller clearance to 0.002 m (1/8 inch). [The no-flow outlet pressure of
2 2
the fan went from 310 N/m (1.25 inches of water column) to 370 N/m (1.50 inches
of water column)].
90
-------
-------
Thermocouple
Mixed Gas Sample Line
Air Inlets
CGR "B
Removable
Choke Plate
Variable
Swirl Vanes
CGR
Inlet "A
CGR
Inlet "C"
Mixed Gas Path
(Ducting Not Shown)
Mixed Gas
Flow Control Valve
5AA34-12/17/74-S1
Figure 32. Photograph of Combustion Gas Recirculation Burner
-------
Experimental Test Matrix
A matrix of tests was planned as tabulated in Table 9 for evaluating the emissions
performance of the CGR burner in the tunnel-fired combustor orientation. Varia-
tions in the magnitude and source of CGR, excess air level, burner choke, and
swirl were planned predominantly at steady-state conditions. Subsequently,
with optimum values for those parameters, it was planned to vary the oil nozzle
and supply pressure, and the chamber length and diameter during cylical testing.
Experimental Test Results
The CGR burner firings actually conducted, are summarized in Table 10. The flue
gas and recirculated gas composition data are tabulated by run number in Appendix D.
After 17 firings (Runs 155 through 171) were conducted, it became apparent that
the CO and UHC concentrations were significantly higher than those from conven-
tional burners, and that the target values would not be obtained with the configu-
rations and conditions specified by the Table 9 test matrix. Therefore, an
exploratory firing series was initiated to investigate methods of improving the
CO and UHC emissions. A baffle was added to the tunnel-fired chamber (Runs 172
through 183) 0.23 m (8.75 inches) from the burner head, hopefully to improve re-
circulation eddies and reflect radiant heat back into the combustion zone. The
result was an improvement in the system operational characteristics, but no signi-
ficant improvement in emissions. The side-fired burner orientation was tried both
with a 1.0 60°-A nozzle and with a 0.75 60°-A nozzle (to increase the chamber
volume/burner output ratio, i.e., a larger effective chamber diameter). The CO
and UHC concentrations were not improved appreciably. Further experiments were
conducted using the tunnel-fired chamber with various changes to burner, combustor,
and run conditions. A 0.15 m (7.5 inches) length of Pyroflex insulation (added
to retain heat in the combustion zone), higher excess air settings, a 0.02 m (1
inch) choke plate extension, and gas recirculation from 0.25 m (10 inches)
93
-------
TABLE 9. COMBUSTION GAS RECIRCULATION BURNER TEST MATRIX
«o
CGR,
Percent
10
20
k
30
1
1
Optimum
Excess Air,
Percent
10
5
15
20
25
30
35
Optimum
Optimum
10,20,40,80
Optimum
5,10,20,30
Optimum
Choke
Diameter,
cm
2.5
3.8
5.1
2.5
3.8
5.1
6.4
2.5
3.8
5.1
6.4
Optimum
Air Swirl
Angle,
Degrees
10,20,30,40
Optimum
i
i
Chamber
Diameter,
m
0.222
i
0.279
1
I
Optimum
Chamber
Length,
m
0.40
i
0.30
0.50
0.75
12
16
20
30
Optimum
Oil
Nozzle
1.0-90A
0.6-90A
0.6-90A
1.1-90A
1.1-90A
1.0- 90A
Oil
Pressure
xlQ5 N/M2
6.89
20.68
20.68
5.52
5.52
6.89
Data Type
Equilibrium
Transient
Equilibrium
Equilibrium
Transient
Chamber
Wall
Air-Cooled
-------
TABLE 10. SUMMARY OF COMBUSTION GAS RECIRCULATION BURNER, HOT FIRE EXPERIMENTS
10
en
Run
No.
155-171
172-183
184-186
187-190
191, 192
193-199
200-212
213-223
224-230
231-238
239-247
248, 249
Transient
CGR
Location
A
B
v
A
A
A
A
A
A
C
0.25 m
shroud
C
0.25 m
shroud
A
A
CGR
Percent
10-30
10-30
15-40
15-25
44, 40
15-45
10-35
10-30
15-45
15-30
10-30
20, 10
Excess
Air,
Percent
5-35
5-15
15-40
1-20
55, 40
15-110
15-70
15-80
15-50
15-50
10-50
40, 25
Choke
Diameter,
m
0.042
0.042
0.051
0.051
0.02m Ext.
0.051
0.051
Swirl
Angle,
Degrees
10-55
10-40
40
40
0
40
40
Oil
Nozzle
1.0-60-A
1.0-60-A
0.75-60-A
1.0-60-A
1.0-60-A
Combustion Chamber
Dia.,
m
.22
.2
1
2
Length,
m
.40
.4
0
Remarks
Air-Cooled (A/C) Tunnel-
Fired
A/C, Tunnel -Fired,
Baffle at 0.23 m
A/C, Side-Fired
A/C, Side-Fired
A/C, Tunnel -Fired,
Baffle at 0.23 m
A/C, Tunnel -Fired,
Baffle at 0.23 m
0.15 m Insul.
A/C, Tunnel -Fired,
Baffle at 0.23 m
A/C, Tunnel-Fired,
Baffle at 0.23 m
A/C, Tunnel -Fired
Baffle at 0.40 m
A/C, Tunnel-Fired,
Baffle at 0.40 m
Water-Cool ed,
Side-Fired
Water-Cooled,
Side-Fired
-------
downstream (position C)* all resulted in CO and UHC emissions higher than
the target values. A water-cooled, side-fired combustor produced very low
NO emissions but showed no improvement in CO and UHC.
Only two cylical tests were conducted with the CGR burner. Minor ignition prob-
lems were encountered because of the high-excess air starts (because of the initial
absence of combustion gas) but no start-to-blue-flame transition problems were
noted.
It was found that the steady-state operation of the CGR burner gave acceptably
low exhaust gas NO concentrations but generally had unacceptably high CO and UHC
concentrations. A few conditions did have low enough CO, but the hydrocarbons
remained high. Examination of the data strongly suggested that the UHC readings
were probably erroneously high. Careful examination of the data revealed the
following factors:
1. Among the various combustion gas sample analyses performed in the
burner/furnace test laboratory, the UHC measurement system has the
slowest response. During previous cyclical furnace tests in which
appreciable UHC levels (~30 to 50 ppm) were measured, recovery to a
precycle zero level took much of the 20-minute off-time of a 10-min on/20-
min off cycle. With continued testing at these levels of UHC emission,
precycle zeros frequently were not recovered, i.e., the system appeared
to be retaining UHC which was desorbed quite slowly, resulting in
apparent zero shifts.
2. In the CGR burner tests, samples were fed into the entire sample system
alternately from the flue gases and from the mixture of CGR and combus-
tion air within the burner. With several configurations and many
*Rather than using tubular extensions from position A holes, as illustrated in
Fig. 31 and 32, position C experiments employed a 0.25 m (10, inches) long, 0.15 m
(7.5 inches) ID stainless-steel shroud around the flame zone. This shroud was
chosen to give apseudo-simulation of the CGR burner reported by Cooper and
Marek (Ref. 10).
96
-------
operating conditions, the UHC levels in the recirculated combustion
gases were extremely high so that the concentration in the mixed burner
gases exceeded the meter range (>3300 ppm). Thus, particularly during
steady-state testing, there were ample opportunities for the UHC sample
system to absorb substantial quantities of UHC that would be slowly
desorbed into subsequent samples, causing them to produce erroneously
high UHC readings.
3. The probable magnitude of the error was estimated, from previous experi-
ence and from a few tests of the zero shift when an air purge was passed
through the system, to be on the order of 20 to 30 ppm, corresponding
to about 0.22 to 0.33 g/kg fuel. This is on the order of one-half the
UHC levels recorded for those CGR burner conditions where the CO, NO, and
smoke were acceptable; the indicated residual UHC concentrations are
still higher than the target acceptable value of 0.1 g/kg.
It was decided that, in future testing, UHC zero shifts of this sort should be
quantified by admitting air to the sample line after the analysis of each sample
having high UHC. However, for these CGR burner experiments, even if the UHC
readings were totally wrong and this pollutant had had acceptable concentrations
in all of the tests, only some minimal design and operating conditions appeared
where CO, smoke, NO and operability were all acceptable. Thus, it was decided
to forego further testing of this concept, since it appeared to generate more
problems than solutions.
FLUE GAS RECIRCULATION (FGR) BURNER STUDY
External recirculation of combustion gases is practical if they are obtained
downstream of the furnace heat exchanger, i.e., from the exhaust flue. Mixed
with the incoming combustion air, recirculated flue gases are a more effective
diluent than an equal quantity of CGR because they are cooler. Peak flame zone
temperatures are reduced by 300 to 400 C by 30 percent FGR, and NO emissions from
burning distillate fuel oils have been lowered by as much as 85 percent. As
97
-------
the level of FGR is increased, two phenomena are commonly observed: a diminish-
ing effect occurs on NO level, and increasing difficulties develop in maintaining
flame stability. The guiding principles for optimizing the level of FGR for given
stoichiometric ratio and flue gas temperature conditions have been delineated in
Ref. 11, but optimization of specific units currently depends more upon empirical
testing than on theoretical analysis.
FGR Burner
The basic FGR burner is a modification of the 1-gph versatile burner (Ref. 5).
The air inlet was modified to provide a mixture of FGR to the air on the suction
side of the squirrel cage fan, and the variable choke plate was replaced by an
interchangeable, fixed-diameter choke plate. The flue gas adapter was designed
to be mated to a 0.1 m (4 inches) diameter galvanized steel flue gas return
duct (Fig. 33).
Experimental Test Results
In view of the experiences with the CGR burner, a preconceived FGR burner test
matrix was set aside and exploratory tests were conducted in various combustion
chamber configurations to determine a suitable combustor to use for performing
the FGR burner/chamber optimization. Flue gas composition data are tabulated
in Appendix E.
Runs 250 to 321 are the preliminary steady-state chamber configuration experi-
ments involving 0.222 m (8.75 inches) diameter chambers which were: (1) side-
fired, water-cooled; (2) side-fired, insulated; and (3) tunnel-fired,insulated.
Tests of the water-cooled chamber (Runs 250 through 285) produced very low NO
concentrations (0.2 through 0.4 g/kg), well below the 0.5 g/kg of fuel burned
target value. However, the CO and UHC emission characteristics were very poor,
and therefore, the side-fired, water-cooled combustion configuration was deemed
as an unlikely candidate for FGR optimization.
98
-------
Primary Gas
Thermocouple
Primary Gas
Sampling Tube
Variable
Swirl Vane
Head
Removable
Choke Plate
,
Flue Gas
Control
\
Air Inlet
Control
«
Flue Gas
Thermocouple
Port
Figure 33.
50P21-2/7/75-S1
Photograph of the Flue Gas Recirculation Research Burner
-------
The insulated, side-fired combustor was tested (Runs 286 through 299) with two
choke plates [0.051 m and 0.072 m (2.0 inches and 2.9 inches)]. Improvement in
UHC emissions was observed with the 0.051 m (2.0 inches) diameter choke, but also
experienced was an unacceptable increase in NO emissions (>0.5 g/kg). The 0.072 m
(2.9 inches) diameter choke proved unacceptable in smoke, CO, and UHC at lower
air settings.
The insulated combustor was then reconfigured into a tunnel-fired configuration,
and the resulting pollutant emissions concentrations were favorable. Three chamber
lengths [0.40, 0.50, and 0.75 m. (16 inches, 20 inches, and 30 inches)] were tested
(Runs 300 through 343) and resulted in the 0.50 m (20 inches) length being selected
for further experimentation. Runs 322 through 343 represent swirler vane varia-
tion results with a 0.051 m (2.0 inches) diameter choke plate. Runs 344 through
357 produced similar data with a smaller 0.042 m (1.5 inches) diameter choke plate.
In the latter set of experiments, the 10-degree swirl vane angle tests were eli-
minated because of the poor CO, smoke, and operational characteristics observed
in the previous set with the 0.051 m (2.0 inches) choke.
An oil spray angle variation from 60 to 90 degrees (Runs 358 through 361) resulted
in a slight improvement in NO concentration, but with a significant degradation
of operational characteristics. Rough combustion was encountered at all settings
below 20-percent excess air.
Hot-fire testing resumed with the 60-degree nozzle and a choke plate change to
0.064 m (2.5 inches) diameter (Runs 362 through 381). This large-diameter-choke-
plate burner configuration was smoky and produced high NO levels at the lower
excess air settings (~10 percent).
These FGR parametric variation experiments converged upon a burner configuration
having a 0.051 m (2.0 inches) diameter choke plate, 40-degree swirl vane angle,
60-degree oil nozzle and 30-percent flue gas, recirculation in an insulated,
tunnel-fired combustor at least 0.50 m (20 inches) in length. Runs 382 through
100
-------
387 were made to recheck this preliminary, steady-state optimization of the FGR
burner/chamber configuration before proceeding to the cyclical testing portion
of the optimization testing. This brief series of tests began to display a
characteristic that had to be corrected before the optimization could be completed:
rough and noisy combustion prevented operation at less than 18-percent excess air.
Since these tests were conducted on a very windy day with strong gusting, the poor
operation was attributed to wind effects on firebox draft. However, it persisted
on the following day, which was calm, and exploratory cyclical testing revealed no
regions of acceptably quiet operation where data-taking runs could be made.
Inspection of the burner and combustion chamber revealed two conditions thought to
be possible contributors to noisy operation; both were corrected. The first was
that a piece of aluminum foil tape, used to thermally protect internal 110-VAC
wiring in the burner, had come loose and was partially obstructing the flow around
the squirrel-cage impeller of the fan. The second was that the Pyroflex liner
of the chamber was partially separated into irregular and displaced layers near
the heat exchanger end, presumably as a result of the many configuration changes
that had been made. The insulation was removed and the chamber relined. However,
rather than installing three layers of Pyroflex to form a 0.22 m (8.75 inches) ID,
a single layer was initially used to form a 0.26 m (10-5 inches) ID chamber to
test before completing the relining.
The 0.26 m (10.5 inches) diameter chamber was tested cyclically in both the tunnel-
and side-fired configurations (Tests 388 through 382). Operation was quite satis-
factory, with smooth starts and smooth, quiet burning to as low as 10-percent
excess air. Carbonaceous pollutant levels were below the target levels at low-
excess-air conditions, but a starting puff of smoke pushed the cycle-averaged
UHC readings above 0.1 g/kg at greater than about 30-percent excess air. Cycle-
averaged NO concentrations exceeded the 0.5 g/kg target value slightly; the
increase over the previous steady-state levels was thought to result either from
the 18-percent increase in chamber diameter or from the spark igniter having
been left on for the entire bumer-on time. Interestingly, the side-fired con-
figuration did not produce appreciably higher NO than the tunnel-fired.
101
-------
For tests 393 onward, the additional refractory lining was installed to reform
the 0.22 m (8.75 inches) diameter chamber, which was then tunnel-fired. Tests
393 and 394 were made to provide a steady-state comparison of NO emissions with
the spark igniter left on versus turning it off after ignition was achieved.
There were no measurable changes in the flue gas composition.
The optimized FGR burner was then cyclically fired and the data (Runs 396 through
398) show some marginal emission level characteristics. The NO concentration
was slightly higher than the targeted 0.50 g/kg, and the CO emission at less than
16-percent excess air began to show some unacceptable concentrations, mainly
because of the "start spike" at ignition. Recirculating gas burners are charac-
teristically poor starters as a result of the high excess air starts when cold
air instead of combustion gas is available in the gas recirculation ducts.
The FGR burner was start-limited at >35-percent excess air, but it displayed no
ignition nor transition difficulties at the lower air settings.
Runs 399 through 403 represent a chamber length variation; the heat exchanger was
moved to the 0.75 m (30 inches) position in the 0.22 m (8.75 inches) ID combustor.
Comparison of these 0.75 m (30 inches) chamber length data with the 0.50 m (20
inches) data shows the 0.75m(30 inches) chamber to be slightly better in CO emis-
sions, but also slightly worse in NO concentrations (from ~0.58 g/kg to -0.61 g/kg
at 30-percent FGR). Although the 0.50 m (20 inches) length chamber did show some
increase in CO as stoichiometric ratio was lowered to about 15-percent excess air
(mainly attributed to "spikes" in the FGR starts), its slightly lower NO levels
makes it the preferred configuration over the 0.75 m (30 inches) chamber.
The nominal cycle-averaged NO emission level of 0.58 g/kg of fuel burned at 30-
percent FGR is slightly higher than the 0.50 g/kg target value. If this target
value is to be strictly adhered to, then the FGR system would require more re-
circulated gases (probably 40-to 45-percent FGR) with electrical or mechanical
compensations (e.g., FGR control valve and programmed start cycle) to produce
102
-------
"smoother" (no emission spikes) ignition characteristics. The oil burner/furnace
industry would be opposed to any added "complexities" and, therefore, the commer-
cial success of a residential-size FGR system would have to be supported by signi-
ficant, saleable advantages and/or required to satisfy stringent pollutant emis-
sion standards.
So that the effect of changes in nominal oil spray dropsize may be studied, oil
supply pressure variations were tested in Runs 404 through 408. Runs 404 and 405
2
were fired with a 1.25-gph, 60-degree A nozzle at 551 kN/m (80 psig) oil supply
pressure, and Runs 406 through 408 were fired with a 0.75 gph, 60-degree A nozzle
Lgr
2
at 2137 kN/m (310 psig). The resulting plllutant emission data show no signi-
ficant differences from the typical 1.00-gph, 60-degree A nozzle at 689 kN/m
(100 psig) oil supply pressure; however, slight operational differences were noted.
2
The 551 kN/m (i.e., larger dropsize) firings showed some combustion roughness (noisy),
2
while the 2137 kN/m firings appeared to light-off more smoothly. These results
show no reason for the development of a different type of oil spray system.
The sensitivity of the optimum 1 ml/s (gph) FGR system configuration [0.051 m
(2.0 inches) choke, 40-degree swirl burner head; 0.22 m (8.75 inches) ID x 0.50 m
(20 inches) long, tunnel-fired insulated combust'or] to changes in firing rate were
intended to be investigated in Runs 409 through 412. However, Runs 409 through
2
420 were found to have been fired inadvertently at 551 kN/m (80 psig) oil supply
pressure. (Adjustments made to the pump system for a previous test series had
not been reset.) Thus several data points were rerun later, so the 0.68 ml/s (gph)
tests (Runs 409 and 410) should be compared with the later 1.21 ml/s (gph) tests
(Runs 425 and 426). The lower 0.68 ml/s (gph) firing rate (as compared to the
0.98 ml/s (gph) optimized firing rate) resulted in increases in all carbonaceous
pollutants. This result has been seen before with oversized burner choke con-
figurations and is caused by lowering the firing rate with fixed burner geometry.
Similarly, the 0.68 ml/s (gph) burner required more excess air for smoke-free
operation. This additional air was probably needed to regain the exit air velocity
(i.e., mixing energy) that was lost by the increase in effective choke diameter
103
-------
brought about by the lower firing rate. The higher 1.21 ml/s (gph) firing rate
(Runs 425 and 426) showed no significant effect upon the pollutant emission
levels; however, it appeared to operate more noisily (combustion roughness) than
the nominal 1 ml/s (gph) firing rate. The carbonaceous pollutant concentrations
were all below the allowable "target" values, and the NO concentration, as expected
from the reduced effective combustor size, increased slightly [from 0.58 g/kg at
0.98 ml/s (gph) to 0.64 g/kg at 1.21 ml/s (gph)]. The optimized FGR burner system
appears to be no more sensitive to firing rate changes than are the conventional
burner systems.
A pair of tests (Runs 413 and 414) was conducted to examine whether turning the
spark igniter off after ignition would lower the cycle-averaged NO emission
from the FGR burner. These two runs were also mistakenly fired at 551 kN/m (80
psig); however, since no operational difficulties were noted, and the resulting
emission levels were respectably low, the results were accepted as representative
data. Run 413 is the "igniter-off" firing with the ignition spark turned off
approximately 10 seconds after burner start, which is similar to an "interrupted
spark" -type burner control.
Run 414 is the comparative "continuous spark" firing; there were no apparent
differences in flue (and recirculated) gas compositions between the two runs.
This lack of an effect on cycle-averaged NO emissions is consistent with that
observed previously in steady-state igniter on/off firings. Although this
evidence points in the direction of a "no effect from igniter" conclusion, experi-
ence gained in an earlier oil burner emissions study (Ref. 5) restrains the adop-
tion of any "blanket" conclusion. It was seen in those earlier experiments that
the igniter-off effect was significant with certain burners and insignificant
with other burners. The FGR burner tested in the current experiments is of the
mechanically pumped, forced-recirculation type, and it has shown no transition
phase (ignition-to-steady-state) difficulties. The air-aspirated, induced-
recirculation-type burners, which are dependent upon gas temperatures, densities,
and resulting velocities may be greatly affected (both in operation and emissions)
by an igniter-off condition during their difficult transition stage. Therefore,
only a tentative conclusion can be made, restricting this "no effect from igniter"
result to this specific (or specific type) FGR burner.
104
-------
Oil furnace manufacturers expressed some concern about the quality of fuel oil
that might be available during periods of petroleum supply crisis. Therefore,
a short series of three FGR firings (Runs 415 through 417) was conducted with
the quality of the No. 2 fuel oil degraded by a 20-percent volumetric (22-percent
weight) contamination with No. 6 "residual" fuel oil. The chemical compositions
of both No. 2 and No. 6 fuel oils are listed in Table 11, together with the calcu-
lated (linear averages) composition and properties of the 20/80 mixture (22/78
by weight) of the two oils.
With the oil mixture, significant increases in all of the flue gas pollutant
concentrations were observed. The carbonaceous pollutant levels may have been
2
influenced upward slightly by the inadvertently low 551 kN/m oil pressure.
However, the No. 2 fuel oil tests immediately preceding (Runs 413 through 414)
were also fired at the lower oil pressure, and the carbonaceous pollutants were
well within acceptable limits. The increase in fuel nitrogen, from No. 2 oil to
the mixed oil, was a factor of about nine. The corresponding increase in flue
gas NO was a factor of about four (0.53 g/kg to 2.21 g/kg at 10-percent excess
air). From previous experimental studies of fuel nitrogen conversion to NOX (see
Ref. 3), it may be assumed that about 85 percent of low concentrations of fuel
*
nitrogen will show up as flue gas NO. For No. 2 fuel oil, then, flue gas NO
from fuel nitrogen was estimated to be about 0.16 g/kg fuel, leaving a balance
of 0.37 g/kg thermal NO. Similar conversion of the mixed oils fuel nitrogen
was estimated to produce about 1.46 g/kg fuel NO, leaving a balance of 0.75 g/kg
NO attributable to thermal formation. This apparent doubling of thermal NO may
have been caused by reduced radiation from the flame zone attending increased
combustion gas capacity as indicated by increased smokiness.
Earlier experiments with a slightly larger 0.26 m (10.5 inches) ID, insulated,
side-fired combustor rekindled some optimism for that configuration. Therefore,
an experiment was made with a 0.15 m (6.0 inches) ID by 0.08 m (3 inches) long, '
>'
side-port extension spool added to the 0.22 m (8.75 inches) ID, insulated com-
bustor to increase the effective combustion chamber diameter. Runs 4i8 through
105
-------
TABLE 11. COMPOSITION AND PROPERTIES OF FUEL OILS USED
IN OIL BURNER EXPERIMENTS
Carbon, % (Weight %)
Hydrogen, %
Nitrogen, %
Sulfur, %
Ash, %
Carbon/Hydrogen
Gravity, API at 60 F
Specific Gravity
BTU/lb Gross
BTU/lb Net
No. 6 Oil
86.66
11.85
0.32
0.97
0.018
7.31
18.8
0.94
18,398
17,548
No. 2 Oil
86.79
13.16
0.009
0.11
-
6.59
36.1
0.84
19,481
18,372
22% No. 6/
78% No. 2*
86.76
12.87
0.08
0.30
0.004
6.75
33.0
0.86
(19,240)
18,190
Calculated values for mixture using linear weight percent
averaging.
106
-------
424 were made with this extension; however, evaluation should be restricted to
Runs 422 through 424 as the prior runs had oil supply pressure (Runs 418 through
420) and chamber warm-up (Run 421) qualification restrictions. The addition of
the standoff spool did improve the smoke characteristics compared to the 0.22 m
(8.75 inches) ID combustion chamber system, which had unacceptably high smoke
emission with the spoolless configuration. However, it did not lower the NO
emissions, and the nominal level remained approximately 15 to 20 percent (0.69
g/kg) higher than the tunnel-fired configuration at 10-percent excess air and
30-percent FGR.
107/108
-------
SECTION VI
PROTOTYPE SYSTEM DESIGN INVESTIGATION
Results from the analytical and experimental studies reported in Sections IV and
V, together with past experience and current contacts with furnace manufacturers,
were synthesized to arrive at preliminary conceptual designs for a warm-air fur-
nace and a hydronic boiler. The decision was made that the optimized conventional
burner head would be used in a cooled combustion chamber for both designs.
Although the different cooling media of" the furnace and the boiler cause substan-
tial differences between the designs, a substantial commonality exists between
them concerning basic objectives and approaches taken to meet the objectives.
Therefore, the areas common to the two designs are subsequently discussed before
specific descriptions for each design are presented.
Each prototype design concept was discussed with one or more representatives of
a manufacturing company that makes and markets residential space heating equip-
ment of the appropriate type. The format for this section first provides a des-
cription of each preliminary prototype design. Following which are given a sum-
marization of manufacturer's comments, criticisms, and suggestions pertaining to
each design. Appropriate responses to those manufacturer inputs, such as concept
revision, design modification, and provisions for data-taking, are then discussed.
PRELIMINARY DESIGNS: AREAS OF COMMONALITY
Objectives
Design objectives were to conceptualize residential heating units with a nominal
1 ml/s (gph) firing rate that would satisfy the following requirements, listed in
order of priority (with low emissions and high efficiency being of equal and
primary concern):
1. Low pollutant emissions
a. Nitric oxide (NO) _<0.5 g/kg of fuel burned
109
-------
b. Carbon monoxide (CO) £1.0 g/kg
c. Unburned hydrocarbons (UHC) <. 0.1 g/kg
d. Smoke £ Bacharach No. 1 scale
2. High efficiency (10 percent or more above conventional)
3. Compliance with safety codes and operational standards
4. Competitive costs, both initial retail price and subsequent operation
and maintenance
5. Quiet
6. Compact, if possible.
Design options have been evaluated on the basis of a newly manufactured product
line, with compromises involving retrofit versatility given a much lower priority.
However, product saleability certainly is of concern and, therefore, unit costs
and acceptability to manufacturers, service personnel, and customers have been
considered in the design selection process.
Burner Unit
The oil burner unit used in both prototype furnace conceptual designs is a con-
ventional burner (Beckett Model AF body) fitted with an optimized, nonretention,
burner head and an air flow/draft loss control device (Fig. 34).
The optimized burner head is described in Section IV, and consists of six air
swirl vanes canted 25 degrees from the blast tube centerline and a firing-rate-
dependent choke diameter. The air swirl vanes are relatively large, approximately
0.05 m (2-inches) long, and extend from approximately 0.03 m (1.2-inch) diameter
out to the diameter of the blast tube. The choke diameter size is dependent upon
the specific installed firing rate according to:
f lo .
D, . 0.0254 2.57 w ., , , / .
(m) L oil(ml/s)J
110
-------
Wire Igniter
Connection
External Flap Lever
and Set Screw Assembly
Air Flow/
Draft Loss
Control Flap
Optimized
Head
Figure 34. Schematic of the Residential-Size Optimized Burner With Positive-
Pressure Air Flow/Draft Loss Control Device
-------
or, equivalently,
r i°-4
D-. , , = 2.7 w ., , ,-.
(inch) L oil(gph) J
The optimum choke diameter for the 1 ml/s (gph) burner is 0.042 m (1.65 inch).
Both the choke diameter and the swirl angle requirements were determined experi-
mentally, and the combination of these specifications result in minimum NO con-
centrations with very low carbonaceous pollutant levels at quite efficient
operating conditions (C02~13.5 percent, SR«1.10). The 1 ml/s (gph) optimized
head configuration has been fitted to both a 0.076 m (3-inch) and a 0.102 m (4-
inch) diameter blast tube with satisfactory results. The anticipated higher
efficiency of the furnace system will stem primarily from the capability of the
optimized burner head to operate reliably at 13.5 percent CO- (with low pollutant
emissions).
Experiments have shown that a nonretention (conventional) type head is lower in
NO emissions than the flame-retention type. The turbulent mixing (well-stirred
reactor) within the retention cone promotes very rapid combustion with no heat
being removed. At near-stoichiometric conditions (desirable for higher efficiency),
this mixing produces quite high adiabatic peak flame temperatures and enhances
production of high NO concentrations (Ref. 5). Therefore, the optimized burner
head is of the nonretention, "plug-flow" type design.
The combustion air flow/draft loss control device (Fig. 34) is intended to provide
improvements in both operating efficiency and combustion roughness. The device
is a weighted flap that drops to a closed position when the burner unit (combus-
tion air fan) is deactivated. Such closing of the combustion air flow passage
eliminates the natural convection air flow through the furnace during the standby
mode and thereby reduces standby thermal losses (~1 to 4 percent). Improvement
in combustion roughness is expected to result from the air flow restriction device
being located on the discharge side of the air fan. Such location allows an un-
restricted air inlet and increases the static pressure within the fan chamber. The
higher operating pressure should reduce the susceptibility of the air fan to
112
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"stalling" (resulting in fan/flame coupled oscillations), and thereby improving
the rough combustion characteristics. The air fan output pressure of the 1 ml/s
2
(gph) optimized burner is approximately 310 N/m (1.25 inch WC) at zero flowrate
(dead-headed), typical of units capable of up to three times the firing rate.
The primary burner control unit will use a solid-state, interrupted-spark ignition
(on-for-start) device with ultraviolet (UV) flame detection cell. Specification
of solid state and interrupted spark reduce electrical power requirements by
about 100 watts (~330 Btu/hr). However, considering the 35 percent efficiency
of the electrical power generation loop, the net overall energy reduction is about
1000 Btu/hr or approximately 1 percent of the furnace output. An ultraviolet flame
detection cell is specified because the optimized burner flames, under certain
conditions, have had low luminosity (yellow spectrum), making applicability of
CdS cells questionable. A potential problem arising from the sensitivity of the
ultraviolet cell to emission from ignition sparks can be avoided by providing a
shorter spark-on duration (~20 seconds) than safety control decision time (30 to
45 seconds). Such an early termination of the ignition spark will allow the
ultraviolet cell to correctly evaluate and verify a "flame-on".condition.
Operational Efficiency Gains
Operational efficiency gains are expected to result from higher cycle-averaged
heat recovery from the furnace flue gases and from reduction of typical energy
losses.
Both conceptual prototype furnaces have burner/firebox designs optimized to operate
much closer to stoichiometric conditions (SRwl.10 to 1.15) than either the average
of the installed existing furnace population (SRsal.9) or the typical current
commerical practice (1.2 <. SR <. 1.6). The effect of stoichiometric ratio on net
cycle-averaged thermal efficiency of a common 1 ml/s (gph) residential furnace
fired with several different burner heads has been illustrated in Fig. 9. Con-
sidering the midrange of the several operating lines in Fig. 9 to be representa-
tive of typical furnace behavior, net thermal efficiency may be estimated to
113
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increase by approximately 4 percent to 13 percent if the stoichiometric ratio is
decreased to 1.15 from 1.4 and 1.9, respectively. For the most part, these
efficiency gains are achieved because the mass flowrate of flue gases is reduced
as stoichiometric ratio is lowered, so that less sensible heat is convected up
the flue. This would be the only reason for increased efficiency if the flue gas
temperature were to remain invariant as stoichiometric ratio is changed. However,
the reduced flue gas flowrate and altered flame combustion characteristics also
alter the heat transfer conditions so that flue gas temperatures are not likely to
stay constant. Normally, flue gas temperatures are somewhat lower at lower stoich-
iometric ratios, which contributes a portion of the actual efficiency gain. The
range of operating line slopes experienced with different burner heads, Fig. 9,
was probably caused by differing flame pattern influences on heat exchanger
effectiveness. Further efficiency increases are being sought by designing the
prototype units to minimize flue gas temperature within the firebox draft, con-
densation, and corrosion constraints discussed in Section IV.
Another substantial contributor to the convection of energy up the flue is the use
of heated and humidified living space air for burner combustion air and for fur-
nace barometric pressure control draft air (Fig. 35a). Energy deposited in room
air used in these ways has been shown (Ref. 12) to be on the order of 10 percent
of the residential heating load. These losses can be eliminated by supplying
outdoor ambient air to the burner and to the barometric pressure control device.
Such an air supply is known as a sealed air system; a typical installation is
shown schematically in Fig. 35b.
Use of a sealed-air system to supply barometric control air should reduce the flue
heat losses, on the average, by an amount equivalent to raising the furnace effi-
ciency by about 8 percent. No operational problems are introduced, with the
possible exception that very cold barometric control air might result in some
temporary local condensation of combustion gas moisture in the flue system.
Insulation of the sealed air system, up to and including the plenum around the
barometric control device, will prevent convective heat transfer from the living
space air and maximize the thermal benefits of using the sealed air system.
114
-------
Cn
Flue
Exhaust Gases
i
Furnace
Barometric
Control
/
Room
Air
Burner
N\\\\\\\\\N\\\\\\\\\\\\\V
(a) Conventional Furnace Installation
v\
Flue
Exhaust Gases
Outdoor Air
Barometri c
Control Air
Furnace
Sealed-Air
Plenum
'Combustion
Air
Burner
\\\\\\\\\\\\\\\\\\\\\\\\\
(b) Installation with Sealed-Air Supply
Figure 35. Conventional and Sealed-Air Furnace Installations
-------
The impact on fuel utilization efficiency of using living space air for burner
combustion air is less than 1/4 that of using such air for barometric control
air. A lower average flowrate and the opportunity to recover the extra input
heat in the heat exchanger contribute to lessening the impact. Nontheless, the
2-percent or so potential fuel saving is worth the little extra effort to include
combustion air in the sealed air system, as less incentive exists for insulating
the combustion air duct from being heated on its passage through the living
space. In fact, in very cold weather, the outdoor air needs to be so tempered
in order to avoid ignition problems and excessive start spikes of CO, UHC, and
smoke. The combustion air will also be sealed and filtered to reduce effects
of dust, lint, animal hair, etc., and maintain consistent combustion air-fan
performance at the near-stoichiometric burner operating conditions. In addition
to providing a modest heating system efficiency gain, an enclosed combustion air
system also beneficially reduces odors, combustion noise, and combustion fan
noise in the vicinity of the furnace.
Still another contributor to the convection of heat up the flue is the cooling
of furnace components by a natural draft flow of air through the burner, firebox,
etc., during the burner-off or standby period. Losses up to about 4 percent of
the residential heat load may be caused by this phenomenon. In the conceptual
prototype designs, the sealed combustion air systems should partially inhibit
the natural draft flow through the system, but the draft loss control flap in
the burner is intended to completely prevent it.
Combined, the foregoing reductions in flue heat losses sum to the order of 14-
to 24-percent increase in cycle-averaged fuel utilization efficiency for the
residential heating system. While the sealed air devices can affect substantial
overall fuel economies, such devices do not contribute to increased thermal
efficiency of the furance unit per se. Thus, reducing the stoichiometric ratio
and stopping standby natural draft losses should raise the cycle-averaged furnace
efficiency by between 6 and 16 percent, depending upon the magnitude of the
stoichiometric ratio change.
116
-------
Additional sources of heat losses which these conceptual prototype designs do
not attempt to alleviate are:
1. The remaining sensible and latent heat in the flue gases (each on
the order of 8 percent). Much larger heat exchangers, corrosion-
resistant heat exchangers and flues, and the provision for conden-
sate disposal and induced-draft, hot-gas fans would be required to
achieve recovery of these losses. Such furnace attributes are con-
sidered not to be cost-competitive at this time.
2. Convention and radiation losses from the external cabinet of the
furnaces (as much as 2 percent for warm air and 3-1/2 percent for
hydronic units). Although such losses are from a furnace unit
efficiency view-point, they commonly contribute to heating the
living space and are not really residential heating system losses.
The reduction of these losses is predominantly an insulation
problem involving economic tradeoff of better furnace housing
insulation and perhaps slightly better fuel economy against increased
initial cost and attendant loss of competitiveness.
3. Transient start-up and shutdown losses, such as a starting puff of
UHC, the dribble of oil from the spray nozzle after the burner is
turned off, a flue gas temperature spike before the warm air fan is
turned on. The aggregate of these losses is well below 1 percent;
they are normally minimized because of emissions and good design
standpoints rather than for the sake of efficiency.
Combustion Chambers
The combustion chamber is the primary component in the path of the combustion
gases for controlling pollutant emissions. Therefore, extensive burner/combustor
optimization experiments were conducted with the 1 ml/s (gph) optimized burner
to provide information for favorable combustion chamber configurations that allow
the burner to operate very near stoichiometric (SRssl.10, CCL^IS.S percent) with
minimum pollutant emissions and smooth combustion (Section IV).
117
-------
A simplified summary of the results of cyclical (10 minutes on/20 minutes off)
burner/chamber optimization experiments has been given in Table 8. From these
data, it was concluded that:
1. Meeting the NO emission goal is more difficult than meeting the other
emissions goals.
2. Cooled combustion chambers are required in order to begin to approach
the NO goal of £0.5 g/kg fuel. (The combustion chamber is then the
first, or primary, heat exchanger.)
3. Combustion chamber diameters of 0.25 m (10 inches) or larger are also
required to reach the NO goal.
/
4. Corabustor lengths in excess of 0.5 m (20 inches) are then required in
order to ensure completion of combustion before reactions are quenched
in the secondary heat exchanger.
5. Either the tunnel- or side-fired configurations can be used.
At this point, the different cooling media force different design approaches to
be taken for the warm-air andhydronic units.
CONCEPTUAL DESIGN STUDY: WARM-AIR FURNACE
The warm-air-furnace design study commenced with evaluation of combustor cooling
methods. A water-cooled combustor/warm-air heat exchanger combination was con-
sidered for the favorable heat transfer characteristics of the water coolant.
However, the fabrication costs for a boiler-code-certified pressure vessel com-
bined with a complicated combustion-gas-to-warm-air heat exchanger were considered
to be prohibitive in the residential-size, warm-air-furnace market. Therefore, a
finned, air-cooled combustor configuration that could more or less simulate the
heating and cooling charcteristics of a water-cooled chamber was conceived. A
preliminary layout drawing of such a combustor, combined with a commercial warm-
air-furnace heat exchanger, is shown in Fig. 36, where the combustor is side-fired,
118
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^=-_
I
/
f
NOTE: Detail dimensions are
called out in inches to
maintain commonality with
material stock.
Figure 36, Preliminary Layout Drawing of the Air-Cooled-Combustor,
Warm-Air Furnace
119
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constructed from standard 0.25 m (10 inch) pipe with a standard pipe cap, and
is rather heavily finned to provide both substantial convective cooling and a
large heat sink.
Rather than this combustor design concept for the prototype warm air furnace
being simply accepted, a finned research combustor was fabricated and tested to
verify the pollutant emissions and performance expectations of a well-finned,
air-cooled combustor. This additional experimental effort and its results are
documented in Appendix F. The results are seen to be very promising, with the
carbonaceous pollutants well under control [L = 0.75 m (SOinches)] at design
operating conditions (10- to 15-percent excess air), and the NO emissions (~0.6
g/kg) approaching the target value of 0.5 g/kg. The total heat extracted from
the finned section [8200 J/sec (28,000 Btu/hr)] was more than the 7300 J/sec
(25,000 Btu/hr) limit established by the earlier water-cooled experiments wherein
no emissions or stability problems were encountered. The differences in heat
transfer profiles between the cooling methods were probably responsible for both
the differences in heat extraction and in NO emission between the two types of
cooled chambers.
A heat transfer analysis of the combustor/heat exchanger combination shown in
Fig. 36, was conducted to verify design estimates; the results are presented in
Table 12. The analysis estimated a heat extraction rate from the combustor of
6260 J/sec (21,400 Btu/hr) at 0.707 m /s (1500 cfm) coolant flow with the 0.076m
(3 inch) fin extension section. The experimental finned combustor, without the
extended fins, removed 8200 J/sec (28,000 But/hr) at 1.18 m /s (2500 cfm) air
flowrate. Use of the Dittus-Boelter equation to account for the differences
in surface areas and coolant flow enables a calculated experimental heat extrac-
tion value for the lower air flowrate (0.707 m /s)of 6500 J/sec (22,200 Btu/hr)
to be derived, which is in good agreement with both the design estimate and the
heat transfer analysis (~3.5 percent).
Although the total heat removal was well within design expectations, some re-
arrangement of the fins and/or a larger combustor diameter was believed to be
appropriate for the prototype, so as to ensure meeting the 0.5 g/kg NO
requirement.
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TABLE 12. SUMMARY OF FINNED COMBUSTOR/HEAT EXCHANGER OPERATING
CONDITIONS HEAT TRANSFER ANALYSIS
Finned Combustor Section
Ta.r in/out, °C (°F)
Tgas 1n/OUt> °C (°F)
q, J/sec (Btu/sec, Btu/hr)
h W/m2-K (Btu/in.2-sec-R)
h > W/m2-K (Btu/in.2-sec-R
T °C (°F)
TWG' ; ( h;9
A,0, nf (in/)
by n n
Asc, m2 (in.2)
Oil Flow, ml/sec (gph)
Wn;,c' k9/sec Ob/sec)
.gas
Wa. , kg/sec (lb/sec)
«
(in)
o
(in/)
a
a i r
A_,
y
Ac,
AP Pa (psi)
u
APg, Pa (psi)
21/28 (70/83)
1871/1578 (3400/2872)
6260 (5.93/21,400)
12.9 (0.440 x 10"5)
24.9 (0.847 x 10"5)
135 (275)
.309 (480.1)
2.311 (3583)
1.05 (1.0)
0.015 (0.035)
0.852 (1.88)
0.050 (78.4)
0.092 (143.3)
81.3 (0.0118)
0.50 (7.21 x 10"5)
Heat Exchanger Section
Tair in/out' °c (°F)
T«e in/out, °C (°F)
gas
q, J/sec (Btu/sec, Btu/hr)
h , W/m2-K (Btu/in.2-sec-R)
h,, W/m2-K (Btu/in.2-sec-R)
sc!
V
*C>
AP
m2 (in.2)
m2 (in.2)
n2 (in.2)
n2 (in.2)
Pa (psi)
Pa (psi)
28/62 (83/144)
1578/224 (2872/435)
28,900 (27.4/98,650)
28.8 (0.978 x 10"5)
71.5 (2.43 x 10"5)
232 (450)
1.713 (2656)
1.713 (2656)
0.040 (63.4)
0.112 (174.4)
95.8 (0.0139)
0.50 (7.31 x 10"5)
121
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Preliminary Warm-Air-Furnace Design
The preliminary conceptual warm-air furnace design is illustrated in Fig. 37.
This design is based on modifying an existing, commercially purchased warm-air
furnace* to incorporate the larger-diameter, forced-convection, air-cooled com-
bustion chamber matched to the 1 ml/s (gph) optimum burner and sealed air system.
The combustion chamber for the prototype warm-air furnace is a carefully optimized,
side-fired, cool-wall configuration. Based on the optimization experiments, the
combustor section is provided with 24 radial fins, increasing the combustor out-
side surface area approximately 6-1/2 times, to remove approximately 20 to 25
percent of the total furnace heat output. The primary purpose of cooling the
combustion chamber is to reduce the peak flame temperatures and thus minimize
formation of NO. However, it has been determined experimentally that removal
of more than 25 percent of the heat from the combustion zone leads to premature
quenching of the chemical reactions, and incomplete combustion products (CO, UHC,
and smoke) in the flue gases.
The combustion chamber inside wall temperature is expected to average about 230
to 260 C (450 to 500 F) even though an inner, hot-wall retort will not be used.
This lower-than-usual metal wall temperature results from using a relatively
large 0.30 m (12 inch) firebox inside diameter (ID), which should be less suscep-
tible to direct flame impingement on the wall, and from using a quite massive
firebox construction. For the prototype design, the combustion chamber wall
consists of a standard 12-inch pipe cap welded to an eccentric 12-inch to 10-
inch pipe reducer, the small end of which is mated to the existing secondary
heat exchanger of the furnace. The eccentric reducer configuration is expected
to be beneficial in several ways. The rear-biased eccentricity allows a packag-
ing advantage by allowing the burner to sit closer to the unit centerline. The
*A Lennox Model 011-140 furnace is the unit selected.
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to
1.45m
fflfcf
SECTIOM A-A
KlO.
Figure 37. Layout Drawing of the Finned-Combustion, Prototype Warm-
Air-Furnace Unit With a Sealed Air System
-------
forward-biased combustor exit induces combustion gases back along the cooled walls
to the burner area enhancing a combustion gas recirculation (CGR) effect to help
ensure lower peak-flame temperatures (i.e., lower NO ). Also, the reduction in
A
flow area caused by the reduction from 12 to 10 inches accelerates the combustion
gas flow and promotes mixing of any remaining fuel species. The reduction in
warm-air passage area on the side opposite could be a problem, because the air
coolant flowrate is reduced there; additional fins have been added in that area
that is likely to have the highest local wall temperature.
With the external fins welded to those firebox components, the combustion chamber
assembly is expected to weigh on the order of 90 kg (200 pounds). This rather
massive construction was intentionally adopted to accomplish two objectives other
than to simply reject heat from the combustion zone. Both are related to the
heat sink aspects of a massive chamber. First, a massive heat sink chamber can
more readily approach uniform inside surface temperatures than can a lightweight,
unlined metal chamber. This fact is important for avoiding excessive NO forma-
tion and wall erosion (or possibly burnout) at "hot-spots", and for avoiding smoke
and UHC formation at "cold-spots".
The second objective is for the combustion chamber to retain most of its stored
heat during burner-off times so that the firebox is considerably wanner at
burner startup than is the usual warm-air-furnace practice. This objective
arose from the observation that start-transient spike emissions of CO, smoke,
and UHC were considerably lower with water-cooled than with uninsulated air-cooled
combustors. In a very real sense, the massive, finned combustion chamber design
represents a way to achieve beneficial effects found experimentally with water-
cooled chambers.
The side-port orientation was chosen principally for its conventional packaging
in "Hi-Boy" or "Low-Boy" flexibility. The opposing wall in the side-fire orien-
tation, like the bias in the eccentric reducer, also helps to induce recircula-
tion of some partially cooled combustion gases back into the air/oil mixing and
combustion zone, aiding in maintaining flame zone temperature uniformity and
combustion stability.
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The prototype warm-air furnace will retain a conventional secondary heat exchanger
for several reasons. The lowering of pollutant emissions is accomplished by the
burner and combustion chamber, and modifications to the heat exchanger have very
little effect on the composition of the flue gas. While increased thermal effi-
ciency will be accomplished primarily by the ability of the optimized burner head
tD operate near stoichiometric conditions, the addition of the cooled combustion
chamber supplements that gain by increasing the heat exchanger capacity by about
25 percent. Thus, the heat exchanger could be made smaller. However, since the
heat exchanger accounts for only 20 to 30 percent of the total furnace volume,
a significant reduction in heat exchanger size (with associated increase in cost)
could result in only a nominal reduction in furnace unit size. Therefore, the
Lennox heat exchanger section supplied with the stock furnace unit will be used
in the construction of the prototype furnace.
Manufacturer Review. The layout drawing and an explanatory memorandum for the
preliminary, conceptual, warm-air-furnace design were sent to engineering per-
sonnel at Lennox Industries for review and comment. The Lennox comments and
recommendations were transmitted verbally to the authors during a June 10, 1975
visit to Lennox's main plant in Marshalltown, Iowa. The substance of discussions
held that day are documented in this subsection.
Firebox. The massive, finned-steel firebox was viewed as being a technically
viable approach, whose experimental test results should prove to be interesting.
Reservations were expressed in two specific areas: metal temperature and weight.
Some doubt was expressed as to whether the heat sink capacity of the firebox would
be high enough to keep its average inside wall temperature anywhere near the 230
to 260 C (450 to 500 F) estimated by heat transfer calculations, and that a sub-
stantially higher-than-average temperature hot-spot opposite the side-fired burner
port may be experienced.
An estimated firebox weight had not been included in the explanatory memorandum.
Upon learning that it would be well in excess of 50 kg (110 pounds), Lennox per-
sonnel recommended that its weight be minimized, within the constraint of ensuring
125
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acceptable pollutant emissions. Three areas of concern were indicated: shipping
weight, supporting the extra weight within the cabinet, and extra bearing loads
on floors of frame residences. The most serious of the three is the increased
shipping weight; it would tend to reduce a manufacturer's competitiveness with
increasing distance from his factory. Two illustrations were given. First,
Lennox has converted from firebrick firebox linings to refractory fiber linings,
largely to reduce shipping weights by approximately 11 to 14 kg (25 to 30 pounds).
Second, Twentieth Century's warm-air furnace, featuring cast-iron fireboxes and
cast-iron heat exchangers, weighs on the order of 180 kg (400 pounds) more than
competing warm-air furnaces and, as a direct result, is essentially limited to a
one-city market.
Burner. Having studied Ref. 5, which details the development and perfor-
mance of the optimized burner, Lennox personnel apparently feel comfortable with
the potential applicability of the technology supporting this burner. However,
they did make comments and suggestions concerning several aspects of its appli-
cation in the preliminary conceptual design.
The gravity-actuated flap within the burner to eliminate draft air flow during
standby was the item of most concern. The group generally agreed that the flap
would substantially change the uniformity of air flow distribution within the
blast tube, possibly degrading the ability of the burner to operate at low
stoichiometric ratios with acceptably low smoke and UHC emissions. Reliability
of flap operation was questioned. In fact, it was stated that such an alteration
of the burner may abrogate its UL listing, and that recommending the use of
unlisted equipment entails substantial legal liabilities for manufacturers. Gain-
ing a UL listing is not very likely unless a positive flap actuator, such as a
solenoid, and an electrical sensor, to confirm the actuator is open, are provided.
A preferred way of eliminating draft air, particularly with the sealed combustion
air supply, is to place a low-cost solenoid-actuated, butterfly valve in the com-
bustion air supply line to the burner.
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Some burner problems may develop as a result of eliminating the standby draft
air flow. Oil spray nozzles are commonly coated with a varnish-like residue if
they are heated above about 95 C (200 F). In conventional equipment, nozzles
may be heated convectively by backflow of combustion gases into a furnace room
with negative ambient pressure [room drafts as low as -25 N/m (-0-1 inch WC)
are not uncommon], radiantly by "viewing" high temperature firebox refractory
surfaces, or both. If a backflow is not present, an inflow of draft air normally
provides enough convective cooling to offset the radiant heating. Positive cut-
off of the burner air supply duct would eliminate the backflow problem but accen-
tuate the radiant heating problem. The low-temperature steel firebox should
reduce the magnitude of this latter problem, which needs to be estimated and
checked experimentally. A related problem is the possibility of overheating the
flame detector cell and the high-voltage ignition components, none of which should
be allowed to reach temperatures above 52 C (125 F). Radiant heat transfer from
the firebox to these components is not expected to be much of a problem because
they have a rather limited exposure to the combustion chamber through the burner
choke plate. However, if the objective of maintaining a warm firebox during
standby is achieved, heat conduction to these temperature-sensitive components
may be a significant problem. Monitoring of burner component temperatures during
prototype furnace testing was recommended.
Another potential overheating problem arises from use of the burner vestibule as
a plenum for the sealed combustion air supply. The vestibule cover, a part of
the external furnace cabinet, is louvered at both the bottom and the top. Nor-
mally, considerably more furnace-room air circulates by natural draft through
the vestibule than is drawn into the burner as combustion air; this flow cools
t
the vestibule. It was doubted if the combustion air alone would provide suffi-
cient vestibule cooling. Three problem areas should be anticipated: overheating
burner components, overheating electrical controls mounted in the vestibule, and
overheating painted cabinet panels. Warm-air-furnace codes and standards specify
maximum temperature for electrical wiring of 60 C (140 F). While higher tem-
perature wiring could be used, at some additional expense, it is unlikely that
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anything but 60 C (140 F) wiring would be used for field installation, so this
temperature limit should be maintained. It was recommended that vestibule tem-
peratures be measured during prototype furnace testing and, if they are too high,
that vestibule louvers having a separate sealed air duct to the burner again be
used.
Ultraviolet flame detector cells are suitable, although more expensive than CdS
cells, largely because the ultraviolet cells are not produced in large quantities.
Concerning the criticality of timing the cut-off of the spark igniter and the
flame-on decision point, it was noted that Honeywell has anultraviolet flame
detector system whose circuit is integrated with the ignition system such that
the circuit evaluates the flame intermittently (60 Hz), between sparks..
Several miscellaneous comments regarding burners may be helpful. Rocketdyne per-
sonnel were reminded that in selecting combustion air fan characteristics, ambient
air density varies with altitude. For example, at the same furnace firing rate
and excess air condition, approximately 10-percent greater volumetric flowrate
of air is required in Denver than in Des Moines. Another cautionary note: the
closer a burner is operated to stoichiometric conditions (high C02), the more
sensitive it is to unintentional in-leakage of air. That is, maintaining seals,
closing peep-holes, etc., becomes more critical.
Furnace Evaluation Test Procedures. An appropriate experimental warm-air-
furnace performance evaluation set-up, minimum instrumentation requirements, and
test procedures are delineated in ANSI Z91.1; safety codes are given in UL 727
for warm-air furnaces and UL 296 for oil burners. There is also a new standard
in preparation, ANSI Z91.2, for oil burners smaller than 6.7 ml/s (7 gph).
Lennox is currently performing extensive experimental laboratory characteriza-
tion of several stock furances. The company is using a set-up similar to that
specified in ANSI Z91.1, but is making substantially more than the minimum meas-
urements. As an example, ambient air is supplied to the burner, for both com-
bustion and draft air, through a forced-draft metering section that has a draft-
balancing system for ensuring zero-draft at the burner inlet. Thus draft-air
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thermal losses can be determined as functions of other variables, such as cycle
frequency, fractional burner-on-time, and firing rate. The first furnace tested,
fired at 1 ml/s (gph), was found to have about 1-1/2-percent draft-air loss dur-
ing 8 minutes on/4 minutes off cycles; this condition increased to about 2-1/2
percent during 4 minutes on/8 minutes off cycles. Rocketdyne personnel were
privileged to examine the Lennox test setup. (Several of their ideas are reflected
in the prototype furnace test assembly to be used in Phase II of this program.)
Compact Heat Exchangers and Achieving Higher Efficiencies. Since the proto-
type furance design is basically a set "of modifications to a Lennox Oil furnace,
which entails essentially no impact on the secondary heat exchanger, Lennox re-
marks on this component were confined to explanations of potential hot spots and
a breif discussion of effective ways of mating the heavy-walled, combustion-
chamber section to the thin-walled, 17-gage heat exchanger. Essentially, it was
agreed that a simple nonwelded mechanical connection could be used in a proto-
type unit, even though this might be unacceptable in production runs because of
high cycle-fatigue or long-term leakage potentials.
Although the parallel-plate, clam-shell, secondary heat exchanger used in the
Lennox Oil furnaces is relatively compact, Lennox is considering using a modified
annular type secondary heat exchanger surrounding the combustion chamber in the
Lennox new furnace series. Mainly, this design change reduces the sheet metal
welding cost and results in a less-expensive furnace. Lennox periodically examines
more-compact, higher-effectivity heat exchanger concepts, but commonly finds that
such concepts either do not meet their proponents' claims, simply cannot compete
economically, or both. Lennox expects an industry response to limited supplies
and higher costs of fuels will be to abandon industry's reliance on conventional
chimney-type flues. Exhausting the flue gases through a short, small-diameter,
corrosion-resistant vent through the outside wall of the residence would require
the use of a pressurized combustor but would permit extraction of most of the
sensible heat of the exhaust gas, would obviate the need for barometric control
air, and would virtually eliminate draft-air losses. A conventional heat
129
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exchanger for such a unit would need at least two to three times the area of
present heat exchangers, so that higher fabrication and shipping weight costs
might then support the use of compact heat exchanger designs.
Prototype Design Modifications. In response to the manufacturer's comments and
suggestions, some aspects of the preliminary prototype design were reassessed
and modifications made.
Firebox. Cyclical furnace performance analyses were undertaken with the
WAFURN warm-air-furnace thermal analysis computer program to assess the thermal
effects of reducing combustion-chamber and cooling-fin thicknesses. It was
determined that, from a zero-dimensional, heat-sink standpoint, a considerably
lighter assembly could provide the desired average thermal behavior. It was
also decided that both the firebox wall and fin thicknesses could be reduced
to 0.0063 m (1/4 inch), whereupon the assembly weight would be about 49 kg (108
pounds).* It was thought that even thinner fins would exhibit significant
temperature gradients and force the firebox to deviate substantially from an
ideal heat sink. This may prove to be no problem, in which case even greater
weight reductions can be considered for the final design.
Burner. A decision was made to eliminate the draft-air control flap from
inside the burner. The intended function of the flap will be performed by a
solenoid-actuated butterfly valve in the sealed combustion air supply duct.
Concerning the possible overheating of burner components, it was decided to pro-
vide insulation between the burner and the firebox to limit conductive heating.
Excessive radiant heating of the spray nozzle, flame detector, and ignition
*This weight is substantially below the originally projected 90 kg (200 pounds).
However, it is based upon using Schedule (Sen) 20 pipe components. Attempts
to locate Sch 20 caps and eccentric reducers have so far been fruitless. If
none is found, the prototype furnace will be built with standard Sch 40 compo-
nents, and the finned firebox will weigh about 57 kg (126 pounds).
130
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system will be watched for during prototype furnace evaluation testing. Similarly,
possible problems from too high temperatures in the vestibule will be monitored.
One change will be made in the design to better cool the electrical controls and
burner motor: the combustion air inlet to the vestibule will be removed from the
left to the right side, where the controls are located.
CONCEPTUAL DESIGN STUDY: HYDRONIC BOILER
Preliminary Hydronic Unit Design
The preliminary conceptual hydronic furnace design is illustrated in Fig. 38.
This design is based on substantially all new construction, principally because
no commercially available unit was identified as being readily adaptable to the
larger-diameter, water-cooled combustion chamber for the optimum burner.
Based on the burner/firebox matching experimental results, a horizontal, tunnel-
fired, water-cooled, combustion-chamber configuration was selected. The tunnel-
fired configuration offers structural, packaging, and pollutant emissions ad-
vantages over the side-fired arrangement. The straight-through, tunnel-fired
chamber can be welded directly to the front and rear boiler tube sheets, con-
tributing to the pressure vessel strength. Analysis of a side-fired arrangement
indicated such an arrangement would increase installation volume by nearly 50
percent because it needed an additional 0.2 m (8 inches) of width for burner
placement as well as a further additional 0.15 to 0.2 m (6 to 8 inches) to allow
room for servicing the burner. By contrast, the tunnel-fired arrangement employs
the space below the flue manifold for the burner and provides all required access
for servicing, and combustor and heat exchanger cleaning from the front of the
unit.
The firebox is designed to remove between 20 and 25 percent of the total heat
output from the combustion gases before they enter an internally finned fire-tube,
secondary, heat exchanger section. It has been determined experimentally that
131
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KJ
SECTION A-A
Figure 38. Layout Drawing of the Horizontal-Pass, Prototype Hydronic-Furnace With a
Sealed Air System
-------
removal of more than 25 percent of the heat from the combustion zone leads to
premature quenching of the combustion reactions and produces high concentrations
of CO, UHC, and smoke. That fraction of the heat can be extracted in about 0.3m,
(12 inches), so the water-cooled combustion chamber is only 0.28 m (11 inches)
long. Combustion reactions are not expected to be complete by then; the flow-
reversing firebox-to-secondary-heat-exchanger manifold is considered to be part
of the primary combustion volume. It is partially insulated to lower the heat
extraction rate and avoid further quenching of the combustion process.
Close confinement of the burner flame ean also lead to premature quenching of
part of the combustibles (fuel contacting the cooled walls). Additionally, flame-
front instabilities (combustion roughness) can be caused by too-high bulk gas
velocities in small-diameter chambers. Therefore, and in conformance with the
experimental burner/firebox matching results, a relatively large 0.28 m (11 inch)
ID, tunnel-fired, combustion chamber is proposed for the optimum hydronic unit.
The combination of the optimized burner, the large-diameter combustion chamber,
the controlled heat extraction, and a combustion-chamber, discharge, choke plate
is expected to prevent the occurrence of combustion roughness commonly experienced
with nonretention head burners in wet-base combustors.
The chamber discharge choke plate is shown in Fig. 38 simply as a 0.178 m (7 inch)
hole in the rear tube sheet. The purpose of the plate is to promote recircula-
tion of the combustion gases back along the cooled walls to the burner insertion
region. Ingestion of these partially cooled gases into the initial flame region
aids in stabilizing the flame and, adding a mild diluent slightly reduces peak flame
temperatures, thereby aiding in the reduction of NO emission concentrations.
Jv
The choke ring also will accelerate the gases leaving the chamber, promoting
better mixing and consumption of residual unburned fuel species.
The secondary heat exchanger is a multiple-pass, horizontal, firetube configura-
tion (see Fig. 38). Three series-connected firetubes are internally finned,
0.15 m (6 inch) diameter pipes*with inner (combustion-gas-side) surface area
*These pipes are manufactured commercially by the Bock Corporation, Madison,
Wisoncin, and are known by the trade name "Turboflue".
133
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3-1/2 times greater than the outer (water-side) surface area for improved gas-
side heat transfer. Successive passes are arranged in an ascending direction
to aid internal combustion gas draft and provide more favorable conditions for
natural circulation of the water during burner-on. The water temperatures will
be hotter on the left side (Fig. 38, front view) and induce a clockwise circula-
tion pattern. The input of return water through the circulation pump is also
biased in a clockwise direction to further -assist this circulation pattern.
The tankless coil will carry a 3-1/2 gpm (2.2 x 10~ m /s) rating. The water
temperature gradients within the tank are not expected to be significant during
the burner-off standby mode. The central height location of the coil is, there-
fore, not foreseen to be a disadvantage as compared to being located at the top
of the boiler. During the burner-on mode, the clockwise circulation of the
heated water places the tankless coil immediately downstream of the last heat
exchanger element, providing very hot water to the coil. The location of the
coil also provides some simplification of the firetube cross-manifolding. The
water temperature sensing element is mounted on the tankless coil flange; this
location was chosen because it is felt important to monitor closely the tempera-
ture of the water that the inhabitants of the dwelling may contact directly.
The baffling within the rear cover is intended to direct the combustion gas
flow so that the full volume of the reduced heat transfer zone is used to complete
combustion before the gases are cooled rapidly in the firetubes. The rear cover
is designed to be removed with only 2 inches of rear clearance required and with
all fasteners accessible from the front. The bolt pattern of the rear cover can
be indexed to prevent misaligned assembly during manufacturing or servicing.
Rear-cover removal is required only if repair of the pyroflex lining is necessary.
All other internal servicing can be done through the front of the furnace.
All water ports are located at their respective preferred locations. The inlet
port enters tangentially at the bottom to assist the clockwise water circula-
tion. The air vent is located at the very top of the tank for optimum air collec-
tion, and the residential heating water outlet is located immediately downstream
of the last heat exchange'r element.
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Manufacturer Review. The layout drawing and explanatory memorandum for the
preliminary conceptual hydronic furnace design were sent to the director of
engineering of the Burnham Corporation for review and comment. Rocketdyne
personnel then visited Burnham on June 9, 1975, to discuss their evaluation
of the conceptual design.
Basically, the preliminary design was assessed as representing a buildable, work-
able hydronic unit. In fact, in 1973 Burnham investigated a hydronic boiler
having many of the same general features. Rocketdyne personnel were given prints
of layout drawings showing to be very similar: vertical front and rear tube
sheets; front-mounted burner, tunnel-fired into a water-cooled horizontal com-
bustion chamber; insulated 180-degree return manifold to a secondary heat ex-
changer section embodying horizontal fire-tubes; and combustor, heat exchanger,
anri tankless coil orientations to promote tangential water circulation in the
boiler. The design was not developed further because it was found to offer no
significant cost advantage over Burnham's existing dry-base, vertical-fire-tube,
production steel boiler design. The ways in which the Rocketdyne conceptual
design differs from the Burnham preliminary design constituted the main points
of discussion. The most significant difference is the use of series-connected
Turboflue sections for the secondary heat exchanger. Although not familar with
Turboflue per se, Burnham has evaluated many candidate extended-surface heat
exchangers and has invariably found them not to be cost-competitive with simple
cylindrical boiler tubes, particularly when the latter are fitted with replace-
able internal turbulators. With installed 0.075 m (3 inch) diameter boiler tube
costing $1 per linear foot, an increased heat transfer area can be obtained at
much less expense by adding one or more boiler tubes than by using extended sur-
face configurations. That this is a valid comment is borne out by Rocketdyne"s
estimate that the installed Turboflue heat exchanger would cost between 2 and 3
times that amount.
Concern was expressed that the Turboflue heat exchanger would have higher pressure
drops than turbulated boiler tubes. Citing Bock Corporation estimates of vertical
Turboflue draft losses was insufficient argument to dissuade Burnham from viewing
135
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this as a probable problem. Further, the series connection involves three 180-
degree flow reversals, which are also high-pressure-loss devices. Combining
these, it was estimated that the firebox would probably be slightly pressurized
and that a larger-size burner (costing an additional 10 to 15 dollars) would be
needed to ensure delivery of adequate combustion air. (Subsequently, these
expectations have been more or less confirmed by draft-loss calculations sum-
marized in the following subsection.)
Another potential Turboflue problem is the avoiding of buildup of hard carbonace-
ous scales in the heat exchanger sections. Burnham experience calls for direct
line-of-slight access for rodding-out (cleaning) such deposits; it is doubtful
that the Turboflue sections could be effectively cleaned by either spiral brush-
ing or water-mist washing suggested by the Turboflue manufacturer. For this
reason, Burnham uses removable, disposable, boiler-tube turbulators.
The refractory linings in the aft flow-reversing manifolds are likely to be
damaged during periodic rodding-out. The fact that the hot combustion gas flow
is being turned against the surfaces most likely to be damaged, increases the
chances of burning through the manifold housing, which would constitute a serious
residential fire hazard. Provisions should be made for ready replacement of
damaged linings.
It was also noted that, for the same reason, the flow-reversing manifolds must
be positively sealed to prevent leakage. One method for achieving such a seal
was suggested: construct a refractory fiber liner such t.hat it has a peripheral
compressed flange serving as a gasket between two bolted metal flanges.
For the foregoing reasons, the aft reversing manifolds might better be built as
separate units than combined in a single metal housing as shown in Fig. 38.
Some comments concerned construction methods. Shearing steel plates rather than
cutting or burning them is less expensive; therefore, square or rectangular tube
sheets are preferable to circular or circular-derivative sheets such as those
136
-------
shown in the preliminary conceptual design. Uncooled protuberances into the
high temperature combustion gases will sooner or later be burned off. Thus,
forming a combustor discharge choke plate as an uncooled section of the aft
tube sheet would be highly unsatisfactory. A bolt-on refractory metal or ceramic
choke plate is preferable. Extended, uncooled ends of firetubes must be kept
to the minimum length needed for welding them to the tube sheets.
Finally, it was noted that building the firebox as sketched in Fig. 38 with a
water-cooled, small-diameter, burner insertion tube, is more costly and both
more difficult to weld and to repair than would be a single cylinder through
both tube sheets. However, the more complicated design provides cooling for the
burner end of the chamber, providing an important contribution to flame-zone
cooling and minimum NO formation.
Prototype Design Modifications. The manufacturer's assessments of the warm-
air and hydronic, preliminary, conceptual designs were discussed with the EPA
Project Officer immediately following their communication to Rocketdyne per-
sonnel. In consideration of those assessments, the estimated relative difficulty
in effecting recommended (and accepted) modifications, the relative population
of warm-air and hydronic units, their production rates, etc., a decision was
made to build and test a prototype warm-air furnace rather than a hydronic proto-
type. As a result, modification of the preliminary hydronic conceptual design
to a producible prototype design became an academic exercise that did not extend
beyond the decision stage.
In response to Burnham's concern about combustion-gas pressure drop or draft
losses, a complete draft loss analysis was performed. Component pressure drops
were calculated for operation at 1 ml/s (gph) firing rate with 15-percent excess
air by means of formulae given in Ref. 13. A limited amount of draft loss data
for a 0.15 m (6-inch) diameter Turboflue were obtained from the Bock Corporation
137
-------
and used to back-calculate friction factors.* The total draft loss from the
2
firebox to the barometric control was estimated to be 7.84 N/m (0.0315 inch
WC) distributed as follows:
Portion of Flow Path Percent of Draft Loss
Choke Plate 11.9
3 180-degree Returns 25.7
3 Turboflue Entrances 13.8
2 Turboflue Exits 9.2
Turboflue Passes 32.8
Flue Connection 6.6
By contrast, comparable calculations for a design with seven, parallel 0.076 m
(3-inch) diameter boiler tubes in a single pass following a single 180-degree
2
return and no combustor discharge choke plate predict only a 1.25 N/m (0.005 in
WC) draft loss. This loss confirms the essential validity of the Burnham expec-
tation of high draft loss for the preliminary hydronic design. Also, the fore-
going distribution table shows that the biggest contributors to the draft loss,
in order of decreasing importance, are the series arrangement with multiple re-
turn bends, inlets, and exits; the Turboflue heat exchanger tubes; and the com-
bustor exhaust choke plate.
The calculated stack effect between the combustion chamber and the barometric
2 2
control is only 1.5 N/m (0.006 in WC). Thus, if a firebox draft of 2.5 N/m
2
(0.01 in WC) is desired, the barometric control device should control to 8.8 N/m
(0.036 in WC) draft. The chimneys on many residences cannot reliably provide
that high a draft at the barometric control; therefore, this condition tends to
confirm the fact that the preliminary design layout would have (or approach) a
pressurized combustion chamber.
*Friction factors calculated were f = 0.24 and 0.51 at Reynolds numbers of
approximately 5200 and 6500, respectively. These values are approximately
an order of magnitude larger than those for relatively smooth pipe.
138
-------
The draft loss could be approximately halved by changing the hydronic unit
design to (1) eliminate the combustor exhaust choke plate, (2) use only one
180-flow return, and (3) use either parallel single-pass Turboflues or one
long one. While such a design would approach giving acceptably low-combustion-
gas pressure drops, the undoubtedly higher fabrication costs and questionable
ability to clean deposits from the Turboflue interior combine to make inadvisable
the pursuit of the application of this device without definitive testing first
being performed.
139/140
-------
SECTION VII
REFERENCES
1. Hall, R. E., J. H. Wasser, and E. E. Berkau, "NAPCA Combustion Research
Programs to Control Pollutant Emissions from Domestic and Commerical Heat-
ing Systems," New and Improved Oil Burner Equipment Workshop, NOFI Tech.
Publ. 108 ED, National Oil Fuel Institute, Inc., New York, N.Y., September
1970, pp. 83-93.
2. Hall, R. E., H. H. Wasser and E. E. Berkau, "A Study of Air Pollutant
Emissions from Residential Heating Systems," EPA-650/2-74-003, Environ-
mental Protection Agency, Research Triangle Park, N. C., January 1974.
3. Martin, G. B., and E. E. Berkau, "Evaluation of Various Combustion
Modification Techniques for Control of Thermal and Fuel-Related Nitrogen
Oxide Emissions," presented at the Fourteenth Symposium (International)
on Combustion, Pennsylvania State University, August 1972.
4. Barrett, R. E., S. E. Miller, and D. W. Locklin, "Field Investigation
of Emissions from Combustion Equipment for Space Heating," EPA R2-73-084a
(API Publ. 4180), Environmental Protection Agency, Research Triangle Park,
N. C., June 1973.
5. Dickerson, R. A., and A. S. Okuda, "Design of an Optimum Distillate Oil
Burner for Control of Pollutant Emissions," EPA-650/2-74-047, Environ-
mental Protection Agency, Research Triangle Park, N. C., June 1974.
6. Hagel, J. A., "Una-Spray Integrated Furnace," New and Improved Oil Burner
Equipment Workshop, Tech. Publ. 108 ED, National Oil Fuel Institute,
New York, N.Y., September 1970.
7. Howekamp, D. P., and M. H. Hooper, "Effect of Combustion Improving Devices
on Air Pollutant Emissions from Residential Oil-Fired Furances," Paper No.
APCA 70-45, Annual Meeting of the Air Pollution Control Association, St.
Louis, Mo., June 1970.
141
-------
8. Turner D. W., et al., "Efficiency Factors for Domestic Oil Heating Units,"
Addendum to the Proceedings of the Conference on Improving Efficiency
in HVAC Equipment and Components for Residential and Small Commercial
Buildings, Purdue University, Lafayette, Indiana, October 1975.
9. Bailey, F. W., "Blue Flame Burner Development at the Blue Flame Labora-
tory," New and Improved Oil Burner Equipment Workshop, Tech. Publ. 108 ED,
National Oil Fuel Institute, New York, N. Y., September 1970.
10. Cooper, P. W., and C. J. Marek, "Design of Blue-Flame Oil Burner Utilizing
Vortex Flow or Attached-Jet Entrainment," API Publ. 1723-A, American
Petroleum Institute, New York, N. Y., January 1965.
11. Locklin, D. W., et al., "Design Trends and Operating Problems in Combus-
tion Modification of Industrial Boilers," EPA-650/2-74-032, U. S. Environ-
mental Protection Agency, Washington, D. C., April 1974.
12. Peoples, G., "Sealed Oil Furnace Combustion System Reduces Fuel Consump-
tion," Addendum to the Proceedings, Conference on Improving Efficiency
in HVAC Equipment and Components for Residential and Small Commercial
Buildings, Purdue University, Lafayette, Indiana, October 1974.
13. Steam/Its Generation and Use, Chapters 3 and 14, 38th Edition, Babcock
and Wilcox Co., New York, N. Y., 1972.
142
-------
APPENDIX A
WARM AIR OIL FURNACE COMPUTER MODEL
An analytical model of the thermal behavior of a warm air residential oil
furnace during cyclical operation has been formulated and programmed for
solution on a Honeywell 440 Timeshare computer. This appendix contains: a
summarization of the model formulation, including assumptions, limitations
and capabilities; a listing of the computer program; and examples of the
program's computer printout.
The computer program is entitled WAFURN (warm air furnace). Given a furnace's
design data, input fuel oil and combustion air flowrates and conditions data,
warm air flowrate, ambient temperature conditions, cycle timing, etc, it marches
in time through two or more furnace operational cycles and computes the unit's
thermal behavior. Principal output data are temperatures of various components
and flow streams as functions of time and values of net heat input,* heat trans-
ferred, and heat stored or lost during one of a series of nearly-identical cycles.
Program input data and all computations are in the modernized metric (SI)
system of units; to facilitate interpretation of computed results, a portion
of the printout is also given in English engineering units.
FURNACE SYSTEM MODELLED
A schematic illustration of the furnace components and flows modelled by the
WAFURN computer program is given in Fig. A-l. The furnace consists of a housing
surrounding a combustion firebox and a heat exchanger, an oil burner, a combustion
gas flue system, and a warm air supply and removal system.
*Net heat input is based on the fuel's lower (net) heating value.
143
-------
Flue Gas
Barometric
Control
Recirculated
Flue Gas
Fuel Oil
Combustion Air
Burner
Warm
Air
Furnace
Housing
Refractory
Return Air
Figure A-l. Schematic of Furnace Modelled by WAFURN Computer Program
144
-------
The warm air is the furnace cooling medium; it flows within the housing,
over the external surfaces of the firebox, heat exchanger, and perhaps part
of the exhaust flue and removes heat from them; it is circulated through the
residential living spaces and returned as not-so-warm air to the furnace. Two
important warm-air components not shown in Fig. A-l are a filter to remove
airborne household dust and a blower to drive the flow of warm air. The filter
does not enter into the WAFURN calculations and the blower is considered only
in the cycle timing control sense of when the specified volumetric flowrate of
warm air is started and stopped.
The oil burner is assumed to be supplied with a specified volumetric flowrate
of a No. 2 distillate fuel oil and with the correct flowrate of combustion air
to support combustion at a prescribed stoichiometric ratio. It is further
assumed to be an efficient burner which causes the input reactant streams to
be mixed thoroughly and reacted completely to thermodynamic equilibrium com-
bustion products soon after their entry into the firebox. A typical No. 2
oil having a net (lower) heating value of 4.273 x 107 J/kg (18,387 Btu/lb)
is built into the program. It may be supplied as a water emulsion (by speci-
fying a nonzero mass fraction of water in the fuel) and its supply temperature
may take on any reasonable value. Similarly, the combustion air supply tempera-
ture and percentage humidity are independent input parameters. As sketched in
Fig. A-l, the combustion air may be mixed with a quantity of externally-
recirculated flue gas by simply specifying the fraction of FGR. Because
the flue gas temperature is a time-varying parameter, the heat content of
the mixed combustion air/FGR reactant stream is time-varying and, as a result,
so is the adiabatic flame temperature. Condensation in the FGR circuit is
assumed not to occur, so the combustion gas composition is maintained at
thermodynamic equilibrium for the dry combustion air/dry oil stoichiometric
ratio whether or not there are humidity in the air, water in the oil and/or
FGR; if one or more of these are present, they are treated as simple diluents
which act to reduce the flame temperature.
145
*M608-B-13 REV. 10-73
-------
The firebox is assumed to have a right circular cylindrical configuration and
to consist of a steel shell with or without a refractory lining for its interior
surface. One end of the cylinder is closed, the other communicates with the
heat exchanger. Configurational details such as tunnel- or side-fired
mounting of the burner and vertical or horizontal orientation of the firebox
are not dealt with explicitly; if they are influential, they must be accounted
for in the expressions for convective film coefficients. Thus the firebox
design information employed are rather limited in scope; they include inside
and outside diameters, total weights of refractory and steel shell, interior
and exterior surface areas for heat transfer, and interior and exterior cross-
sectional areas available for gas and air flows.
Consideration of the heat exchanger is similarly rather vaguely nondescript!ve
concerning design details. As will be seen, the convective exchange coeffi-
cients included in the heat transfer model assume that the heat exchanger is
one or more firetubes with external air crossflow. The user is given the
responsibility for translating his actual heat exchanger into a consistent
set of design parameters for that heat transfer model, or for changing to a
more appropriate model.
The flue system is assumed to be fitted with a weighted-damper barometric con-
trol device for automatic regulation of firebox draft. Draft considerations
enter the model calculations only when the burner is turned off. Then, a
natural draft flow of air is drawn in through the burner, firebox, etc., and
passes out the flue. Draft air flowrate is determined by the draft pressure
in the firebox and by the effective open area of the combustion air supply
louvers in the burner's blower shroud. This effective area is given as a
specified constant at unity stoichiometric ratio and is thereafter treated
as a variable proportional to actual stoichiometric ratio. The firebox draft
is calculated using conventional stack effect formulae, but is divided into
two piecesa firebox-to-heat-exchanger stack effect and a heat-exchanger-to-
barometric-control stack effect.
146
'RM 608-B-13 REV. 10-73
-------
THERMAL ANALYSIS
The thermal analysis approach taken in the WAFURN computer model is essentially
an input, output and storage accounting method. Input reactants provide a
known flowrate of a combustion gas stream having known composition, temperature,
heat content, specific heat, etc. to the firebox. While flowing through the
firebox and, subsequently, the heat exchanger, heat is transferred from that
gas stream to these components. Any residual heat content left in the gases
as they are exhausted up the flue constitutes a thermal loss. Meanwhile, a
more-or-less simultaneous flow of warm air is passed over the outside of the
firebox and heat exchanger; heat is transferred from the components to the
airflow. Heat taken up by this airstream represents the useful portion of the
heat supplied.
If the heat transferred to a particular furnace component differs in magnitude
from that being transferred from it, that component's heat content and, there-
fore, its temperature change. Such changes constitute heat storage. Depending
upon its ultimate disposition, storage may become part of the useful heat or
part of the lost heat. To minimize the impact of time-varying storage terms
on the overall accounting of cyclical heat disposition, WAFURN is structured
to analyze sequential cycles until the total heat storage during a cycle is
less than one percent of that cycle's input heat.
The only thermal losses presently accounted for in WAFURN are the flue gas and
draft air losses. While there are certainly natural and forced convective
heat transfer from the warm air to the furnace housing as well as radiant
transfer from the firebox and heat exchanger to the housing, conventional
practice is to insulate the inside of the housing sufficiently well that these
losses are limited to less than one percent during steady-state operation.
Although continued losses after burner cutoff will magnify that value by
perhaps as much as three times, the total furnace setting losses were believed
to be small enough and well enough understood that their omission would not
impair the utility of the WAFURN solutions.
147
B-13 REV. 10-73
-------
Heat Transfer Models
As indicated earlier, heat transfer analyses in WAFURN are based on a rather
generalized approach that does not get very involved in specifics of component
design. The heat transfer models incorporated as subroutines in the program
were adopted from Chapters 4 and 14 of Ref. A-l .
The rate of heat transferred from the combustion gases to the inside surface
of an unlined firebox or of a metal heat exchanger was calculated by means of
the following expressions appropriate for heat transfer inside cylindrical
tubes. The heat given up by the combustion gases at a given instant is:
Simultaneously, the same quantity of heat is accepted by the inside surface
of the furnace component
= U Asurf
where
U = U + ir
+ U (J/s-m2-K) (A-3)
- [2Tsteel + (^ + (Tout)A
- 30.385 Fpp F, G°g8/ ^^ (A.5)
= M + 0.58 (XH J 1/0.195 + 2.79 x 10"5 A?
L \ H2UlcgJ\ s
7500/ATsurf (A-6)
148
-ORM 608-B-13 REV. 10-73
-------
I
Hg+ Hogl/l2""*'
Gcg = *cg /(Acs>inside
U = K (0.0572 AT
0.096
AT
1m
leg " I0"* C9
[Meg ~Tsteel]
0.8
- 47.3)
A ItM. - H/
Kr ' °-57< D1ns1de
/P]'/'
(A-7)
(A-8)
(A-9)
(A-10)
(A-ll)
This system of equations is solved in subroutine HEXHT by an iterative
process for the heat transfer rate, Q, and the combustion gas exit tempera-
ture, (1"out)cq. Note that the gas stream and metal component temperatures
are treated as bulk mean averages, with no accounting for axial or radial
gradients. The same subroutine is used also for unlined firebox and heat
exchanger cooling by draft air.
For the refractory-lined firebox, essentially the same system of equations
is used but the temperature at the refractory's inside surface is also
included as a variable to be solved for in the iterative solution procedure
in subroutine RFBHT. Simplifying assumptions have been made in order to
approximate transient heat conduction through and heat storage by the
refractory. When the burner is turned on, the inside surface temperature
is assumed to reach very quickly a value corresponding to that it will have
later during steady-state operation. It is this pseudo-steady-state tempera-
ture that is solved for in subroutine RFHBT. Heat transferred from the
149
"08-B-13 REV. 10-73
-------
combustion gases to the refractory is conducted radially outward as a thermal
wave. It is assumed to be stored in the refractory such as to establish a
radial thermal gradient consistent with the eventual steady-state distribution
T Tstee1 1n (r/rsteel) .
Tsurf " Tsteel ln (rsurf/rsteel }
In other words, until that temperature profile is established in the refractory,
no heat is transferred to the firebox's steel shell. Thereafter, steady-state
conduction is assumed, with the conduction rate maintaining consistency with
both the inside refractory and outside steel temperatures.
When the burner is turned off, the hot refractory is cooled from the inside
by convection to the draft air. Then, a cooling wave propagates into the
refractory, penetrating more deeply as time progresses. It is assumed that
the same form of radial temperature profile (but with reversed slope) is
established from the inside to the penetration radius, and that the tempera-
ture at that penetration point corresponds to that of the earlier steady-state
conduction when the burner was on. Meanwhile heat continues to be conducted
down that same earlier gradient to the steel shell. This somewhat simplified
solution to a rather complicated problem is accomplished by subroutine RFBHT
in conjunction with subroutine REFKUL to define the penetration point. During
this process, radiation from the refractory to the air is neglected, since
the only constituents of air which will absorb radiation are the low concen-
trations of COg and HpO.
For the external cooling of the firebox steel shell and the heat exchanger,
convective heat transfer coefficients for crossflow of a gas outside a bank
of cylindrical tubes is used. A single tube was assumed to represent the
firebox and, in consideration of the flow pattern through the Lennox OF7
heat exchanger, it was assumed to correspond to four tubes. The system of
150
DRM608-B-13 REV. 10-73
-------
equations is very similar to that for heat transfer inside the metal heat
exchanger, except that
n c~\ f\ OQ
U = U = 86.14 F F G . ' / DU<: . .
c uu-'^ pp a air i outside
Fpp = 0.0695 + 8.19 x 10"5 ATS(Jrf - 395/AT^ (A-14)
G D °'167
Fa o.587 air. outside (A-15)
and all other parameters are defined for air and conditions outside the tubes.
This problem is solved iteratively for the rate of heat transfer to the warm
air stream and for the output temperature of the warm air by subroutine AIRHT.
The foregoing heat transfer calculations are carried out by subroutines
RFBHT, HEXHT, and AIRHT at each time step under the assumption that the
firebox metal and heat exchanger metal temperatures are known. Following
their completion, the difference between the heat transfer to and heat
transfer from each component is used to calculate a new component metal
temperature for the next time step. Temperature gradients in the metal
are not considered.
Cycle Thermal Efficiency
As the thermal analysis proceeds in time, a number of heating terms are
integrated and, after the cycle time is reached, are used to calculate
thermal efficiency and the distribution of losses. Cycle thermal efficiency
is defined as the sum of heat accepted by the warm air divided by the net
heat input to the furnace. As noted earlier, losses accounted for are the
heat transported out of the furnace by the incompletely-cooled combustion
gases and by the draft air flows.
151
-------
PROGRAM LISTING AND SAMPLE PRINTOUTS
Table A-l is a listing of the WAFURN main programs and all of the nonlibrary
subroutines and function subprograms which it calls. It is included fn this
report simply as a matter of record. Although it obviously has not been
documented well enough for formal delivery, the design and operational
data on the first two pages are easily understood, so the program certainly
could be used by others.
Following the program listing is Table A-2, a reduced-size reproduction of
a portion of the computer printout for the 4 minute firing, 12 minute cycle
of the Lennox OF7 furnace at 1.0 ml/sec dry oil.
REFERENCES
A-l. Steam. Its Generation and Use, 38th Edition, The Babcock and
Wilcox Co., New York, N. Y., 1974.
NOMENCLATURE
2
A = area, m
cp = heat capacity at constant pressure, J/kg-K
D = diameter, m
Fa' Fpp' FT = emPirical coefficients in convective heat transfer equations
G = mass flux, kg/s-m
= empirical radiant heat transfer coefficient
= combustor static pressure, N/m2
= partial pressure of water, N/m2
Q = heat transfer rate, J/s
r = radius, m
T = temperature, K
AT - temperature difference, K
152
t>RM 608*6-13 REV. 10-73
-------
2
U = heat transfer coefficient, J/s-m -K
w = flowrate, kg/s
y
= mass fraction water vapor
_o
u = viscosity, kg/m-s (centipose*10 )
[-] = mean value of parameter
Subscripts
air pertains to warm air
c pertains to convective heat transfer
eg pertains to combustion gas
in, out pertain to inlet, outlet conditions
inside, outside pertain to surfaces of components through which heat is being
transferred
1m pertains to logarithmic mean temperature difference
r pertains to radiant heat transfer
steel pertains to metal shell of firebox or heat exchanger
surf pertains to surface of firebox or heat exchanger material
153
-------
TABLE A-l. LISTING OF NAFURN COMPUTER PROGRAM
2 -
bAFURN
WATURN CONTINUED
OIL-FIRED WARM AIR FURNACE THERMAL ANALYSIS PROGRAM
cn
L. P. C0M6S 0/522-198 EXT 5014 6 JAN 1975
CINPUT PARAMETERS CORRESPOND TO LENN0X 0F7 FURNACE*>
DATEC3)* TSAT(IO>*bRSAT* TTSR<8>*TTEOC8>
IOOC
IECC
HOC
I60C
I70C
180C
800 DIMENSION
B20C
840 DATA TSAT/400* 460. 490* 500, 510* 520* 525*
£60 DATA bRSAT/0.0, .0023* .0034* .OOS|. .0077*
C80i .0157* .0187. .0224'
300C
320 DATA TTSR/I.O* 1*2* 1.4* l>6» 1.8* 2.0* 2*2* 2.4X
340 DATA TTEO/2280.. 2105., 1914.. 1754., 1630.. 1512.«
360* 1425., 1340./
380C
400 CALL DATIME
420C
440C TIME-LINE PARAMETERS
530,
Oil*
535* S40/
.0131*
460
480
500
510
S15C
520
540
S60C
580C INPUT C0ND1TI0NS F0R 0IL i
T0N -0.0
TFAN1 1.0
TOFF 4.0
HITIME
(IF TOFF
TCYC «
DELT
12-0
IS LESS THAN TCYC,
12*0
0.1
SET HITIME EfiUAL TO TCYC.)
AIR
600 FV0L I.OE-06
620 FTEMP 273.2
640 FRHO 84S.
660 CPOIL I860.
680 HVN « 4.E73E»07
700 CPbL ' 4190.
720 FXH20 « 0*0
740C
760 SRI 1.20
780 CATMP 273.2
800 CAHUM 0*6
geoc
840 HAVOL » 0.5
60 WATI * 293.2
880 bAHUM 0.3
900 NAT8FF 320*
920C
940C FRACTION OF FLUE GAS RECIRCULATED
960 RECIRC « 0.0
980C
tOOOC DESIGN PARAMETERS FOR BURNER * FURNACE
1020 CAREA 0.0013
1040
1060
1080
1100
1120
1140
1160
1180
1200
I220C
1240
1260
1£80
1300
1320
1340
I360C
1380
1400
1420
1440
1460
1480
ISOO
I5EOC
I540C
1560
1580
1600
1620
I640C
I660C
1680
1700
1720
1740
1760
1780
1800
1820
1840
1860
1880
1900
1920
1940
I960
1980
2000
2020
HFBHE 0.50
HHEBC 1.5
DIFB 0*2635
AIFB 0.3068
TIIFB 450.
bRFB 1.88
CSIFB 0.0545
OCR 0.065
CPR 837.
DSFB - 0.3048
ASFB 0.365
TISFB 375.
bSFB » 4.. 54
CS0FB <> 0. 143
CPSFB » 500.
DIME 0. 14
AIHE £.25
TIHE 300.
bSHE <> 21.0
CSIHE 0.085
CS0HE 0.088
CPSHE 500.
ADDITIONAL DATA REQUIRED
TAMB * 273.2
DSR 0.3
NSR 4
IPRD 0
DERIVED PARAMETERS
SR SRI
FWO i> FV0L*FRHO
FbDW FWD*FXH20
FWDO FWD*(1. - FXH20)
CAWRU XTRK 1.8*CATHP*TSAT«bRSAT>*CAHUM
CALL CPAIR(CAhRW*CACA.CACB,CACC)
KAWRU XTRPt 1.8*bATI,TSAT,hRSAT)*UAHUM
CALL CPAIR
WAXK * WAbRb/d. » bAbRb)
bARHO 6364. 9//C2.*eCR>
TFUNC /
-------
TABLE A-l. (Continued)
- 3 -
- 4 -
bAFURN CONTINUED
bAFURN CONTINUED
CAL0G(D1FB/DSFB>* CALL HIFIRE<2.HITIME.TCYC.FWD.FWDW.FWD0.CAWD.CAWDW,
2111* CGbDW.CGWD.CGGFB.CGGHE.CREF.OINPUT)
2120 TMIFB a TIIFB
2140 TMSFB a TISFB
2160 TMHE a TIHE
2180 CAWD a (J. » CAbRb)*l4.49*SR*FbD0
2200 CAXb a CAWRb/CJ. * CAWRW)
2220 CAWDW a CAWD*CAXW
2240 CAV0L a I.570E-04*CAWD*CATMP*<18.02 * IO<94*CAXU>
2260 CGWRW a 0*14604 - 0.09205»SR * O.OI934*SR»*2
2280 CCWDW a FbDb * CAWDW * CGbRb*(FbD0 « CAbD*(I.-CAXW))
2300 CGWD a CFbD » CAWD)/(|. - RECIRO
2320 CGXb « CGbDU/(FtaD * CAbD)
2340 CGGF6 a CGWD/CSIFB
2360 CGGHE a CGWD/CSIHE
2380 CGTE6 a XTRPCSR*TTSR*TTEO)
2400 OREF a (FWD0*CP0IL » FUDW»CPWL)*(FTEMP - 296.2)
242C* » CAWD*CPKEAN*
2620* CPMEAN(CGTX*298«2*0.40S2*|.3654E-04*0«0)>/((FWD *
2640* CAbD)*CPMEAN(CGTEC.CGTX.CGCA.CGCB.CGCC» ,
266C 10 CONTINUE
£670 CALL HIFIRE(I.HI TIME.TCYC.FWD.FWDW.FWC0.CAWD.CAWDW,
2671* CGWDb.CGWD.CGGFB.CGGHE.CREF.CINPUT)
2680 IF(SR.NE.SRI) GO TO 100
2700C
2720C PRINT INPUT DATA
2740 PRINT 15*
2760 PRINT 20* DATE
2780 15 FORMATC///)
2800 20 FORMAT(4X*"b ARM
2820* "CYCLE
2840* "ANALYSIS DATEI "»3A6/>
2860 PRINT 25*
2880 25 F«RMAT<2X."FURNACE DESIGN DATAl FIREBOX".9X*
2900* "FIREiOX".10X."HEAT"/28X."REFRACTORV".8X»
2920* "STEEL"»9X."EXCHANGER")
2940 IFUR.LE.O) GO T0 35
AIR 01 L-F 0 R N A C E
ANALTSI S'V/23X.
2960
298 Ot
3000
3020C
30404
3060*
3080
3100
3120*
3140
3160*
3180*
3200*
3220
3240
3260*
32801.
3300*
3320
3340*
3360
3380*
3400*
3420*
3440*
3460*
3480
3500
3520*
3540*
3560*
358 OC
3600
3620
3640
3660
3680*
3700
3720
3740
3760
3780
3800
3820
3840
3860
3870
3871*
3880
3900
PRINT 30. DIFi.DSFE.DIHE.WRFB.WSFi.WSHE.AIF6.ASFB.AIHE.
CSIFB.CSIHE.CS9FB.CS9HE.TMIFB.TMSFB.TMHE
30 F0RMATCIOX»"DIAHETER.« I".1P3EI5.4/14X."MASS»KG»".
3EI5.4/9X."H.T. AREA.M2J".3EI5.4/9X."INT. C«S..M2l~.
E1S.4.I6X.EI5.«/9X»"EXT. C.S..M2I".16X.2EI5.4/
7X."INIT. TEMP..K I",OPF13.2.2F16.2/)
G0 T8 45
35 PRINT 40. DSFB.DIHE.WSFB.WSHE.ASFB.AIHE.CSIFB.CSIHE.
CS9FB.CS8HE.TMSFB* THHE
40 F8R«AT'HOISTURE*KG/St"*E15.4.16X.E15.4/SX.
"INIT. TEMP..KI".OPFI3.2.2F16.2//2X,"RECIRCULATED FLUE"*
" GAS*Zt"*2PF9«2//2X*"0UTD00R AMBIENT TEMP..KI"*OPF9»2/>
PRINT 60. T0N* T9N. TFANI. TCYC* T0FF* WAT0FF. CELT
60 F0RHAT<2X."CYCLE TIMING (MINUTES)I CYCLE".10X."BURNER"*
9X."W.A. FAN"/I8X*"0N I".OPF13.2.2F16.2/16X,"flFFl".
F13.2*F16.2»7X»"«"»F6.|»" K"//BX»
"MARCHING INTERVAL.MINI".F6.3)
100 C0NTINUE
JCYC a TCYC/DELT
SOFSR a O'O
IFUR.EC.l) SOFBR a WRFB*CPR«
-------
TABLE A-l. (Continued)
- 6 -
WAFURN CONTINUED
WAFURN CONTINUED
01
SOFL
OF SRI
QFBSl
OHESI
CGTl
CGT2
CGT3
TFAN2
T « DELT
IPR 30
JPRI 0
IRC 0
NHI « 4
-60.*flREF*(TOFF - T9N)
S8FBR
* sores
SOHES
CCTX
CGTX 100.
400.
T9FF » 5.
3980
3940
3960
3*80
4000
4020
4040
4060
4080
4100
4120
4140
4150
4I60C
4I80C PERFORM TRANSIENT HEAT TRANSFER ANALTSISCISTl F.B.. 2ND! H
4EOO DO 400 J*l«JCrC
IF(T.LE*TFAN2> GO TO 108
(PR 10
IF(T«LE«(TFAN2 * 2.*DELT» JPRI J/1PR
106 OELT » 60«*DELT
HESMIN <> WSHE»CPSHE*WAT2
lFtT.CT.TFAN2.9R.CCT2.l.T.WAT2> HCSH1N WSHE*CPSHE*CGT2
ODFBS 0.0
ODF6G 0*0
1FCT.GT.T9FF) CGTl CATHP
1F(T«LE.H1TIME.OR.NHI.GT.4> G9 T8 109
CALL HIFlRE(NHI*HITIME*TCYC»FViD*FbCti*FbDO>CAWD*CAHDb»
CGVDW*CGbO*CGGF6«CGGHE*QREF*OINPUT>
109 CONTINUE
E.)
4220
4240
4C60
4280
4300
4320
4340
4360
4380
4385
4390
439 U
4395
4400 IF(T.LE.T0N.0R.T*GT.TOFF> 09 T9 120
4420 CGTl ((FWD*CAbD>*CPMEANC298.2*CGTX«CGCA*C6CB*CGCC**
4440* (CGTX-298.2) » RECIRC»CGWD**
4460* CPNEAN(298.2*CGT3»CGCA*CGCB*CGCC»/(CGbD*
4480* CPMEAN(29S.2*CGTI*CGCA*CGCB*CGCC>> » 298.2
4SOOC BURNER ON* UNL1NED FIREBOX
4520 IFUR.NE.O) GO TO 110
4540 CAUL HEXHT(SR*CGTI»CGGFB»TMSFB*AlFB*DIF6*CGhD*CGXb*C6CA*
4560* C6CB*CGCC*CGT2»QDFBS*IPRD>
4580 SOFBS S0FBS * DELT*ODF6S
4600 Gfl TO 140
46£OC BURNER 8N< REFRACTflRr-LINED FIREBOX
4640 110 CALL RFBHT OOFBS « CDFBG
4720 SOFBR « SOFBR * DELT*(ODFBG - ODFBS1
4740 SOFBS SOFBS * DELT*6DFBS
4760 G8 T« I4O
4780 120 CALL DRAFTtSR.CAREA.HFEHE»HHEBC.CGTJ.CGT2.CGT3*
4800* CATMP.CAXW,DAWD>
4820 DAGFB DAbD/CSIFB
4840 DAGHE DAWD/CS1HE
4860 IF<1R.NE.O> G« TO 130
4880C BURNER OFF* UNLINED FIREBOX
4900 CALL HEXHT(-I»CATMP,DAGFB.TMSFB»AlF6»DlF6.DAKD»CAXh.CACA.
4920* CACB.CACC»CGT2»ODFBS.IPRD>
4940 IF(ODFBS.GT.O..aR..GT.FBSHlN) GO TO 125
4960 CPDA CPMEAN(CATMP.CGT2.CACA.CACB»CACC>
4980 CGT2 CGT2 * (SOFBS»OELT*ODFBS-FBSHlN>/
5000 OOFBS /DELT
5020 125 S6FBS SSFBS * DELT*ODFBS
5040 GO TO 140
S060C BURNER OFF. REFRACTORT-LINED FIREBOX
5080 130 IRC IRC » I
5100 IFURC.GT.I) GO TO 135
5120 TMP TMSFB * «WA1I-THSFB>/TFUNC
5140 RRP RRFB
5160 COEFF WRFB*CPR
5180 135 CALL REFKULURC»TMSFB,DSFB.TMIFB.DIFB.AlFB»SeFBR.COEFF.
5200* TMP.DP.AP.RRP)
5220 CALL RFBHTt-I.CATMP.DAGFB.TMIFB.TMP.AIFB.AP.D1FB.DP.RRP.
5240* DAWD.CAXW.CACA»CACB.CACC* CGT2.ODFBG.IPRD)
5260 ODFBS CTHP-TMSFB)*ASF6/RRFB
5280 SCFBR « SOFBR » DELT*(flOFBG-ODFBS)
5300 SOFBS SOFBS * OELT*ODFBS
5320 140 CONTINUE
S340C FIREBOX HEATING OF WARM AIR
5360 IF GO TO 180
5380 CALL A1RHTCWATI.WAGFB.TMSFB.ASFB.DSFB.WAWD.WACA.WACB.WACC.
5400* WAT2.0DFBA.IPRD)
5420 S6FBS SOFBS - DELT*ODFBA
5440 180 CONTINUE
5460C (CALL F.B* REGION RAO'N LOSS SUBR.)
5480 IFCT.LE.T9N.9R.T.6T.TOFF) GO TO 200
S500C BURNER ON* HEAT EXCHANGER COOLING COMBUSTION GASES
5520 CALL HEXHT.GT.HESHIN> GO TO 210
5680 CPDA CPMEAN
5700 CGT3 CGT3 » (SOHES * DELT*ODHES - HESMIN>/(DELT*DAUD*CPDA>
5720 ODHES " (SOHES - HESM1N>/DELT
5740 210 SOHES SOHES * DELT*QDHES
5760C HEAT EXCHANGER HEATING OF WARM AIR
5780 220 IF(T.LE.TFAN|.OR.T.CT.TFAN2) GO TO 260
5800 CALL AIRHT(VAT8«bAGHE*TMHE*AIHE*DlHE*toAUD'UACA»MACB»l
-------
TABLE A-1. (Continued)
- 7
- 8 -
VAFURN C0NTINUED
bAFURN CONTINUED
cn
5820* bAT3»ODHEA*1 PRO)
5840 1F(ODHEA.LE.O..OR.(SOHES-DELT*ODHEA).GT.HESMIN) GO TO 240
5860 CPbA a CPMEAN(bAT2.bAT3*kACA.bACB»bACC)
5880 WAT3 « bAT3 CSOHEA-DELT*ODHEA-HESMIN)/
.6*CGTX - 459.7
6*TMIFB - 459.7
.8*TMSFB 459.7
5960
6000
6020
6040
6060
6080
6100
6120
6140
6160
6180
6200
6220 280 FORMAT(////2X«"STOICK. RATI0 a"*OPF5.2>"*
TMHE <
THSFB
ITPR1
ITPR2
ITPR3
ITPR4
ITPR5
ITPR6
ITPF7
ITPR6
6*TMHE
.6*CGTI
.8*CGT2
6*CGT3
S*bAT3
459
459
459
459
459
IFtNCYC.EC.O) GB T9 32C
IF(J.EC.l) PRINT 260* SP.CCTX.ITPR1
6240*
6260*
6280*
6300*
6320*
6340
6360
6380
6400
ADIAB. FLAME "
"TEMP. <*»F7. I." K"/44X*I6*" F"/14X*
"CURRENT AVERAGE TEMPERATURES* DEGREES K **.
*
6500 G« T0 320
6520 310 PRINT 315. T»TMSFB.TMHE»CGT I.CGT2.CGT3. WAT3.
6540* ITPR3«ITPR4»ITP.R5»ITPR6«ITPR7«ITPR8
6560 315 F0RMAT
6580 320 CONTINUE
6600 IF(T.GT.TFAN|.AND«T.LE.TFAN2) SOUA a SOWA * DELT*WAUD*
6620* CPHEANbACB.bACC)*(WAT3 - UATI)
6640 1F*
6660* CPMEAN(CCT3.£98.2.CGCA.CGCB.CeCC)»(CGT3-298.2>
6680 IF(T.LE.T0N.9R.T.GT.T9FF) SODL ' SODL * DELT*OAbD*
67004 CPMEAN(CGT3«CATMP>CACA*CACB*CACC>*CCGT3-CATMP)
6720 IF(T.GT>TOFF.ANO.TFAN2.GT.T.AND.bAT3.L£.UAT0FF) TFAN2»T
6740 DELT DELT/60-
6760 T T » DELT
6780 400 CONTINUE
6800C
6820 OSTFBR a SOFBR - CFBRI
6840 OSTFBS a SOFBS - OFBSI
6860 OSTHES a SOHES - OHESI
6880 OSTTOT ° OSTFBR * OSTFBS * OSTHES
6900 fiSUM a OSTTOT * SOWA » SODL * SOFL
6920 OFRST a OSTTOT/OINPUT
6940 IF(NCYC.GT.O) GO TO 4|0
6960 IF(OFRST.LE.O>0|.eR.ITER.E0.3) NCYC a HER * I
6980 GO TO 105
70COC
7020C CALCULATE * PRINT THERMAL EFFICIENCY DATA
7040 410 ETATH a |00.*SObA/OINPUT
7060 ETADL a |00>*SODL/OINPUT
70SO ETAFL a |00>*SOFL/OINPUT
7100 OINBTU a' OINPUT/I05S.OS6
7120 KABTU a SObAX1055.056
7140 PRINT 415. NCYC '
7160 415 FORMAT(/IOX*"PRINTOUT IS FOR CYCLE NO."*I2)
7180 PRINT 420* OINPUT.OINBTU.SCWA.ETATH.UA6TU*OSTT0T*
7200* 6FRST*SOFL*ETAFL*SODL*ETADL
7220 420 FORMATC/" CYCLE THERMAL RESULTSl HEAT* J ".
7C40* " EFFICIENCY*V3IX*""*IOX*"PERCENT"/8X."NET ".
7260* "HEAT INPUTl"5X. 1PEI0.4/3CX.'C".OPF7.0»")'V6X*
7280* "WARM AIR GAIN I",5X,IPEI0.4,OPF13.2/30X."(".F7.0*
7300* ")"/8X."STORED OR LOSTI-/12X."COMPONENTS".6X. IP
7320* EI0.4*2PFI3.2/I2X*"FLUE GASES".6X.IPE10.4*OPFI3>2/
7340* IEX."ORAFT AIR"*6X*1PE10«4*OPFI3.2)
7360 SR a SR » DSR
7380 500 CONTINUE
7400 END
7420
7440
7460
8000 FUNCTION XTRP(TX*T*0)
8020C
8040C PERFORMS LINEAR INTERPOLATION
806 OC
8080 DIMENSION T(t>*0(l>
8100 OXaO(l)
8120 IFCTX.LE.T+<0(I2>-0
-------
TABLE A-l. (Continued)
- 9 -
- 10 -
WAFURN CONTINUED
fcAFURN CONTINUED
01
00
340
8360
6500 FUNCTION CPHEAN(TI*T2* A.6.O
6S20C
8540C CALCULATES MEAN GAS SPECIFIC HEAT (J/KG-K) BETWEEN
8560C TEMPERATURES Tl AND T2 (K)
8580C
8600 CPHEAN 4186.8*CA * B*(T2 » Tt>/2. * C/>
6620 RETURN
8640 END
8660
8680
8700
9000 SUBROUTINE CPAlR
90COC
9040C CALCULATES COEFFICIENTS A* B* C FOR MB1ST AIR SPECIFIC HEAT
9060C CPMCCAL/G-K) « A > B*T * C/T**2 AT TEMPERATURE T
9080C
9100 FAC ' !/22I9 » 0.40S2*bRH>
9140 CB FAC«l«OE-05«C2.602 » 13.654*bRH>
9160 CC FAC*<-384.)
9180 RETURN
9200 END
9220
924O
9260
9500 SUBROUTINE CPC6CSR*SXH.CA*CB*CC>
9S20C
9S40C CALCULATES COEFFICIENTS A* 6. C FOR EOUILIBRIUM DRY OIL/
9S60C DRY AIR COMBUSTION PRODUCTS WITH ADDED MOISTURE
9S80C CPMCCAL/G-KI » A » B*T » C/T**2 AT TEMPERATURE T. TA. TB<8>» TC<8>
9640 DATA TSR/I.O* I.2. 1.4. 1.6* 1.8* 2.0* 2.2* 2*4/
9660 DATA TA/ .2538. .2491* .2454* .2425* .2403* .2335* .237* .2358/
9680 DATA TB/ 4.23* 4.02* 3.83* 3.69* 3.57* 3.47* 3.39. 3.33X
9700 DATA TC/-II58.-I102.-1006.-926.-862.-8CV.-765.-730/
9720C
9740 FAC I. - SXW
9760 CA FAC*XTRPTA> » 0.4052*SXb
9780 CB FAC*I*OE-05*XTRPCSR»TSR*TB> » I.36S4E-044SXU
9800 CC » FAC*XTRPtSR.TSR.TC)
9820 RETURN
9840 END
9860
9880
9900
10000 SUBROUTINE HEXHT
10010 DIMENSION TG(25)*DTC(2S>*TSR(8>*TPRC8>
CK SORT<0.574*D*XTRPCSR.TSR.TPR»
T6C1) TM * I.
|.SSE»OS*t.B*TF -
I0020C
I0030C SOLVES HEAT TRANSFER INSIDE METAL HEAT EXCHANGER
I0040C
10050 DATA TSR/1.0. 1.2* 1.4* 1.6* I>8* 2>0* 2*2. 2.4/
10060 DATA TPR/.236I* .2136* .1868* .1649* .149. .1331* .1224*
10070* .I1I6/
I0080C
10090 CK 0.0
10100 IFCSR.GE.I.)
10110 10 DELG 10
10120 TG(|> » TOUT
10130 IF
IOI40C
10150 DO 50 J-1.24
10160 TAV » (TIN » TGCJ»/2.
10170 TF /2.
10180 FT CTAV/TF)**0.8
10190 FPP <|. * 0.58*XW)»(0.|95
10200* 24300./**2>
10210 UG 20.2S7*FPP*FT*GG**0.6/D**0.2
10220 UG l«5*UG
10230 IF CSR.LE.O.) GO TO is
10240 TRATI0 « (TIN TM)/(TG - TM>
10250 IF(TRATI0«GT.O.O> DTLM (TIN-TG(J»/ALOG(TRATI0>
10260 UR CK*(0«OS72*DTLN » 0.096*TM - 47.3>
10270 USUM UG * UR
10280 IFCUR.GT.O.) UG USUM
10290 15 CPM CPMEAN(TIN*TG(J>*CA*CB*CC>
10300 TDUM (TIN**UG*A*TM>/.LE«ABS<6MAXM GO TO 20
10340 CDUM OMAX
10350 TDUM TIN > «DUM/(M*CPM>
10360 20 CONTINUE
10370 IF(IPR.LE.O> GO TO 40
10380 IFCJ.EO.I) PRINT 22*
10390 22 FORMATf SUBROUTINE HEXHT"/" J TG
10410 PRINT 32* J»TGCJ>*TDUM*CPM*ODUH
10420 32 FORMAT(I4*1P4E10O>
10430 40 CONTINUE
10440 DTG(J> TG(J> TDUM
10450 IF GO TO 100
10460 CALL TADJ(J*2*DTG«TG*DEL6*XI*Y|*X2*Y2*S1*S2*OS*IF>
10470 SO CONTINUE
10480 100 ODOT ODUM
10490 TOUT TDUM
10500 RETURN
10510 END
TDUM"*
-------
TABLE A-l. (Continued)
- ii -
bAFURN CONTINUED
WAFURN CONTINUED
10520
IOS30
10540
11000 SUBROUTINE RFBHT(SR«TIN*GG»TMI.TMS*Al.AS*DI*DS*RR*k*XW*
11010* CA,CB.CC.TOUT.OD8T.IPR>
I1020C
II030C SOLVES HEAT TRANSFER INSIDE REFRACTORY-LINED FIREBOX
I1040C
11050 DIMENSION TG<25)*TI<25>»DTG(25>*DTI(2S>«TSR(8>*TPR(8>
II060C
11070 DATA TSR/1.0. 1.2* 1.4. |.6* 1.6* 2.0* 2.2* 2-4/
11060 DATA TPR/.236I* .£136* .1868* .1649* .149* .1331*
11090* .1224. . IU6/
IIIOOC
IIIIO CK 0.0 -
11120 IFCSR.GE.I.) CK SORT THI » |.
11170 TICI)- TMI
01 11180 lie o
VO 11190 III 0
II200C
lieiO DO 50 Ja|*£4
11220 JLI J 1
11230 JLG « J « 1
11240 TAV a » TAVI/2.
11260 FT *(O.I9S » I.8*t.S5E-05*TF -
11280* 24300./C!.6*TF>*a2>
11290 UG a 20.257*FPP*FT*GG*»0.8/DI**0.2
11300 UG |.S*UG
11310 IF GO TO 15
11320 TRATIO (TIN -TMI)X(TGCJ) - THI>
11330 1F(TRATI0.GT.O«0> DTLH (TIN-TG(J»/AL8G(TRAT10>
11340 UR - CK*(0»0572*DTLM * 0.0964THI -47.3)
11350 USUH UG * UR
H360 iF(UR.GT.o>) UG USUM
11370 IS CHI a CFMEANCTIN*TGCJ)*CA*CB*CC>
11380 TDUH CTIN*(W*CPM-UG*Al/2.)»UG*AI*Tt
11400 CMAX fc«CPH»CTIN - THI)
11410 lF
11440 20 CONTINUE
11450 TDUMI a IMS » ODUHI*RR/AS
11460 IF GO TO 40
11470 IF CJ.EO.l) PRINT 32*
11480 32 FORMAT*" SUBR0UTINE RFEHT"/" J TG TDUM".
11490* " CPM ODUMI TI(J) TDUMI")
11500 PRINT 42* J.TGCJ>.TDUM»CPM»OOUHI*TICJ)»TDUMI
IISIO 42 F0RMAT(I4>IP6EI0.4)
11520 40 CONTINUE
11530 DTG(J) a TGCJ) - TDLM
11540 DTI(J) :' T1CJ) - TDUMI
11550 IF Ill a |
11580 IFUII.NE.O) JLG a J - |
11590 IFCHG.EO.O) CALL TACJ< J*JLG*DTG*TG«Dt*XI*Y I*X2*T2*SI.S2*02*N2>
11600 IF(IPR.GT.O> PRINT 52*
11610* - J*N2*SI.S2.02.X|.Y|.X2.Y2*TG*TG(J*I>
11620 52 FORMAT(2I4*OP3F4.0/OP6EI|.4>
11630 IF(IIG.EG.l) TG
11640 IF(Ul.ES.O) CALL TADJt J. JLI »DT1. TI.D3.X3. Y3.X4. Y4.S3. S4.04*N4>
11650 IF PRINT 52.
11660* J*N4*S3*S4*04*X3*Y3.X4*Y4*Tl(J>*TKJ»l>
11670 IF(III.EO.|> TI(J»1) a TKJ)
11680 IF(IIG.EO.I.AND.IIl.EO.I) GO TO 100
11690 SO CONTINUE
11700 100 OD0T a ODUMI
11710 TOUT a TDUH
11720 THI a TDUMI
11730 RETURN
11740 END
11750
11760
11770
12000 SUBROUTINE AIRHT/2.
12260 FPP <> 0.0695 * I.8*4.S5E-05«TF - 1279./**2
12280 UA a S7.429*FPP*FD*FA«(GA**0.61/D**0«39>
12300 UA a |.5*UA
12320 CPM a CPHEAN(TIN*TA(J)*CA*CB*CC>
12340 TDUM a (TIN*(b*CPM-UA*A/2.)*UA«A*TH>/(U*CPM*UA*A/e.)
12360 ODUN a W»CPM«
-------
TABLE A-l. (Continued)
- 13 -
- 14 -
WAFURN CONTINUED
VAFURN CONTINUED
12380 CMAX fe*CPM*CTH - TIN)
12400 IFCABS.LE.ABS> G» 19 SO
18420 OOUM a QMAX
12440 TDUM TIN » ODUM/CWaCPM)
12460 £0 IFCIPR.LE.O) G0 TO 40
12480 IFCJ.EO. |> PRINT 22*
I2SOO 22 FORHATC" SUBROUTINE AIRHT"/~ J TACJ) TDUM"
12520* - CPU CDUM")
12540 PRINT 32* J. TACJ), TDUM. CPM.CDUM
12560 32 FORMATU4. IP4E10.3)
12560 40 CONTINUE
12600 DTACJ) TACJ) - TDUM
12620 IFCABSCDTACJ».LE.|.) GO TO 100
12640 CAUL TADJCJ.2.DTA.TA.DELA.XI.Y1.X2.Y2.SI.S2.0S.IF)
12660 SO CONTINUE
12680 100 CDBT a OOUH
12700 TOUT a TDUM
12120 RETURN
12740 END
12760
12780
12800
13000 SUBROUTINE TADJC I.IL.DE.T.DT.X I, Y I.X2. Y2.S I.S2.0S. IFP)
I3020C
I3040C METHOD 0F FALSE POSITION FOR ITERATION C0NVERGENCE
I3060C
13080
13IOOC
13120
13140
13160 IFCDEC I).LT.O.) SI
13180
13200
13220
13240
13260
13280
DIMENSION DEC25).TC25)
GO T0 20
IFCI.GT.l)
SI a i.
IFCDEC I).LT.O.)
OS a -SI
T<2> a TCI) » SI*DT
XI « TO)
Yl a DEC I)
IFF a o
GO TO 100
13300 20 S2
13320
13340
13360
13380
13400
13420
IFCDEC i >.LT.O.> se a - i
IF IFP a J
1FCI.GT.IL.AND.IFP.E0.1)
X2 a TCI)
Y2 a DEC I )
60 TO SO
13440 40 IFCS2.NE.0S) GO T0 60
13460 XI TCI)
13480 Yl a DEC I)
13500 G0 TO 80
13520 60 X2 a TCI)
13540 Y2 DEC I)
60 TO 40
SUBROUTINE REFKULC1RC«TS»DS»TI*DI» A1»SO»C»TP»DP»AP»RR)
CALCULATES REFRACTORY COOLING PARAMETERS
GO TO 40
13560 BO B CY24XI - YI*X2>/CXI - X2>
13580 SM a CYI - B)/XI
13600 TCI* I) « -B/SM
13620 IFCABSCDECI)).GT. 10.) TCI*I> a CTCI»1) * TCI»/2>
13640 100 RETURN
13660 END
13680
13700
13720
14000
I4020C
I4040C
I4060C
14080 DD a CDS - DD/IOO.
14100 JLIM a 100"
14120 IFCIRC.GT. I)
14140 SURF a TP
14160 SHELL a TS
14180 HEAT " SO
14200 FRR ' RR
14220 DP a Dl * DD
14240 JL0W a |
14260 JLIM « 10
I4280C
14300 40 DELTAO a HEAT - SO
14320 DO 60 JaJLObtJLIH
14340 IFCJ.GT.JL0b) DP « DP *
14360 JL0W a J
14380 CR a C*CDP*«2 - OI**2)/CDS**2 - Dl**2>
14400 T a SHELL * CSURF - SHELL)*ALOGCDP/DS>/ALOGCDI/DS>
14420 DO a CR»C SURF-SHELL)*CCDP*C ALOGCDP/DS)- I . )-DI»C AL0GCDI/DS)- ! )
144401. CCDP-DI>*AL0GCDI/DS)» * CR*C SHELL - CT+TD/2.)
14460 IFCIRC.GT. I. AND. DO.GE. DELTAO) GO T0 60
14480 60 CONTINUE
14500 80 TP a T
14520 AP a A1+DP**2/DI**2
14540 RR a FRR*DP*ALOGCDP/'DI)/CDS«ALOGCDS/DI))
I4560C
14580 100 RETURN
14600 END
14620
14640
14660
ISOOO SUBROUTINE DRAFTC SR. A.H I.H2.T I. T2.T3. TA.X..WD)
I5020C
I5040C CALCULATES DRAFT AIR FLOwRATE INTO FIREBOX
I5060C
15080
15100
15120
DD
H a
TB1
TB2
I./CC1. - XU)/28-96 » XW/18'02)
a CTI » 2.*T2 » T3»/4.
a CT2 » T3 » 2.«TA>/4.
-------
TABLE A-l. (Continued)
- 15 -
bAFURN CONTINUED
15140
15160
15180
15200
15220
15240
15260
15280
15300
15320
15340
15360
16000
I6020C
16040
16060
16080
16100
16120
16140
16160
16180
16200
16220
16240
16260
16280
16300
RH91
RH02
RH0A
DPI «
DPS :
WD B
12.|93*H/TBI
12.|93*M/T62
I2.|93*M/TA
9.80S*HI»RH0l*/
= 9.605*H2*RH02*(T2+T3-2.*TA>/(T2*T3»2.*TA>
o.o
IF«DPI*DP2).GT.O.O> UD .6*A*SR*SQRT<2.*RH0A*
C »
1F 69 T9 40
Fl
F2
F3
Cl
C2
C3
C4
CS
C6
OI*(TH + l.5*< TC - TH»/TC
«2* 60 T0 90
16440 C " 1.5
16460 N 5
16480 90 Fl C*FOI
16500 F2 C*FD2
16520 F3 C*FD3
16540 Cl C*CD1
16560 Ce C*CD2
16580 C3 C«CD3
16600 C4 C*CD4
16620 C5 C*CD5
16640 C6 C*CD6
16660 100 RETURN
16680 END
-------
TABLE A-2. SAMPLE TABULATION FROM THE WAFURN COMPUTER PROGRAM FOR A
12-MINUTE FIRING CYCLE IN A LENNOX MODEL OF7 FURNACE
CONFIGURATION
W fl R n R I R 01 L-F U P N A C E CYCLE ANALYSIS
ANALYSIS DATEI DEC 05. 1974
FUFtlrtCE DESIGN DflTBl F1FEVOX F1PEKJX HEAT
PEFPftCTOPY STEEL EXCHANGER
DlftrKTER.il I i.eJIOE-Ot 3.046UE-OI t .4000E-OI
MAtS.f.Gl l.6eO('E«00 4.S400E«00 2.IOOOE*OI
H.T. AFEA.H2I 3.0cfcOE-01 3.630(-OI ».2500E»00
IHT. C.S..H2I 3.4SOOE-02 2.SOOOE-02
EXT. C.S..M2I 1.4300E-OI 9.9000E-02
IN1T. TEMP..t 1 430.00 377.00 300.00
FB-HE HT. - 0.30 M. HE-BC HT . I .SO Hi C.R. APEA - 0.00130 H2
UAPtl AIR FLOWPATE .MS'S I 0.3000E«00 INLET TEMP . .K 1 £93.30
tG-S: 6.0076E-01
tnPtJT FUEL i COMBUSTION A1R<9 SR«1.2>!
EMULSION DRY OIL COMB. AIR
FLOUPATE >H3'S 1 l.OOOOE-06 1.1396E-02
VG'Si 8.4300E-04 8 .4500E-04 1 .4723E-OS
NO I STUPE -MS/Sl O.OOOOE'OO 3.261IE-09
INIT. TEMP..KI 373.20 273.20 273 .SO
BECIPCULATED FLUE GAS.*! 00.00
OUTDOOR AMBIENT TEMP.iKl Z73.20
:VO.E TIRING
TIME.
HIM.
0.10
0.20
o.so
1.00
1.50
2.00
2. SO
3.00
3.50
«.oo
3.00
6.00
7.00
8.00
9.00
10.00
11.00
12.00
FlPEtOX FIREBOX
PEFFACT. CTEEL
1777.8 4*5.9
£740 378
1823.4 463.9
2822 378
1823.3 463. 9
2822 378
1823.3 463.9
2622 378
1812.9 389.3
2803 240
1804.9 347.7
2789 IK
1800.3 "324.3
2781 124
1798.3 311.3
2777 1 00
1796.9 303.7
2774 86
1796.1 299.3
2773 79
1269.1 334.7
-1824 142
1122.2 366 .£
1360 199
1033.9 392.4
1404 246
930.7 414.6
1231 286
894.0 433.7
1130 320
832.9 430.3
1 073 330
811.3 464.7
1000 376
784.1 477.4
931 399
PRINTOUT IS FOR C'
CYCLE THEPflAL PESULTS I
NET HEAT INPUT!
WARN AIR GAIN 1
STORED OP LOSTl
COMPONENTS
FLUE GfilES
DRAFT AIR
HEAT FLAHE
EXCHANG. ZONE
343.7 2080.6
162 3283
261.3 2080.8
191 3283
408.4 2080.8
273 3283
484.6 2080.8
412 3283
416.4 2080.8
289 3283
392.6 2080.8
247 3283
383.6 ' 2080.8
230 3283
379.8 2080.8
223 3283
378.1 2080.8
£20 3283
377.3 2080.3
219 3283
330.7 273.2
133 32
330.7 273.2
133 32
330.3 273.2
133 32
330.4 273.2
133 32
330.3 273.2
134 32
330.2 273.2
134 32
330.0 273.2
134 32
329.9 273.2
134 32
rCLE NO. «
FIREBOX
EXIT
1993.9
3133
1993.0
3127
1993.0
3127
1993.0
3127
1989.3
3121
1986.9
3116
1983.3
3114
1984.7
3112
1984.2
3111
1984.0
3111
316.9
110
312.7
103
309.9
93
307.0
92
304.9
89
303.3
86
301.7
83
300.6
ei
FLUE
GAS
370.0
36*
379.4
383
608.2
633
636.1
721
628.0
670
610.0
638
£02.9
629
600.4
621
399.3
61*
399.8
618
326.0
127
324.3
124
323.4
122
322.3
120
321.3
119
320.9
117
320.2
116
319.7
US
UARH
AIR
319.2
114
319.2
114
319.2
114
319.2
114
373.1
211
393.9
177
346.6
164
343.9
198
342.1
IS*
341.4
194
319.3
119
319.3
119
319.3
119
319.3
119
319.3
119
319.3
IIS
319,3
119
319.3
IIS
STOICH.
RATIO 1.30. ADIAB. FLAME TEMP. -
1809.6 K
2797 r
CUP-PENT AVEPAGE TENPEPATUPES DEGREES K
TIME.
HIN.
0.10
0.20
0.30
1.00
l.SO
2.00
2.30
3.00
3.30
4.00
3.00
6.00
7.00
e.oo
9.00
10.00
11.00
12.00
FIFEEOX FIPEI-OX
PEFPACT. STEEL
1317.7 436.8
2272 326
1337.3 436.6
2343 32b
1337.6 436.8
2344 326
1337.6 436.8
2344 326
1347.8 373.6
2326 212
1940.3 339.0
2313 ISO
1336.2 319.3
2303 113
1533.8 308.4
2301 93
1332.4 302.0
2298 83
1331.6 298.3
2297 77
1097.8 327.6
1316 129
976.1 333.6
1297 176
892.9 379.2
1147 213
834.3 393.3
1042 248
788.0 409.3
998 277
733.7 422.9
896 301
719.6 434.8
833 322
697.0 443.2
794 341
HEAT FLAME
EXCHAHG . ZONE
342.6 1809.6
136 2797
337.6 1809.6
183 2797
401.8 1809.6
263 2797
473.3 1809.6
392 2797
408.7 1809.0
273 2797
386.3 1809.6
236 2797
378.3 1809.6
221 2797
373.0 1809.6
213 2797
373.3 1809.6
212 2797
372.8 1809.6
211 2797
328.7 273.2
131 32
328.3 £73.2
131 32
328.4 273.2
131 32
328.2 273.2
131 32
328.1 273.2
130 32
327.9 273.2
130 32
327.7 273.2
130 32
327.6 273.2
129 32
FIPEBOX
EXIT
1749.3
2689
1747.1
2689
1747.1
2689
1747.1
2689
1744.6
2680
1742.9
2677
1741.9
2679
1741.3
2674
1741.0
2674
1740.8
2673
310.3
98
306.9
91
303.6
86
301.4
82
299.6
79
298.2
77
296.8
74
293.8
72-
FLUE
GAS
603.4
633
617.8
692
646.4
703
693.1
787
669.3
737
647.6
706
640.9
693
638.9
689
637.4
687
636.9
686
322.0
119
320.3
117
319.3
119
318.4
113
317.6
112
317.0
110
316.4
109
319.9
106
UflRN
AIR
317.8
lit
317.8
lit
317.1
112
317.8
lit
367 .9
201
349.8
170
343.2
IM
340.3
133
339.3
130
338.7
149
317.8
112
317.8
lit
317.8
lit
317.8
112
317.8
lit
317.8
lit
317.8
lit
317.8
lit
PRINTOUT IS FOR CYCLE NO. 4
HEAT. J EFFICIENCY.
l»TU>
8.3401E*M
< 7903. >
6.860IE»06
< 6302. >
6.8034E»04
1.3762E*06
3.4779E*04
PERCENT
CYCLE
THERMAL RESULTS!
NET HEAT INPUT!
62.23
0.82
16.30
0.42
W
nRn AIR GAIN 1
STORED OR LOSTl
COMPONENTS
FLUE GASES
DRAFT AIR
HEAT. J EFFICIENCY.
llTU>
8.3684E»0*
< 7932.)
6.4201E»06
< 6083.)
4 .9988E*04
l.6610E»06
3.7972E*04
PERCEMT
76 .72
«.60
22.24
0.49
!M608-B-13 REV. 10-73
162
-------
TABLE A-2. (Continued)
STOICH. RATIO - I.6
mint. FLAHE TEMP.
CUPREMT AVEPAGE
TIME. FIFEtOX FIPEION H£AT
Mirt. REFPliCT. STEEL EXCHAMG.
415^
287
415.2
X67
415.2
287
413.2
287
361.9
191
332.3
138
313.6
108
306.3
91
300.8
81
297.*
76
322.3
120
344.2
139
362.4
192
377.9
1609.4 K
2430 f
0.10
1928
0.20 1339.1
1986
0.50 1360.1
1988
1.00 1360.1
1988
1.30 1351.1
1972
£.00 1343.9
1959
2.30 1339.8
1931
3.00 1337.4
1947
3.30 1336.1
1943
4.00 1333.3
1943
971.1
1288
866.1
1102
797.8
976
3.00
6.00
7.00
8.00
».00
10.00
11.00
12.00
748.4
887
709.4
817
680.3
765
631.7
713
632.7
679
220
391.2
244
ZONE EXIT
339.7 1603.4 1360.7
131 2430 2349
333.8 1603.4 1539.1
177 2430 2346
393.6 1605.4 1309.1
232 2430 2346
462.6 1605.4 1339.1
372 2430 2346
401.6 1603.4 1537.3
263 2430 2343
380.9 1603.4 1336.0
223 2430 2341
373.4 1603.4 1355.3
212 2430 2339
370.4 1603.4 1334.9
206 2430 2339
369.1 1605.4 1334.7
204 2430 2338
368.% 1603.4 1334.5
203 2430 2338
273.2 303.2
32 89
326.3 273.2 301.7
128 32 83
326.4 273.2 299.1
127 32 78
273.2 297.2
127 32 73
326.0 273.2 293.6
127 32 72'
325.8 273.2 294.4
126 32 70
273.2 293.1
32 67
273.2 292.3
32 66
326.7
128
326.2
GAS
623.8
663
632.9
679
660.6
729
703.3
809
678.4
7*1
661.3
730
655.2
719
652.9
713
651.9
713
631.4
712
318.3
113
317.0
110
319.9
toe
313.1
107
314.3
106
316.4
10*
31*.4
109
316.4
10*
316.4
10*
362.3
198
346.1
1*3
340.1
132
337.7
148
33*.*
14*
336.1
US
31*.4
10*
316.4
10*
316.4
10*
31t.4
10*
31*.4
. 10*
402.r
263
412.e 323.6
283 126
421.4 325.4
298 123
313.2
103
3U.7
103
316.4
109
31*.4
10*
STOICH.
Tint.
niH,
0.10
0.20
0.30
l.OO
1.90
c.oo
£.30
3.00
3.30
4.00
s.oo
.00
7.00
e.oo
9.00
10.00
11.00
12.00
RATIO 2.10. APIA*. FLAME TEMP. -
CUWEMT AVEPAGE TEMPERATURES
FlPtrDX FIPE10X. HEAT FLAME
REFPACT .
1176.9
1638
1204.1
1707
1203.1
1709
1203.1
1709
1197.7
1696
1190.8
1683
1187.2
1677
1183.1
1673
1183.9
1671
1183.2
1669
849.8
1069
764.9
917
706.9
812
666.3
739
634.3
682
610.9
639
386.9
996
963.9
934
STEEL
391.6
243
391.6
243
391.6
24-5
391.6
243
349.0
168
323.2
123
311.7
101
303.9
87
299.4
T9
296.8
74
318.1
112
336.2
143
331.1
172
363.8
195
374.6
214
384.1
231
392.2
246
399.4
239
EXCHANG.
343.2
138
336.6
182
395.8
232
458.3
363
396.3
234
376.1
217
368.8
204
366.0
199
364.9
197
364.4
196
331.7
137
331.3
136
331.2
136
330.9
133
330.7
133
330.4
134
330.1
134
329.8
133
zone.
1443.8
2139
1443.8
2139
1443.8
2139
1443.8
2139
1443.8
2139
1443.8
2139
1443.8
2139
1443.6
2139
1443.8
2139
1443.6
2139
273.2
32
273.2
32
273.2
32
273.2
32
273.2
32
273.2
32
273.2
32
273.2
32
1443.8 K
2139 F
. DECREES K
FIP£»OX FLUE MARTI
EXIT
1409.3
2077
1408.3
2073
1408.2
2073
1408.2
2075
1407.0
2072
1406.1
2071
1403.6
2070
1403.3
2069
1405.2
2069
1405.1
2069
300.1
80
297.1
75
294.8
70
293.1
7
291.7
65
290.7
63
289.6
61
288.5
99
CAS
629.9
674
638.6
689
664.8
737
706.8
812
678.6
761
661.7
731
633.7
720
633.3
716
632.6
714
632.2
714
319.3
113
317.9
112
316.8
110
316.0
109
313.3
107
314.7
t-oe
314.1
103
313.3
104
AIR
319.9
116
319.9
It*
31*. 9
116
319.9
116
338.9
186
343.0
137
337.3
147
335.1
143
334.1
141
333.7
140
319.9
116
319.9
116
319.9
116
319.9
116
319.9
116
319.9
116
319.9
116
319.9
116
PRINTOUT IS FOR CYCLE NO. 4
CYCLE THERMAL RESULT SI
MET HEAT IMPUTl
UARM AIR GAin I
STOPED OR LOST I
COMPONENTS
FLUE GASES
DRAFT AIR
HEAT J
<«TU>
8.3588E»06
< 7923.>
6.0I33C»06
< 5701.)
3.64336*04
2.2e77E»06
4.0I26E*04
EFFICIENCY.
PERCENT
71.96
0.44
27.13
0.48
PRINTOUT IS FOR CYCLE HO. 3
CYCLE THERMAL RESULTS
MET HEAT INPUTl
IMRfl AIR 6AIM I
STORED OR LOST I
CDnPOMEMTS
FLUE GATES
DRAFT AIR
HEAT. J
4.3732£*04
2.ol27E*06
4.586£E*04
EFFICIENCY.
PERCENT
67.49
0.53
31.42
o.ss
163/164
-------
APPENDIX B
FLUE GAS COMPOSITIONAL ANALYSIS
The sample flow train used for analyzing flue gas composition is illus-
trated in Fig. B-l. A 0.006 m (1/4 in) diameter stainless steel tubing
sample probe was inserted near the combustor centerline, downstream of
the heat exchanger. Flue gas aspirated into the sample probe flowed
through a line to an air-cooled condensibles trap where particulates and
heavy oils were separated out. Next, the gas passed into an ice-cooled,
stainless-steel condensibles trap where most of the water and any condensible,
low-volatility hydrocarbons were removed, After the condenser, the gas
passed into a pyrex wool-filled glass cylinder which served as a final
separator for heavy oils and particulates, and provided a visual indication
of the cleanliness of the gas being admitted to the analysis instruments.
Table B-l gives a summary of the gas analysis instruments used. The gas
leaving the glass wool filter was split into three parallel paths. One path
led directly to the total hydrocarbon analyzer. A second path led through
a Drierite bed where water was removed, and then into the series-plumbed
carbon monoxide, carbon dioxide, and oxygen analyzers. The third path
o
passed through a combined Drierite and 3 A molecular sieve bed for total
water removal, and then into the nitric oxide analyzer. The gas was
pumped through the system by three diaphragm pumps located downstream of
the nitric oxide analyzer, total hydrocarbon, and carbon monoxide + carbon
dioxide analyzers. The system also has the capability for dilution of gas
passing through the carbon monoxide, carbon dioxide, and oxygen path. The
165
-------
SAMPLE
LINE
SMOKE
SAMPLE
GLASS
WOOL
FILTER
CD-
SIGHT
TUBE
GLASS
STAINLESS STEEL
CONDENSIBLES TRIP
ICE-COOLED
CONDENSIBLES'
TRAP AIR-COOLED
MOLECULAR SIEVE 3A*
+ INDICATING DRIERITE BED
3-WAY SELECTOR
VALVES
ROTAHETERS
DILUENT..,
AIR IN ""]
AIR PURIFIERS
DUAL DILUTION
SYSTEM
VENT
TO
HOOD
DRIERITE
BED
CO
ANALYZER
/ \
\ /
C02
ANALYZER
u
INAL
I
ALL LINES 1A-INCH STAINLESS STEEL TUBING (THIN WALL)
Figure B-l Analytical system for fuel oil burner emissions analysis
-------
Table B-l. EXHAUST ANALYSIS INSTRUMENTS
Type
Range
Sensitivity
Calibration
CO
MSA
Nondispersive IR
LIRA
Model 300
0 to 1500 ppra
(nole)
30 ppm minimum
detectable
1000 ppm CO in
NL standard gas
co2
MSA
Nondispersive IR
LIRA
Model 300
0 to 20 mole %
0.25% minimum
detectable
14% C02 in N2
standard gas
NO
MSA
Nondispersive IR
LIRA
Model 200
0 to 500 ppm
(mole)
10 ppm minimum
detectable
0.82% C2H4 in
N- used as
simulant for
410-ppm NO
standard
Total HC
MSA
\\2 flame
ionization
detector
0.2 to 800 ppm
total HC by
volume as CH.
10 ppm minimum
detectable
3% CH4 in helium
used as a
standard
Oxygen
Beckman
polarographic
0« to 100'°
-0.1%
Air - 21%
N2 = 0'.
Smoke
Bacharach
(manual)
0 to 9
1
Ten spots of
monotonically
varying
darkness
-------
dilution allowed use of those analytical instruments on gas samples more
concentrated than the highest range of the instruments. It was achieved
by admission of air metered through parallel rotameters.
When the analytical system shown in Fig. B-l is used to analyze gases which
may have been quenched before combustion was completed, there are two factors
that must be considered in reducing the data: (1) only burned or partly
pyrolyzed fuel is included in the analysis, since minute quantities of liquid
or vapor fuel may be removed by the cold trap, and (2) water formed from hy-
drogen and oxygen during the combustion process is also removed from the
analyzed sample by the cold trap.
Values calculated from the measured flue gas compositional data included:
the overall stoichiometric ratio, the weight of nitric oxide per unit weight
of burned fuel, and the weight of carbon monoxide per unit weight of burned
fuel. The method of calculation to obtain these values is described below.
The calculations were based on air having the following nominal composition:
Component Mole Percent Wt. Percent
N2 78.08 75.63
02 20.95 23.19
Noble gases (Ar.He & Ne) 0.94 1.13
C02 0.03 0.05
100.00 100.00
168
-------
The composition of the fuel was assumed to be characterized by the formula
CH where, for the No. 2 fuel oil burned in this program, x = 1.814. The
/\
following symbols were used in the calculations:
AIR = moles of air to produce 100 moles of dry flue gas
FUEL = moles of fuel to produce 100 moles of dry flue gas
CO = moles of carbon monoxide in 100 moles of dry flue gas
C02 = moles of carbon dioxide in 100 moles of dry flue gas
NO = moles of nitric oxide in 100 moles of dry flue gas
$2 = moles of oxygen in 100 moles of dry flue gas
HC = moles of hydrocarbon, as CH^, in 100 moles of dry flue gas
The values of CO, C02, NO, 02, and HC were obtained directly from the analysis
instruments. In the following, it is assumed that all hydrogen is oxidized
to water and condensed out of the system at the cold trap, prior to analysis.
An oxygen balance yields:
0.2095 AIR = C02 - 0.0003 AIR + 0.5 CO + 0.25x (C02 + CO -
0.0003 AIR) + 0.5 NO + 02 (B-l)
The left hand side of the above equation represents the total free oxygen
contributed by the air. The first two items on the right side represent
moles of oxygen tied up in C02, less the amount of C02 originally present
in the air. The third term represents moles of oxygen tied up as carbon
monoxide. The fourth term represents oxygen consumed to oxidize hydrogen,
yielding the water condensed out in the cold trap. The fifth term is the
oxygen tied up in nitric oxide. The sixth term is free oxygen remaining
169
-------
in the sample reaching the analysis treatments. Eq. (B-l) can be arranged
to yield:
(1 + ) CO, + (1/2 + T) CO + 1/2 NO + 09
AIR = - - - - - - - -
0.2095 + 0.0003 + 0.0003 x/4 (B-2)
A carbon balance can be used to calculate the moles of burned fuel per
100 moles of dry flue gas:
FUEL = C02 - 0.0003 AIR + CO (B-3)
The moles of air available per mole of burned fuel in the sample gas can
be obtained by taking the ratio of the values from Eq. (B-2) and (B-3).
AIR must be calculated first, before calculation of FUEL. If the combustion
were in stoichiometric proportions, the moles of air would be, by an oxygen
demand calculation:
AIR = (1 * x/4) FUEL (B-4)
A1Kstoich 0.2095 l^;
The stoichiometric ratio of the locally sampled burned gases is a param-
eter frequently used in this report. It is defined as the ratio of AIR
to AIRstoich:
SR = AIR (B-5)
AIRstoich
Combination of Eq. (B-2) through (B-5) yields a direct calculation of the
burned gas stoichiometric ratio in terms of the measured parameters:
(1 + £) C02 + (1/2 + £) CO + 1/2 NO + 02
SR = 0.2095 + O.Q003 + 0.0003X/4
0 + ft
0.2095
(1 + T) CO, + (1/2 x f) CO + 1/2 NO
C00 + CO - 0.0003 ~ *
2 uu " u-uuuo 0.2095 + 0.0003 + 0.0003X/4
(B-6)
170
-------
According the the above definition, when the sample contains just a suffi-
cient amount of air to oxidize all of the fuel in the sample to COg plus
condensed-out water, then SR = 1. As a second example, if there is twice
the required amount of air for complete oxidation of the fuel, then SR = 2.
Note that the stoichiometric ratio, as calculated from Eq. (B-6) does not
require that the products in the flue gas be in chemical equilibrium.
Note that the accuracy of the stoichiometric ratio calculation would be
affected very little if all terms in Eq. (B-6) containing the factors
0.0003 and NO were ignored. These factors represent the carbon dioxide
originally present in free air, and the oxygen tied up in nitric oxide,
respectively.
One partially questionable assumption made in the formulation of Eq. (B-6)
was that all hydrogen originally present in the fuel becomes oxidized to
water and is removed in the cold trap. This was a necessary assumption,
since there was no instrument available to measure the actual hydrogen
content of the sample gas. The assumption is very good under the combined
conditions of air-rich stoichiometric ratios (SR > 1) and chemical equilib-
rium. To test this assumption, a Rocketdyne thermochemical computer code
was used to calculate the species concentrations under conditions of chemi-
cal equilibrium for stoichiometric ratios from 0.8 to 2.8. These calcula-
tions included the equilibrium presence of free H;>. The actual stoichio-
metric ratios of these combustion gases, compared to those calculated by
Eq. (B-6) (which does not recognize the presence of 1^) are given in Table
B-2, where it can be seen that Eq. (B-6) is quite accurate except for SR < 1.
Calculated equilibrium conditions are tabulated in Tables B-3 and B-4.
171
-------
TABLE B-2. VALIDITY OF STOICHIOMETRIC RATIO CONDITIONS
Actual Stoichiometric Ratio
0.800
1.000
1.200
1.400
1.600
2.000
2.400
2.800
Stoichiometric Ratio Calculated
from Eg. (B-6)
0.844
1.003
1.197
1.400
1.600
2.002
2.404
2.804
The primary cause of the inaccuracy at SR < 1 is the unaccounted for pres-
ence of H2. In nonequilibrium gases, there is likely to be H2 present even
where none would be indicated from equilibrium calculations and, at fuel-rich
conditions, there could be more or less than indicated from equilibrium calcu-
lations. Because of this likelihood of nonequilibrium, no attempt was made
to correct the calculations of Eq. (B-6) by means of equilibrium calculations.
The concentration of C02 (dry basis) in the flue gas is the parameter most
often used in the space heating industry as an indication of combustion
conditions. To illustrate the relationship of %C02 to the Stoichiometric
ratio, equilibrium data from Table B-4 were used to calculate the curve
shown in Fig. B-2; a calculated %02 curve is also shown. A number of
values of measured C02 concentrations in actual furnace flue gases are also
plotted on Fig. B-2. The measured data are seen to be very well correlated
by the calculated equilibrium curve at SR ,> 1.1 (the calculated maximum
172
-------
TABLE B-3. EQUILIBRIUM COMBUSTION GAS PROPERTIES FOR
NO. 2 DISTILLATE FUEL OIL BURNED WITH AIR
(CHn 01,, 18,443 Btu/lb Net Heat of Combustion
1. o!4
with Air at 14.67 psia)
Stoich.
Ratio*
0.8
1.0
1.2
1.4
1.6 Air
2.0 Rich
2.4 1
2.8 T
0.8
1.0
1.2
1.4
1.6
2.0
2.4
2.8
0.8
1.0
1.2
1.4
1.6
2.0
2.4
2.8
Oil + Air
Inlet Temp. ,
F
0
70
200
Flame
Temperature,
F
3429
3614
3290
2940
2649
2209
1897
1663
3778
3649
3336
2991
2703
2765
1955
1722
3867
3709
3418
3085
2802
2369
2061
1831
c
Frozen,
Btu/lb-R
0.346
0.341
0.333
0.324
0.318
0.307
0.298
0.291
0.347
0.341
0.333
0.325
0.318
0.308
0.299
0.193
0.347
0.342
0.334
0.326
0.320
0.309
0.301
0.295
Y
Frozen
1.261
1.254
1.260
1.267
1.275
1.288
1.298
1.308
1.261
1.254
1.259
1.267
1.274
1.286
1.297
1.306
1.260
1.257
1.259
1.266
1.273
1.284
1.294
1.305
Viscosity,
centipoise
0.0666
0.0687
0.0653
0.0615
0.0581
0.0527
0.0487
0.0456
0.0671
0.0691
0.0658
0.0621
0.0589
0.0535
0.0495
0.0464
0.0681
0.0698
0.0668
0.0652
0.0600
0.0548
0.0509
0.0479
Thermal
Conductivity,
Btu/hr-ft-F
0,0702
0.0711
0.0661
0.0610
0.0567
0.0500
0.0452
0.0415
0.0709"
0.0715
0.0667
0.0617
0.0574
0.0509
0.0461
0.0425
0.0720
0.0725
0.0678
0.0629
0.0588
0.0524
0.0477
0.0441
Prandtl
Number
0.7946
0.7984
0.7954
0.7915
0.7880
0.7820
0.7771
0.7730
0.7948
0.7984
0.7956
0.7918
0.7884
0.7825
0.7778
0.7738
0.7951
0.7985
0.7958
0.7925
0.7890
0.7854
0.7-90
0.-"51
Molecular
Weight
27.73
28.80
29.00
29.03
29.03
29.02
29.01
29.00
27.72
28.77
29.00
29.03
29.03
29.02
29.01
29.00
27.71
28.73
28.98
29.02
29.02
29.02
29.01
29.00
*Stoichiometric ratio is unity at 14.49 masses of air per mass of fuel, and
proportionately greater than unity for increasing relative mass of air.
-------
TABLE B-4. CALCULATED EQUILIBRIUM COMBUSTION GAS COMPOSITION, VOLUME OR MOLE PERCENT
Stoich.
Ratio
0.8
1.0
1.2
1.4
1.6
2.0
2.4
2.8
0.8
1.0
1.2
1.4
1.6
2.0
2.4
2.8
0.8
1.0
1.2
1.4
1.6
2.0
2.4
2.8
Oil + Air
Inlet Temp. , F
0
70
200
H
0.0630
0.0397
0.000
0.000
0.000
0.000
0.000
0.000
0.0737
0.0455
0.0000
0.0000
0.0000
0.0000
0.0000
0.0000
0.0964
0.0577
0.0000
0.0000
0.0000
0.0000
0.0000
0.0000
0
0.0000
0.0313
0.0217
0.0000
0.0000
0.0000
0.0000
0.0000
0.0000
0.0362
0.0261
0.0000
0.0000
0.0000
0.0000
0.0000
0.0000
0.0468
0.0356
0.0000
0.0000
0.0000
0.0000
0.0000
Ar
0.821
0.866
0.882
0.890
0.895
0.902
0.907
0.910
0.821
0.866
0.882
0.890
0.895
0.902
0.907
0.910
0.821
0.864
0.882
0.890
0.895
0.902
0.907
0.910
OH
0.0499
0.2816
0.1862
0.0757
0.0790
0.000
0.000
0.000
0.0613
0.3072
0.2082
0.0885
0.0351
0.000
0.000
0.000
0.0878
0.3579
0.2533
0.1157
0.0493
0.0000
0.0000
0.0000
H2
2.016
0.250
0.030
0.000
0.000
O.OCO
0.000
0.000
1.996
0.269
0.036
0.000
0.000
0.000
0.000
0.000
1.964
0.304
0.048
0.000
0.000
0.000
0.000
0.000
H20
12.263
11.690
10.141
8.832
7.799
6.297
5.276
4.541
12.271
11.647
10.121
8.824
7.795
6.297
5.276
4.541
12.275
11.562
10.078
8.806
7.787
6.295
5.276
4.541
CO
7.243
1.393
0.161
0.0203
0.000
0.000
0.000
0.000
7.268
1.501
0.195
0.026
0.000
0.000
0.000
0.000
7.318
1.710
0.270
0.042
O.COO
0.000
0.000
0.000
C°2
8.687
12.052
11.247
9.841
8.679
7.000
5.864
5.046
8.659
11.934
11.210
9.835
8.678
7.000
5.863
5.046
8.604
11.705
11.127
9.816
8.6?2
7.000
5.863
5.046
NO
0.000
0.253
0.390
0.2955
0.2080
0.0829
0.0339
0.000
0.017
0.272
0.404
0.322
0.223
0.096
0.041
0.018
0.027
0.310
0.451
0.373
0.268
0.125
0.05?
0.028
N2
68.837
72.522
73.784
74.465
74.947
75.603
76.028
76.326
68.901
72.456
73.751
74.447
74.933
75.596
76.023
76.323
68.796
73.328
75.683
-4.405
74.905
75.582
76.015
76.319
°2
O.COO
0.619
3.160
5.566
7.444
10.107
11.888
13.161
0.000
O.C66
3.159
5.553
7.432
10.100
11.884
13.159
0.000
0.754
5.162
5.526
-.406
10.085
11.676
15.154
-------
15
o
CJ
c
0)
ai
o.
o
14
13
CO
CD 12
=1
11
10
8
1.0
- COp Data from Ref. B-l
- C0? Data from Ref.
Calculated CO
B-2
Data; Ref. B-3
14
to
12 5
10
8
1.5 2.0
Sto1ch1ometr1c Ratio (A1r/Fuel)
CNJ
o
O)
o
Figure B-2. Flue Gas C02 and 02 Concentrations for
No. 2 Fuel 011 Burned 1n Ambient A1r at 1 ATM
175
-------
C0? concentration as the stoichiometric condition is approached by reducing
excess air is not normally observed 1n furnace testing).
Other parameters of interest for the flue gases are the mass ratio of nitric
oxide to burned fuel, the mass ratio of carbon monoxide to burned fuel, and
the mass ratio of unburned hydrocarbons (as CH, ) to burned fuel . These
ratios are generally expressed herein as grams of nitric oxide per kilogram
of burned fuel (g NO/kg fuel), grams of methane per kilogram of fuel
(g CH,/kg fuel), and grams of carbon monoxide per kilogram of burned fuel
(g CO/kg fuel). These parameters are calculated by aid of Eq. (B-2) and
(D-3) from the following relationships:
g MO (1000) (NO) (MWNQ)
kg fuel
g CO
kg fuel
g HC
(C02 - 0.0003 AIR + CO) (MWp)
(1000) (CO) MWrn
uu
(C02 - 0.0003 AIR + CO) (MWp)
(1000) (HC) (MWCH )
kg fuel (C02 - 0.0003 AIR +
where
MWNQ = molecular weight of NO = 30.01
MWp = molecular weight of fuel
= 12.01 +1.01 x = 13.84
MWCO = molecular weight of CO = 28.01
MWCH = molecular weight of methane = 16.04
For calculation of the above quantities, the term 0.0003 AIR can be neglected
without introducing more than about 0.1-percent error in the calculations,
176
-------
or AIR can be computed from Eq. B-3 and included in the calculation.
The numbers given in this report include the effect of the term. The
experimental data were reduced, according to the above equation, by
means of a remote terminal timeshare computer program.
In addition to the gaseous pollutants described above, the smoke content
of the mixed gases was also measured. The instrument utilized for this
purpose was a Bacharach smoke meter. (It is manufactured by the Bacharach
Instrument Company, Pittsburgh, Pennsylvania.) This is a hand-held device
which, when pumped, sucks flue gases from a 1/4-inch-OD, uncooled sample
probe through a piece of white filter paper; 10 strokes of the pump, over
3
a period of about 15 seconds, causes the passage of 57.2 m of flue gas
o 32
per m of filter paper (2250 in /in ). The smoke particles deposit out
on the filter paper. A reading is taken by comparing the darkness of the
smoke deposition spot to a scale of 10 such calibrated spots provided with
the instrument. The readings vary from 0 to 9. A reading of zero corres-
ponds to no visually detectable deposit on the filter paper, while a reading
of 9 corresponds to a dark black deposit. Intermediate readings are vary-
ing shades of black and gray, increasing in darkness with 'Increasing read-
ing numbers. A reading of 1 is generally accepted by the industry as a
very acceptable degree of smoke. At the opposite extreme, a reading of 9,
which is totally unacceptable, still does not correspond to sufficient smoke
to be easily visible from observation of the exhaust stack outlet. In some
instances reported herein, when the reading was obviously greater than 9,
the number of strokes was halved and smoke spot reading doubled, thus
extending the smoke scale to a maximum of 18.
177
-------
REFERENCES
B-l. Barrett, R. E., S. E. Miller and D. W. Locklin, "Field Investigation
of Emissions from Combustion Equipment for Space Heating," EPA R2-73-084a
(API Pub!. 4180), Environmental Protection Agency, Research Triangle
Park, N. C., June 1973.
B-2. Hall, R. E., J. H. Wasser, and E. E. Berkau, "A Study of Air Pollutant
Emissions from Residential Heating Systems," EPA-650/2-74-003,
Environmental Protection Agency, Research Triangle Park, N. C.,
January 1974.
B-3. Dickerson, R. A., and A. S. Okuda, "Design of an Optimum Distillate
Oil Burner for Control of Pollutant Emissions," EPA-650/2-74-047,
Environmental Protection Agency, Research Triangle Park, N. C.,
June 1974.
178
-------
APPENDIX C
DATA TABULATION: OPTIMUM BURNER EXPERIMENTS
Combustion chamber design parameters, burner operating conditions and flue gas
composition data are tabulated for the optimum burner experiments described and
discussed in Section V. Except where specifically called out as being different,
the nominal burner firing rates were 1.0 ml/s (gph) and experiments were 10 min
on/20 min off cyclical tests. All flue "gas composition data, including smoke,
are cycle-averaged values.
179
-------
LEN.
m
.75
.75
.75
.30
.30
.30
.40
.40
.40
.SO
.SO
' .50
00
O .50
.30
.30
.30
.40
.40
.40
.50
.'so
.SO
RUN S101C. C92
NO. RATIO X
..... o_2?n I.D
s
3
*
S
6
7
8
9
IOC
1 1
ie
13
14
15
16
17
IB
19
eo
21
22
l.ll 13.9
1.25 IS. 6
1.16 13.3
I.IO 13.9
1.12 13.8
1.02 14.3
1.07 14.1
1.15 13.4
1.18 13.6
1.12 13.6
1.12 13.6
1.08 14.1
I.OS 14.5
- - - 0.281 I
1.04 14.2
1.14 13.3
1.04 14.2
1.13 13.6
I.IO 13.8
1.16 13.0
1.06 13.8
1.22 12.6
l.ll 13.9
02 ca NO
I PPM PPM
. , Air-Cooled, Tunnc
2.3 83 46
4.5 57
3.0 53
1.9 464
2.4 505
0.6 >I600
1.6 1124
3.0 776
3.6 447
0.0 133
0.0 118
0.0 335
0.0 » 1600
.0., A1i-Cooled
0.9 k.1600
2.8 k!600
0.9 k!600
2.7 671
2. 1 k 1 600
3.1 «I600
1.3 k!600
4.1 1442
2.2 1192
40
17
33
18
49
26
18
22
30
33
UHC CO
PPM GM/KGH
il-F1red - - - -
7 1.24
24 0.96
21 0.83
70 6.72
206 7.47
1395 t2l.72
127 IS. 99
493 11.83
328 7.00
61 1.97
78 1.76
38 36 4.80
44 96 k22.20
. Tunnel-Fired - - -
30 blO k.22.03
30
35
20
13
8
12
22
19
1804 k24. 12
538 k22.06
222 10.07
1 17 k23.27
161 k24.S2
187 122.43
1751 23.28
315 17.42
NB UHC
GM/KGM GM/KGM
0.734 0.059
n.715 0.227
0.294 0.183
0.513 0.579
0.292 1.741
0.725 10.819
0.405 1.031
0. 309 4. 290
0.369 2.929
0.484 0.515
0.532 0.660
0.588 0.294
0.664 0.761
0.447 4.012
0.486 15.539
0.517 4.239
0.334 1.903
0.207 0.972
0.140 1.410
0.184 1.498
0.381 16.153
0.313 2.629
BACH.
SM0KE
0.0
0.0
0.0
0. 1
1.0
4.0
0.3
0.3
0.3
0.3
0.0
0.0
1.5
1.5
2.0
2.5
O. 5
1.2
1.7
2.0
0.2
1.0
TFO
C
260
250
235
260
280
270
315
310
310
295
310
310
310
157
217
203
174
179
176
186
188
183
LEN HUN STOIC. CB2
n NO. KATIO I
.75
.75
.75
.40 3/4
.40 1/2
.40 1/4
.40 1/4
.40 1/2
.40 3/4
.40 3/4
.40 1/2
.40 1/4
.30
.30
.30
.30
.40
.40
.40
.50
.50
.50
23
24
25
26
27
28
29
30
31
32
33
34
35 C
36
37
38
39
40
41
42
43
44
- - - U.2/«
1.03 14.7
1.14 13.6
1 . IH 13.0
- - - 0.22n
1.12 13.8
l.ll 13.9
l.ll 13.9
1.17 13.2
1.19 13.0
1.20 12.9
1.04 14.6
02 CB N8
X PPM PPM
I.D. , Air-Cooled, Tu
0.7 175 38
2.7 HI
25
UHC CO
PW GM/KGM
2 2.40
46 1.23
NO
GM/KGM
0.570
0.404
3.4 228 II 198 3.57 0.187
I.D., Fractional Insulation, Tunnel-Fired -
2.S 1304
2.2 1064
2.3 1160
3.3 2S3
3.7 210
3.8 234
1.0 k!600
ee
23
eo
24
24
26
40
I.OS 14.5 1.2 k!600 38
1.06 14.3 1.4 k!600 35
- - - - - 0.22» I.O.. Insulated.
I.OS 14.6 1.2 1483 87
1.14 13.5
1.09 14.4
1.12 14.2
1.21 13.1
1.21 12.9
I.OS 14.1
l.ll 13.7
1.22 12.6
1.14 13.4
2.8 96
1.8 192
2.5 9S
3.9 35
4.0 41
1. 1 >1600
2. 3 422
4.0 37
2.8 43
40
47
40
49
46
61
59
50
54
179 19.33
126 15.55
1ST 17.04
184 3.92
204 3.32
192 3.73
326 k22.08
393 k22.24
322 122.50
Tunnel-Fired -
3 20. 67
13 1 . 47
7 2.76
9 1.42
3 0.57
4 0.68
170 » 22.21
2 6.22
3 0.61
5 0.67
0.358
0.361
0.323
0.400
0.416
0.445
0.600
0.574
0. 542
1.302
0.661
0.738
0.641
0.845
0.802
0.918
0.939
0.871
0.890
UHC
GM/KGM
0.016
0.395
1.766
1.516
1.052
1.317
1.630
1.844
1.746
2.571
3. 12V
2.587
0.024
0. 1 1 1
O.O5S
0.073
0.024
0.039
1.349
0.015
0.032
0.048
PACH. TFG
SMBKF. C
0.3 181
0.4 )85
0.2 186
0.2 169
0.2 1W
0.2 171
0.5 *71
O.S 17*
0.5 "8
2.0 18«
2.5 186
2.0 186
3.0 192
0.1 IBS
1.5 192
O.S 189
0.1 189
0.0 192
7.O 175
S.O 192
O. 1 189
O. 5 . 1*4
-------
CD
LEN
m
.75
.75
.75
.75
.75
.50
.50
.50
.40
.40
.40
.75
.75
.75
.30
.30
.30
RUN SIB 1C
NO. RATIB
4$ 1.33
46 1.20
47 1.10
48C 1.41
49C 1.07
SO 1.73
51 1.36
52 1.23
53 1.11
54 1.50
55 1.77
56 1.60
S7 1.32
58 1.17
59 1.60
6O 1.23
61 1.10
C02 02 CB NB
X I HPM PPM
0.22» I.D., Insulated, Tum
11.7 5.6 II 54
13.0 3.7
13.9 2.1
11. 1 6.5
14.1 1.3
- 'OlTnl.O
9.0 9.5
11.4 5.9
12.6 4.8
13.9 2.2
10.2 7.5
8.7 9.8
9.9 8.6
18.0 5.5
13.7 3.4
9.8 8-6
12.6 4.8
12.7 2.0
18 66
184 72
17 59
15 73
., Insulated,
103 46
45 68
155 77
401 81
155 42
406 16
12 64
IS 75
21 89
1153 10
313 »4
k!600 72
UHC CB
PPN GM/KGM
6 0.21
3 0.21
1 «.70
5 0.34
0 0.23
TunfMkl.Flrwd
40 2.42
7 0.81
3 2.54
S 5.89
9 3.12
57 9.71
13 0.28
1 0.28
1 0.34
1 56 24. 67
5 S.IO
158 123.33
NO UHC
GM/KGM GM/KGM
1.041 0.059
1.123 0.024
1.136 0.008
1.194 0.050
I.I IB 0.000
1.162 0.538
1.322 O.O78
1.349 0.033
1.273 0.042
0.911 0.099
0.433 TJ.784
1.491 0.159
1.409 0.010
1.488 0.009
0.244 1.906
0.951 O.OSI
1.135 1.266
BACH.
SMBKE
0.0
0.0
4.0
0.0
0.0
0.3
0.7
4.0
6.0
O.S
0.2
0.0
0.0
0.3
4. 1
4.0
9.0
TFG
C
226
213
209
847
230
217
213 t
213 I
1
202
207
EDO ,
847
234
224
815
187
202 '
LEN
n
.30
.40
.40
.30
.50
.SO
.75
.75
.75
.75
.75
.30
.30
.30
.40
.40
.40
.50
.50
.50
RUN
N0.
62
63
64
65
66
67
68
69
70
71
72
73
74
75
76
77
78
79
80
81
S10IC. C02
RATIB Z
- - - - 0.22m
1.67 9.2
1 . 69 9.1
1.43 11.0
1.38 11.3
1.42 10.8
1.66 9.1
1.44 ll.l
1.67 9.5
09?n
l.b« 9.8
1.12 13.7
1.26 12.0
1.28 12.1
1.11 13.9
1.51 10.1
1 . 28 1 8. 1
1.18 13.8
1.49 10.3
1.49 10.4
1.36 11.4
1.13 13.7
az co NO
I PHM PPM
I.D., Insulated, Tu
9.2 »I6OO 10
9.2 6H3
6.8 76
6.3 275
6.6 28
8.7 130
7. 1 12
6
35
1 1
54
28
67
9.2 22 63
I.D.. Naur-Cooled,
7.7 28 30
2.4 17
4.6 13
5.0 519
2.3 110
7 . 6 i 1 600
4.9 71
2.4 33
7.3 387
7.4 75
6.0 28
2.3 25
50
40 '
10
31
15
24
41
9
IB
30
46
UHC C0
PPM GM/KGM
409 t35.88
88 15.50
2 1.47
28 5.07
1 0.55
9 2.89
10 0.25
23 0.52
Tunnel ~F1 red
10 0.60
0.27
1 0.83
110 8.84
1.62
820 132.20
10 1.22
S 0.50
65 7.72
13 1 . 49
3 0.53
7 0.37
NO
GM/KGM
O.P56
0. 169
0.734
0.820
1.116
0.681
1.397
1.535
0.676
0.806
0.734
0. 194
0.490
0.348
0.438
0.656
0.805
0.393
0.594
0.735
UHC
GM/KGM
5.243
1. 141
0.022
0.290
0.011
0. 112
0. 108
0.894
0. 1 18
0.008
0.010
1.069
0.013
9.431
0.097
0.042
0.744
0. 149
0.027
0.060
BACH. TFG
SMBKE C
2. 1 PS 6
1.0 224
0.8 228
1.3 221
0.0 273
0. 1 873
0.0 288
0.0 298
0.0 170
0.0 168
0.0 168
0.0 811
I.I 210
0. 1 200
0.0 200
0.0 187
0.0 208
O.O 204
0.0 204
0.0 196
-------
oo
LEN
m
.75
.75
.75
.75
.75
.75
.75
.50
.50
.50
.40
.40
.40
.30
.30
.30
RUN
NO.
R2
83
84
. .
85
86
87
88
89
90
91
92
93
94
95
96
97
STOIC.
RATIO
1.20
1.27
1.61
- - - -
1.28
1. 17
1.02
1.35
1.36
1.07
1.17
1.18
1.39
1.06
1.13
1.37
1.12
C82 02 CO NO UHC CO N8 UHC
X % PPM PPM PPM GM/KGM GM/KGM GM/KGM
0.17m I.D.. A1r-Cooled. Side-Fired
13.4 4.0 295 41 59 4.71 0.702 0.538
12.6 4.9 314 38 14 5.29 0.701 0.131
9.1 8.2 k!600 19 1630*34.47 0.462 20.068
12.0 4.9 30 21 4 0.51 0.383 0.039
13. 3 3.2 25 30 3 0.39 0.506 0.026
15.0 0.0 k!600 36 IBI k2l.69 0.537 1.402
11.7 5.9 35 31 15 0.63 0.597 0.151
11.5 6.0 55 29 16 0.99 0.564 0. 1 t>3
14.1 1.4 59B 27 24 8.44 0.409 0.190
13.1 3.3 26 35 2 0.42 0.583 0.022
13.1 3.4 33 29 4 0.53 0.486 0.035
11.1 6.3 78 25 16 1.46 0.497 0.168
14.1 1.4 M600 21 146 k22.5l 0.317 1.174
13.4 2.6 k!600 15 470 k23.86 0.254 4.005
11.5 6.2 170 17 72 3.11 0.350 0.747
14.0 2.4 363 IS 27 5.38 0.243 0.229
BACH.
SM8KE
0.3
0.3
8.5
0.0
0.0
3.9
0.0
O.I
1.6
0.0
0.0
O.I
2.1
£.2
O.I
0. S
TFG
C
172
187
128
196
243
230
243
230
'209
217
224
234
226
221
245
243
LEN
m
.75
.75
.75
.50
.50
.50
.40
.40
.40
.30
.30
.30
.40 3/4
.40 3/4
.40 3/4
.40 1/2
.40 1/2
.40 1/2
.40 1/4
.40 1/4
.40 1/4
RUN
NO.
98
99
100
101
102
103
104
105
106
107
108
109
110
111
1 IP
1 13
114
IIS
116
117
118
STOIC.
RATIO
1.35
1.02
1.62
1.07
1.52
1.26
1 .25
1.07
1.49
1.46
1.07
1.22
I.OB
1.32
1.49
1.07
1.24
1.44
1.44
1.25
1.07
. C02
I
0.22B I.
11.7
14.9
9.5
14.4
10.2
12.0
12.2
14.3
10.4
10.7
14.2
12.5
- 0.22m
14. 4
11.9
10.5
14.5
18.7
11.1
11.0
17.5
14.5
02
1
CO
PPM
D. , Air-Cooled
5.9
0.6
8.6
1.5
7.7
4.5
4.5
1 . 4
7.4
7.1
1.5
4.0
I.D.i
1.7
5.5
7.5
I.S
4.5
7.1
7.0
4.6
1.4
II
1532
25
SO
31
28
35
127
65
101
184
NO
PPM
UHC
PPM
CO
GM/KGM
NO
GM/KGM
UHC
GM/KGM
BACH. TFG
SMOKE C
, S1de-F1red
41
SI
40
56
45
54
58
54
51
46
50
60
39
I
8
5
6
4
19
83
6
0.22
20*78
O.S4
0.71
0.65
0.48 '
0.58
1.81
1.31
1 .98
2.68
0.794
0.745
0.953
0.862
0.996
0.966
0.938
0.831
1.092
0.970
0.764
45 52 6 0.73 0.912
p__ Af_n.| |tuitlBf>4jM Clita F1 nut m.
0.010
0.465
0.485
0.012
0.093
0.048
0.057
0.036
0.216
0.256
0.047
0.055
0.4 192
6.0 202
O.I 230
2.9 230
O.I 234
0.5 821
0.8 286
4.8 241
0.8 251
0.5 256
4.5 258
2.8 243
139
PS
60
456
PI
40
37
ft
545
34
4«
39
if
47
47
54
52
40
S
1
10
13
1
2
f
i
38
2.00
0.46
I. ft
6.4f
0.36
0.77
O.A3
0.38
7.69
0.579
0.874
0.853
O.«4|
0.847
0.9RO
1 . 1 3 1
0.938
0.617
0.045
0.015
0. 110
0. 108
0.014
0.074
0.07P
0.010
0.909
2. 5 IB9
0.9 198
0. 7 713
3.8 189
0.5 198
0.1 719
0. 1 774
0.9 196
4.0 . 196
-------
_,
00
CO
m»
00
UN
in
.30
.30
.30
.40
.40
.40
.50
.50
.50
.75
.75
.75
.30
.30
.30
.40
.40
.40
.50
.50
.50
RW STOIC.
NO. RATIO
0.
1 19 1 33
120
m
IP?
IP*
125
I7A
127
I2H
129
130
- » -
131
132
133
134
135
136
1 11
1 «J '
I3R
139
1.14
I.S1
1.38
1.31
1.57
1.17
1.35
1 .48
1 .19
1 .44
1.14
- - 0
1.21
1.35
1.49
1.65
1 .44
1.16
1*18
1.59
1.35
C02 0? CO
* I PPM
.22m I.O.. Insulated,
13.3 6.3 23
13.7
10.5
11.3
11.9
9.9
13.3
11.5
10.7
13.1
II. 0
13.7
p.a »I600
7.8
6.?
5.4
8.1
3-3
5-9
7.4
3.A
7.0
BS
25
15
10
20
II
12
15
15
?.R 18
n I__..l.*&*4
NO
PPM
Slde-
97
89
68
1 IP
108
97
116
101
97
110
no
136
C4.4. I
1HC CO
PPM GM/KGM
F1 r»H ...
6 0. 4P
45 S&4.03
31 1.77
IP 0.46
6 O.P*
10 O.PI
1 0.33
0.22
5 0.26
0.25
3 0.29
0.29
NO
GM/KGM
I.R33
1.436
I.4RO
P.POB
C.03P
P. 191
1.926
1.953
2.052
1-864
2.275
2.206
IMC
GM/KGM
0. OSA
O.M6
0.356
0* IP6
0.060
0. IPO
0.009
0.015
0-056
0.009
0.030
0.013
PACK. TFG
SMOKE C
0. 1 POO
9.0 170
00 POB
0. 1 IR3
1.0 IA6
0..1 P04
0.8 172
0.3 181
O.I 189
q.l 275 !
0.0 296
O.I 279
13.1
11.7
10.6
9.5
10.8
13.4
13*3
9.7
II .3
3.9
6-0
7. A
9.0
A. 9
3.2
3 *
8.3
5.H
134
65
125
57
32
60
35
so
95
88
73
66
96
107
112
81
106
2 2.15
2 1.17
27 2.51
24 1 .29
3 0.63
2 0.92
0 70
14 0-75
2 0.90
1.627
1.699
1.572
1.625
1 .984
1 .775
1 H 7*i
1 * D ' J
1-852
7.059
0.016
0.021
0.308
0.31 1
0-038
0.015
0 »0 1 3
0.175
0.026
7.0 185
1.7 192
0.7 198 [
O.I 226
O.I 215
6-0 192
?S Oil
.3 is 1 J
O.I 241
O.I 232
,EN RUN STOIC. C02
" NO. HAllfl I
0.17m I
.75 140 1.74 I.T.?
.75 loi
.75 14?
.40 I4"<
.40 144
.40 145
.30 I4A
.30 147
H "R
.50 149
.50 ISO
.50 151
.75 IS?
.75 153
.75 154
1 .72
1 .54
1 .47
1 .05
1 .74
1 .73
1 .SR
1 .OA
1 .1 1
1 .42
1 -25
1.74
l.ll
1 .54
1. 1
10.4
0.22H I
10. A
14. P
17.6
17.7
9.9
14.5
14.4
1 1 .7
13.0
13. P
1 4. A
10.4
. . rt o^ ?
RON
NO.
427
J
y, 426
o
i 429
430
" Ol
jo 432
- U.C.OBB1 1 .
STCIC- CC2
RATIO
1.04
l.£4
1 .14
1.40
1.13
1.27
X
14.7
12.4
13.5
II .0
13.5
12.1
02 CO
* PPM
.D.. Insulated
4.6 IS
9.5
R.P
DUji
. t *a
7. P.
1 .!>
4. 4
4.3
R . T
1 .4
2.4
A.R
4.7
4. A
7.1
H.7
OM»*
i IMh
02
t
0.9
4.4
2.7
6.4
2.5
4.7
20
IS
14?
1307
60
117
3AS
1532
131
RQ
4A
30
45
191
«&r~cooi
CO
PPM
1083
503
298
951
85
150
NO UHC
PPM PPM
, S1de-F1red
129
75
IOH
led SI
pi
40
34
26
8
30
45
28'
3R
4?
4A
12
ed Sid
NC
ppn
16
. 10
10
4
4
3
A
rf*_F1n
ae~r i n
1 A
49
14
16
III
49
2
7
2
2
38
»*f1r»i
e~Tire<
UHC
PPM
I 13
124
44
155
4
29
CO NO UHC BACH. TFG
GM/KGM GM/KGM GM/KGM SMOKE C
0.75
0.46
0.31
fid - - -
2.RO
18.24
0.99
1 .V3
7. 71
PI .52
1 .94
1.53
0.78
0.51
0.66
3.94
CD
GM/KUt
14. 90
U.JO
4.48
1.7.74
1.28
2.52
?.?74
1 . KH 1
P. 403
0-441
0.603
0.599
0.45A
0. 193
0.453
0.713
0.569
0.6*2
0.745
0.774
0.?*!!
NC
GK/KOM
0.249
O.I lib
0.172
0.086
0.072
0.068
0.009
0-04V
0 .01 H
0. 1 79
0-3R9
0.134
0.146
1 .343
0.396
0.013
0*073
0.019
0.019
0.006
0. 447
UHC
UI/KGM
0.888
1.167
0.382
1.652
0.034
0.283
O.I 739
0.? 768
0.0 264
0.1 1 79
5.2 170
0.7 172
2-B 177
0-3 IR3
5.0 189
2.7 172
O.I IB3
0-5 175
0-5 194
1.0 109
0.0 209
BACH. TFL
SMCKE C
0.4 2S7
0-2 2S9
0.2 249
0.1 269
0.2 274
0.2 2M
-------
APPENDIX D
DATA TABULATION: CGR BURNER EXPERIMENTS
Combustion chamber design parameters, burner operating conditions, and flue gas
and recirculated combustion gas composition data are tabulated for the CGR (com-
bustion gas recirculation) burner experiments discussed in Section V. Except
where specifically called out as being different, nominal burner firing rates
were 1.0 ml/s (gph). All experiments "were steady-state runs except Runs No.
248 and 249, which were 10 min on/20 min off cyclical tests.
Two rows of data are given for each test. Composition data in the first row per-
tain to the exhaust flue gas while those in the second row pertain to the mixture
of recirculated gas and fresh air supplied to the burner. The parameter listed
as "RR" is a calculated recirculation ratio (recirculated burned gas mass divided
by fresh reactant mass) expressed as a percentage. It is calculated by
RR - (100.) (14.49 + 1.0)/(14.49 (SRM-l.OJ)
where SRM is the stoichiometric ratio for the gas mixture supplied to the burner.
(Values of SRM are not tabulated but can be derived from the mixed gas composi-
tion data using the methods of Appendix B.)
185
-------
PI'RNFRI TGR-A f .042 H
I.O-AO-*
.22 M ri«.. TUNNEL-FIRED,
.40 M LENGTH,Air-cooled
Fl'fcNFF! TKR-A (.042 M
N0ZZLFI 1.0-60-0
-2Z M ri*., TUHHEL-FIRED,
.40 M LFNGIH. A1r-cooled
PCN JTflir.
N0. RATIO
ISS 1.09
PP = PAX
cap 0? ra NB ivr ra NO IMP F*PW.
I I PPM PPM PPM GM/KGM GM/KGM GM/KG* . 15A 1.5
PR = 31* P.8 1A.3 »1600 35 »3P99
IS8 I.P3 IP.4 4.p |R7 16 POO 3.07 O.P9A 1.868 I.O
PR = ISt 1.1 |1.9 I7P t POO
159 1.07 IP.7 1.3 klAOO 38 3POO>PP.S4 0.5X7 P4.ISP A. 5
RF = II* 0.5 19.0 CIAPO PI 1750
160 I.II IP.3 P.OJlfOO 4| k3P99 *P3.40 O.A5S»P7.W1 R.O
RR > IPX O.S 18.9 ilfiOO IS 1500
161 1.06 14.5 I.P P78 31 P70 3.89 0*4*5 P. 156 1.0
RR « PA* 1.9 17.0 t1600 PI 1550
16? I. IP 13.4 p.3 397 10 300 5.87 0. IA9 P.S3I 0.5
FP * II* 0.7 19.0 47P 9 750
163 1.07 IP.6 1.3 11600 49 »3P99 kPP.5A 0.7/6*P6.59O 0.5
P.R P3X 1.4 |7.3 A1600 40 13P99
TrG
Swirl
Angle
40°
FI-N STBIP.
N0. FPTI0
164 1.10
rap 0? re
* * PPM
No Go
NP ivr. ra NB IMP PACH. TFG Swirl
PPM PPM GM/KPM GM/KCM GM/KGM 5M0KF C Angle ,
20°
RR - 30*
P4|
l?P4.04 I . |OI»P8.338 P.O 1AO >' ,
7P »??99 1 If
rcO
35 480 10.32 0.559 ?.<>?7 f,0 P?9 **
5 370 93
38 »3P99 kpg.AO 0.6)4k33.7l3 6.0 138 5S
P8 P40O 88
PA 1500 IP4.07 0.4PO IP. 894 p. 5 168 55
PI 1500 88
36 1500 »P?.P9 0.55P 11.939 I.S P38 ^
PI 1PSO AA
7P ».1P99 kP4.47 |.19lkP8.8/4 p. 0 I8P 10
4O 3000 ' 74
10°
-------
PllfrNFPl CGR-B ( .042 M
N0ZZLF: I. 0-60-A
CeMBUS10F-« .22 M niA.. TUWa-FIRED.
.40 M UFNGTH, Air-cooled,
Baffle at .23n
Pl'RNEF: TGF-A ( .042 M 01 A)
NBZZLF: 75 -60-A (184-186)
1.0-60-A (187-190)
reMRl'JTOFi .22 M M A.. SIDE-FIRED.
.40 M LFNG1H. Alr-COOled
00
PI >N stair.
N0. RATle
I7P
RP
173
RR
17*
PP
175
RR
176
RH
177
RR
178
RR
179
ff
180
RP
181
RR
182
RR
183
RR
LIP
« P97
1-08
» 31*
1.16
- 31*
1. IP
= P7*
1.1*
» II*
1. 17
10*
1.07
« II*
1.08
= IS*
1.06
P6J
1. 10
?«1
1.0*
96*
l.ll
« P6»
r0p
*
19.
P.
13.
P.
IP.
9.
13.
9.
13.
1.
13.
0.
1*.
1.
13.
1.
13.
1.
IP.
9.
I*.
p.
IP.
P.
9
9
*
8
P
8
1
6
6
1
3
9
3
1
6
P
3
9
7
P
1
5
P
6
OP
I
P
16
1
16
3
16
9
16
P
19
3
19
1
19
1
18
1
17
1
17
1
17
P
17
. *
.6
.6
. *
0
*
.*
.8
.8
1
.P
.3
.5
.0
.7
.S
.3
.n
.9
.0
0
.0
.0
.0
r0
PPM
M600
41600
» 1600
1600
11600
ill 600
»1600
iu.oo
no
»IAOO
383
397
11600
730
k IAOO
»1600
*I600
k1600
k!600
k1600
k!600
k|60C
*1600
NO
PPM
*0
10
35
?6
39
18
.16
10
*3
n
*l
5
*0
7
36
9
3?
PO
31
10
*0
8
31
5
mr r0 HI two FAPH. TFB Srlrl!
PPM GM/KGM GM/KGM GM/KGM 5M9KF T Angle
POOO »P3.75 0.6*9 16.961 1.0 1.18 10°
7 SO P18
1300 »PP.79 0.5*9 10.580 3.0 17* 20°
PbOO P16
IP50 i?*.6l O.SP9 10.989 3.0 17* 30°
PISO PIS
750 »93-7l 0.587 6.351 3.0 17* 30°
900 P1R
980 1.66 0.699 1.898 5.0 PO* 40°
POO 0.77 0.691 1.76* 1.0 135 20°
180 IP9 |
900 5.63 0.6PO 1.619 3.0 199 10° '
630 1*3
PIO I0.*7 0.568 I.7PO 5.5 PI6 10°
RSO 15*
1800 »P?.*9 0.*83 |*.*S7 P.O 199 20°
PISO 191
600 »P3.Pl' 0.*>J3 *.97* *.0 IHS 20°
IP50 193
390k?P.I1 0.598 3*080 P.O 1RP 30°
5*0 P07
P700 »?3.*P 0.*87 PP.581 *. 0 40°
RUN ST0IC.
N0. RATI6
18*
RR
18 S
RR
186
RR
187
RR
188
RP
189
RR
190
RR
I.I*
= 1**
1. 16
= 38*
l.*l
?3»
1.15
» 131
I. 01
= ?3*
I.OP
= P4S
1. 19
= I3»
C0?
*
IP.
P.
IP-
S'
R.
1-
13*
1.
IS.
P.
IS.
3.
13.
1.
6
*
1
6
6
6
7
9
1
6
0
1
1
6
9P
*
P.
18.
3.
IS.
S.
17.
3.
18.
0.
17.
0.
17.
3.
18.
A
A
0
6
3
*
0
7
*
3
6
P
6
8
re
PFM
11600
11600
» 1600
>I600
»1600
k|600
55
10
k|600
315
%1600
503
35
137
N0
PPM
31
S
PI
5
68
91
PI
5
P6
6
PR
8
P7
S
cue r0 Ne INT
PFM GM/KGM GM/KGM GM/KGM
1650 kp*. OS 0.501 M. 17?
*30
1550 kp*.6* 0.3*7 13.6*0
*50
»3P99 430.10 1. 378k35. *79
I7SO
ISO 0.8* 0.351 1.305
l
100
850 kPI.51 0.376 6.SP9
160
900 »PI.69 O..M5 A. 973
3*0
130 O.SS 0.*57 1.170
ISO
PAfH. TFG SKlrl
»BK K r Angle
3.C I3P 40°
1 10
0. 5 1 P 1
PIO
10*
ISP
O.O PPI
POP
0.0 PPI
PP*
0. 1 P07
PP*
0. 1 PP7 4
0°
179
7 10
PI6
-------
PURNFRl CGR-A ( .051 M 01 A>
NOZZLFI I. 0-60- A
C8NPl'ST9Ri .22 M DIA.. TUNNEL-FIRED,
.40 n LFNGTH. Air-cooled,
Baffle at .23 m
Pl'FNFRl CGK-A < .051 M nl A>
NOZZLE! 1.0-60-A
C0MPUST0R! .22 M PI P., TUNNEL-FIRED,
.40 M LENGTH. A1r-coo1ed,
Baffle at .23 m
RtN STOIC.
COP
N0. RATI0
f
1
1
I
S
e
s
3
191
RR
19 P
RR
193
RR
194
RP
195
RR
196
RR
197
RR
198
RR
199
RP
1
e
1
e
1
a
p
e
1
8
1
1
»
1
3
1
.55
10
z
.0
/4I 4.6
. 4p
40X
.94
34*
.09
IS*
.51
PIT
.51
45*
.54
39*
. 16
P7*
.P7
171
10
3
7
P
7
1
10
P
9
3
B
e
IP
p
9
1
.p
.8
.5
.8
. 1
.7
.P
. 4
. 1
.6
.0
.6
.6
. 1
.?
. 1
0S
R
IS
A
IS
in
16
II
IB
7
17
7
14
6
16
3
16
3
IR
T
. 1
.0
.4
.4
.5
.0
.P
.4
.5
.A
.0
.9
.5
.P
. 1
.«
.8
.P
C9
PPM
1099
IAOO
.1*00
* IAOO
.1600
.1600
PO
629
PO
.1600
.IAOO
1600
.1600
klAOO
.1600
a IAOO
.1600
.1600
N0 IMC C0 MB I'HC RACK.
PPM PPM GM/KGM GM/KGM GM/KGM SM0KF
P3 P80 PP.BP O.SP4 ?.??! 0.5
II SO
33 1400 »3P.P7 r.*<10 1^.134 P.I
31 1900
PI 110*4P.06 C.59P 1.65P r.O
38 P350
3P 73 0.57 0.975 1. IH 3 0.0
8 IPO
33 64 0.42 0.7P3 0.736 0.0
9 P60
18 570 -3P.37 0.400 6.589 0.0
31 P.3SO
P6 »3P99 13P.99 0.576*38.883 0.0
50 >3P99
31 380 1P4.63 O.SPI 3.343 0.0
P6 POSO .
35 »3P99 »P7.03 0.6SP.3I.854 3.0
A3 .3P99
1FG SW
r An
4
PP9
K '
P16
PIO
154
168
138
IBB
IP7
"93
PIR
1AO
177
141 t
I3P
Irl RIW .'
gle N0. f
0° POO
Rh
pot
RR
POP
RR
' P03
RP
P04
RR
90S
' RP
I
P06
| PR
; e07
RP
0° P08
RR
T0IC.
-------
Pl'RNEPl CGP-A ( .051 M niA>
N9ZZLE: 1.0-60-A
C0MPl'?T0Fl .22 M DI A., TUNNEL-FIRED.
.40 H LENGTH. Air-cooled
Baffle at 23m
BURNER: roR-A c .051 M DIA>
3/4" Choke Ext.
Nozzle: 1.0-60-A
C0MPHST0M .22 M HA.. TUNNEL-FIRED.
M LFNGTH, Air-cooled,
Baffle at .23M
00
vo
RUN <
NB. F
209
RR
210
RR
211
RR
pie
PR
STB 1C.
(ATI0
1.49
3 34X
1.32
- 26X
I.7O
2SX
1.46
21Z
Cl
10
3
1 1
2
9
2
10
1
92
X
.4
.4
.7
.6
.0
.4
.4
.8
1
7
16
5
17
9
17
6
17
02
X
. 4
.0
.5
.0
.1
. 1
.9
.6
ce
PPM
61
1051
80S
10
IS
10
17
*1600
N0
PPM
21
10
22
8
21
S
21
6
IHC CO N0 IWT
PFM GM/KGM GM/KGM KM/KGM SM0KF
ISO
160
1.23 0.448 1.70S
120 14.18 0.414 1.203
100
110 0.34 '0.515 1.437
100
95 0.35 0.448 I.OSS
440
ACH. TFG Slflrl RMN ST0IC.
M0KF r Angle N0. RAT 10
o.o |46 40° 213
199
0.0 179
196
0.1 !*»
193
RR
P14
PR
215
RR
0.0 157 40° 216
166
*
KR
PIT
RR
218
RR
219
RR
PPO
RR
PPI
RR
PP2
RR
223
RR
1
a
i
3
1
3
1
3
1
=
1
3
1
3
1
a
i
3
1
=
1
B
.57
31X
.52
317
.29
33X
.43
22X
. 19
I3X
.37
19 X
.51
17X
.76
161
.32
131
. 17
P.4X
.78
I2X
CO 2
X
10.
3.
8.
P.
10.
2.
8.
1.
11.
0.
10.
1.
10.
1.
8.
1.
II.
0.
11.
1.
8.
0
1
5
8
/
2
2
2
2
0
9
8
2
1
4
3
8
7
9
3
4
3
9
0?
I
8.
16.
3
3
6.8
16.
4.
16.
5.
17.
3.
18.
6.
17.
7.
IB.
9.
IR.
5.
IB.
P.
17.
9.
18.
3
4
I
9
5
1
7
0
9
A
1
1
3
5
7
9
2
A
9
ce
PFM
25
153
.,1600
1600
.1600
1600
>1600
1600
1600
31600
k|600
31600
20
10
20
10
40
1600
31600
1600
50
10
N0 IMC C0 N0 IHf IrACH. 1FG StflH
PFM PPM GM/KGM GM/KGM GH/KG^1 JM0KF T Angle
IS 23 0.53 0.357 0-P76 0.0 154 40°
S 30 185
18 1850 332.50 0.402 Pl.TI O.I M?
41 >3P99 IP?
22 3POO *P7.32 0.^0? :>1.22S 0.0 135
40 .3299 IP9
38 >3299 331.61 0.824X37.260 ?. 0 116
45 »32*9 91
31 2350*25.15 0.5P3 PI.105 6.0 160
10 1000 71
27 350 329.24 0.531 3.655 3.0 171
I? 1250 10?
26 130 0. 42 0.5*3 1.497 0.0 199
5 110
IS IPO 0.47 0. 4QP 1.621 0.0 171
5 110 1*3
26 IPO 0.70 0.49? 1.207 P.? IK5
5 350 I 1O
26 2600 3p4.76 0.43? P2.988 0.5 IfO
IB 1850 113
PS 110 1.20 0.644 1.507 0.0 210 40°
5 105 * 118
-------
PMPNEPI rGR-C* ( .051 M DIA)
.15m D x .25m L Shroud
Nozzle: 1.0-60-A
VO
o
FUN !
MB. i
?24
RF
EPS
RR
PP6
RR
827
PR
228
RR
P29
PR
230
RR
!T9IC.
; ft TIB
1.51
= I*T
1.53
= 252
1.29
15X
1.45
° 45Z
1.21
= 25*
1.07
« 311
1.15
" 321
C02
X
9.9
1.4
9.5
1.8
8.P
0.9
10.6
3.8
IP. 9
P. 8
14.5
3.2.
13.6
3.4
02
I
7.4
18.6
7.5
17. 1
3.6
18.4
6.9
14.9
3.9
17.1
1.4
16.4
3.0
16. 2
rt)
PPM
693
PI
k|600
»I600
klAOO
»1600
355
» 1600
90
41
IS2
»I600
60
41
N0
PPM
15
12
30
9
26
1 1
es
II
21
10
26
9
ei
9
.22 M 011>.. TUNNEL-FIRED
.40 M LFNGTM. Air-cooled
l«C T9 N9 IHC BACH. TFG StflH
PPM GM/KGM GM/KGM GM/KGM fMBKF C Angle
130 13.99 0. .T43 1.499 1.0 213 0
68 99
80 »3f>.73 0.659 0.935 O.I 1*8
P80 1?9
1800 kl>7.40 0.479 17.614 f-.O lf>9
720 88
60 6.86 0.518 0.«6S> 0.0 ISP
POO 154
/8 1.44 O.^O 0.439 O.I POP
I 00 I 46
51 2,58 0.394 0.4)1 O.I 202
135 174
56 0.91 0.343 0./88 0*0 196 0°
95 191
PUPNFR! CGR-C*t .
.15m D x .
Nozzle: 1.0-60-A
RIM
NO.
231
PR
232
UP
23?
RR
234
RR
235
RR
S36
RR
237
RR
238
PR
STBIC.
RATIB
).?B
= P4*
1.36
= P4T
1. 13
° ?6*
1.24
» 31*
l.4|
» 33*
1. 18
= 32*
1.52
14*
1.18
= 151
CB2
12
P
1 1
P
13
7
I
.0
.6
.5
.6
8
.6
IP. 6
3
9
2
13
3
10
1
13
1
.0
. 7
.5
. 1
.1
. 1
.4
. 1
.2
051 M DIA>
25m L Shroud
a?
4
17
6
17
P
17
4
16
6
16
3
16
7
18
3
18
T
.9
.P
.0
.>
.6
.0
.3
. 4
.0
.1
.5
.2
.6
.6
.5
.5
ce
PPM
20
10
31
503
30
355
17
30
k|600
k|600
PO
57
IS
10
20
139
NB
PPM
30
9
9
31
9
31
10
21
10
31
II
35
9
50'
9
.22 M DIA., TUNNEL-FIRED.
.40 M LFNC.TH. A1r-coo1ed
we re we UHC PACH. TFG Strlrl |
PPM GM/KGH GM/KGM RM/KGM 5MBK F C Angle
53 O.H4 0.548 0.515 0.0-191 40° j
49 185
53 0.58 O.AOP 0.5« C.O 177
70 168
5? n. 45 n. 506 0.444 0.0 18?
69 160
55 O.P9 0.546 0.515 0.0 18S
S3 210
II 00*30.16 0.4P4 11.8*9 ?.Q 157
8PO 168
53 0.31 O. 522 0.475 0. n 17'
55 199
46 0.31 0.785 0.535 0.0 189
43 163
4B 0*31 0.345 0.430 0.0 IS? 40
61 14|
-------
PURNFfi: CGR-A ( .051 M DIA)
N0ZZLE: 1.0-60- A
(0MBUST0r« .22 .4 DI A.. SIDE-FIRED.
.40 M LFNSTH. Water-cooled
ElTNFIx: r.GR-A < .051 M MA)
N0Z2LF: 1.C-60-A
C0MPL'ST0Rt .22 M MA.. SIDE-FIRED,
.40 M LFNGTH, Water-cooled
FIN 5TBIC.
N0. RATI0
191/192
239
RR
240
RR
241
RR
242
RR
P43
RR
244
RR
P4S
RR
246
RR
P47
RR
I.P9
a SIX
1. 10
a 311
1.17
« III
1.89
1.39
» 111
1.16
= P2X
1.49
= PIT
1.4)
= P2T
1.39
a S91
C02
X
1 1
2
.9
.8
14.5
3
II
0
7
0
10
0
14
0
9
0
9
1
.0
.9
.5
.3
.4
.8
.9
5
.7
.7
.5
.0
.4
10.4
1
.1
0?
X
5.
1A.
2.
16.
3.
19.
9.
6.
19.
3.
17.
6.
17.
5.
17.
6.
16.
0
3
2
3
0
0
7
9
1
O
4
5
9
A
4
5
0
6
CO
PPM
838
412
»I600
404
177
30
*1600
P20
130
10
80
77
665
PI4
157
57
44|
231
N0
PPM
11
5
PI
5
40
2
14
2
21
2
PO
2
8
2
6
?
9
2
IHC re NO t«r PACH. TFG Swirl
PPM GM/KGM GN/KGM GM/KGM «M8KF C
RUN stair.
N8. RATI0
ca?
X
no NB
PFH PPM
IMP ra N0 twr PACK. TFG Swirl
PPM (M/KGH GM/KGM GH/KGM SM8KF C Angle
?7 I*. 3? O.PPS O.P63 O.I 17.4 40°
210 163
IISO »23.27 0.327 9.SS7 0.1 177
POO I 49
47 P. 76 '0.680 0.4U 7.0 POP
SO 77
800*40.93 0.407 11.693 O.I PP1
140 74
80 P. 4| 0.417 0.846 0.0 ?4|
46 SP
34 I.P3 0. .138 0-P99 9,0 20?
39 107
130 13. PI O.I7C 1.474 0.0 P04
100 IP4
48 P.98 0.130 0.517 1.0 163
36 113
110 8.PO 0.191 1.167 0.5 19* 40°
90 ISP
?4B 1.38 11.0 6.1
RR = 191 ?? 17.9
P49 I.P6 IP.I 4.6
RR a IP.Z I.I 18.9
753
369
65
20
10
4
P7
P
P16
9
13.83 O.PIO P. 391
1.O9 0.486 P. 086
O.I 174
11'
P.2 157
79
40°
40°
-------
APPENDIX E
DATA TABULATION: FGR BURNER EXPERIMENTS
Combustion chamber design parameters, burner operating conditions and flue gas
composition data are tabulated for the FGR (flue gas recirculation) burner exper-
iments discussed in Section V. Steady-state data are given in Tables E-l through
E-3. Cycle-averaged data are given in Tables E-4 and E-5. Cyclical tests were
all 10 min on/20 min off experiments. -
Two rows of data are given for each test. Composition data in the first row per-
tain to the exhaust flue gas while those in the second row pertain to the mixture
of recirculated gas and fresh air supplied to the burner. The parameter listed
as "RR" is a calculated recirculation ratio (recirculated burned gas mass divided
by fresh reactant mass) expressed as a percentage. It is calculated by
RR = (100.) (14.49 + 1.0)7(14.49 (SRM - 1.0))
where SRM is the stoichiometric ratio for the gas mixture supplied to the burner.
(Values of SRM are not tabulated but can be derived from the mixed gas composi-
tion data using the methods of Appendix B.
193
-------
TABLE E-l. STEADY-STATE FLUE GAS RECIRCULATION DATA FROM A 0.22 M I.D.,
WATER-COOLED SIDE-FIRED COMBUSTOR WITH THE 1.0-60°-A NOZZLE FGR BURNER
0.051m Choke L = 0.40m
0.042m Choke L = 0.75m
10
Rt
P.7 klAOO ?< .PCOO k?'.f»l 0.531 17. l« 7.0 P31
15.9 <95 10 AID S'
4.7 k1600 30 PIOOk30.7l O.AIf 73.03' P.O 735
IT.S 16* 10 61S SP
8.0 »l«00 PO 1150*33.77 0.1" 13.87? O.S 73?
19.0 307 5 300 35
1.9 7A* 15 'PO 11.11 0.7*6 .1.117 O-O PP1
IA.5 1ST S PPO '6
P.' 703 10 P30 '.50 0.170 1.9*7 P.O PP^
IT. 6 70 » 130 y
'.7 P78 II I'D '.77 0. PO' 1.75* 1.0 707
18.1 'S ' *0 P9
«rl PIN ST8IP.
NO. RP110
0° PA* ,.!'
PR » PHI
P67
KR
PA1
RR
769
*K
PTO
RR
*TI
RR
P7P
RR
3. A Sll 11 P*0 R.I3 0.191 P.7<3 O.S PPI 40° "3
17.5 70 7 110 'A ft
27'
PR
P7S
RR
P76
RR
1.1*
1
1.1?
791
1.37
III
l.?5
- ?U
i.np
ri-i
I.O*
* ?u
i.ia
" P7I
i.-p
in
1.76
. lit
1.PS
301
PO?
I
13.'
3.3
1'.
0.
13.
3.
11.
1.
17.
7.
14.
1.
1'.
1.
IP.
P.
10.
0.
IP.
0.
17.
p.
3
0
'
1
1
1
0
P
P
»
3
*
*
6
6
7
0
7
0
6
e?
i
7.7
16.7
'.0
PO.*
3.0
16. A
6.0
19.0
'.'
17. f.
O.S
17.0
1.3
17. A
3.5
16.8
6.5
19.0
'.5
19.0
'.S
16. S
CP
f-ftl
197
10
3O
f
709
30
80
1*00
1 179
PPO
'0
1099
P3I
503
50
70
10
1'03
383
NP
13
7
O
15
10
10
5
10
5
M
5
H
PO
10.
1 '
5
15
S
1*
10
IMP
105
37
0
710
70
'5
1 10
50
1900
no
110
55
AOO
160
110
55
SO
'3
100
160
re N3 mr f»PH. iff, Sxlrl
i/KrM ri/Kf.M rn/,
'i
3.09 0.17* 0.813 O.O f07
'I
17. PI 0.3" 5. 7*6 O.I ???
*»
».S3 0.7PI I.IK* 0.0 17*
P*
1.17 O.P«' O.J77 0.0 16*
P7
?3.P* 0.3S5 7.SB.1 O.I ?!>T
«0*
-------
TABLE E-l. (Concluded)
10
Ul
0.051m Choke L = 0.75m
RIW stair, rep tf re \e I-MC ra NO IMC p»n4. rrr. Swtrl
N8. |i*TIO 1 I PPM PPM PW (M/KGH CM/i
.56 10 110 17.94 0.775 1.739 n. 1 7*7
70 4 40 79
P.?*
P. i
260 1.36 11.1 6.011600 70 «CO »f
KP Set 3.5 16.7 544 ft 700 54
761 1.41 10.« 6.6 I??T 15 3*0 76-17 O.370 3.765 O.I 7'A
RR 1*1 8.* 17.9 754 6 110 4«
?6P 1.18 13.1 3.5 315 13 110 4.95 0.773 0.9)14 0.0 774
RR 301 3.6 I*.5 90 5 75 57
7A3 1.74 17.4 4.4 177 in 73 7-93' O.IRS O.ADt P. C 733
RR PI I 7.« 17.6 40 5 «3 43
764 1.79 11.9 5.0 75 73 ?4 |.7* 0.437 P.579 0.1 713
RR 171 I.A Iff.9 70 5 4A 7*
765 1.13 1.1.K 7.6 70 71 57 1.05 0.336 0.444 0.1 7P4
tr 711 7.4 17.6 70 * 44 37
Irl
0°
J°
FI-N MBIT.
777 |.?A
FF 1 l>
77". 1.74
FF 717
779 |.?ft
FF 777
FF 111
Wl 1.49
FF 77T
7«7 I.'P
FF 79T
7«3 1.10
FF. « 73T
7.4 ,.-.
««. ..«-
FF jo»
re 7
11.5
1.4
ir.4
?. *
17.0
P. A
1.7
IP.I
1.9
ir.i
7.K
13.7
7.7
II. 1
II.*
7.X
ep re KJ i»r r.
,.t ,.- « 77 A..?
19.0 41 7 1*
4.< »I»OC ?» 1?PC »?A.?7
IA.4 711 9 tr
4.9 ,n.- ,p ?ir n.??
17.5 177 7 1TP
19. P ?7» « I/P
7.3 «i*rp i* *rp »*»i.79
1
17. « /7? « f?f
«.» »iApc ?i i»«r »r9..f
IA.» 711 r^p
7.P IJCC1 l« *1P tf». l«
17.* ir«9 « ?rr
».7 P*4 IP I'P /.f<
!.. T7 7 | CO
..9 .UCO l« IIT ,r,.,
,,.. M, . «,r
NS t^r rfry. nr 5»1
P. lf» P.7/? P.P PT7 0°
79
P. 4f | I/. 174 P. 1 ff9
A"
o. it1* r».^7« P. o f*i*
*7
41
r.''^" .on-> n. i "T
«!=
n. />»p >*./»/ p. i f ^ /
'3
P.7*. J..0..9 C^ 1«
49
P.?I1 I.7A? 0. 1 »!«
P.^AI M.f'» P.I f*9 0
AT
-------
TABLE E-2. STEADY STATE FLUE GAS RECIRCULATION DATA FROM A 0.22 M I.D.,
INSULATED COMBUSTOR WITH THE 1.0-60°-A NOZZLE FGR BURNER
Side-Fired
0.051m Choke L = 0.40m
Tunnel-Fired
0.051m Choke L = 0.75m
UD
FI« :mc.
t». FATIC
M* 1.71
RF 1I»
m l.«3
RR 701
Tfld i.«7
RP PIT.
709 1.55
FF 301
790 1.38
FF « 301
791 1.31
FR « fit
79? 1.44
FR |4»
. -O 1.8?
FF » 75T
re? es> re MB iwr re v* mr FOPM. im 5*1rl M»« iitlc.
7 T PHI Pltl Pffl CM/KG* rx/<6H R*/«B« p«r r Ni . »>MI8
8.8 9.0 ID 47 1 0.44 1-190 O.C13 0.0 «9 '0° 30°
1.4 19.0 10 1 t ?9
9.:» H.5 17 90 0.?« 0^704 0.01* O.T ?9.i
3-4 It.; 10 9 p *»
9.4 9.2 17 3| | 0.40 O.74I 0.019 O-O 791
P.4 17. « 10 8 V M»
9.7 7.7 40 P3 « 0.83 O.JP3 0.0?« *.0 M»
3.4 16.5 ' 10 10 f />
3.4 1H. 5 P*4 A 80 "
11.5 *. 1 130 41 t- f.t* O.777 O.P** K.« ?M
P. 4 17.5 «O 5 14 3?
10. 1 *.7 ?0 5? 9 0.4A 1.^91 0. lOt ^.0* ?7 1
1.8 1B.H 10 * 17 ?9
RR
301
FR
see
R(f
303
I.R
304
RR
305
RR
304
RM
H.7 I0.0kl>00 45 ?«OO »?9.PP I.I9R ?5.CI* C. 1 T79 Vf 3("
'5 l?.5 tIAOn ?6 7000 **
308
RR
30*
Fft
310
RR
311
RR
312
FR
i*e4
3C1
1.30
SSI
1*47
111
1*01
4*1
1.13
301
1*15
eet
i.eo
121
I.I*
311
1.44
301
1.53
eei
1.47
- in
1.38
421
1.21
411
c«e
1
ie.4
3.4
te.o
e.7
II. 0
l.e
14.4
4.1
13.8
3.5
13.5
B.4
IE. 9
1.8
ia. i
3.8
10.8
3.8
10. B
2.*
«.S
1.4
11.8
4.8
1C. 7
5.9
*e
i
4.4
14.5
5.e
17.5
7.4
l*.o
0*2
ie.«
B.5
14.5
e.v
17.5
3.8
ie.»
3.4
14.4
4.*
16.5
7.8
17.5
9-e
19.0
4.5
15.2
3.*
13.4
C9
eo
10
eo
10
eo
1C
»l*00
i|4OO
65
15
CO
10
eo
to
eo
10
20
10
eo
10
17
10
eo
10
eo
10
H»
H-W
ee
et
5
38
5
El
5
ei
5
28
e
35
e
15
2
15
4
ei
4
27
e
15
3
10
5
UMC C* N8 UHC
PHI Ctl/KGM tM/KCH Gn'KbN
3 0.33 0.393 O.CB9
4
2 0.34 0.462 O.C20
3
0 0.39 O.809 0.000
1
S 121.33 0.300 0.041
4
2 0>37 C.335 0.017
1
1 0.30 0.4*5 0.009
1
1 0.32 0.614 0.00*
1
1 0.31 C.246 O.COf
0 O.38 0.327 0.000
1
0 C.4I 0.459 0.000
1
0 0.40 0.450 0.000
0
0 0*37 0.313 C.OCO
1
1 0.32 O.IC3 O.OO9
e
BACH, ire Sxtr
SKIKt C
0.0 387 «0'
0.0 338
0.0 341
35
0.0 302
85
0.0 314
0.0 313
52
0.0 313
35
0-0 313
44
0*0 335
82
O.O 341
48
0*0 34*
38
0.0 332
**
o-o 3e« <
10S>
-------
TABLE E-2. (Concluded)
Side Fired
0.072m Choke L = 0.40m
Tunnel-Fired
0.051m Choke L = 0,40m
H* fTOir.
MO. F.MIO
794
PR
»5
RF
r9i
lilt
797
M
798
RR
799
PR
1. I?
SIX
1.31
7I«
I.7D
III
I.C4
fll
1. *?
?9T
l.«4
III
CO 7
J
13.1
4
1 1
0
.9
7.8
17
1
9
7
9
T
1
.?
A
. 1
.«
. 1
.7
.7
.7
J
?« "
16. i
5.4
17. f
5.0
19.0
9.0
17. A
fl.6
IA.A
1C. 1
19.0
re
PP«I
1B7
55
170
30
IIP
17
1*19
37*
> 1*PC
749
31«
50
HZ
ff
y\
10
a
10
6?
A
14
9
7?
10
10
5
IMC re N« IMC **fH. TFT S«f1rl
pp^ r^/»cf>i r-M/KG*4 P^/KG^ fiaKF r
7« 7.9Z 0.530 C.PM 1-P 7«« 40°
7?C "-7
74 7.09 O.H95 0.??9 9.0 77/
7? «
2? I.K7 I.IA* 0.714 «.0 t'P
11 ' ?'
T^P ?I.7P O.-»«I ?.<7J 0.1 ?O7
100 **
«PO »?4.7« 0.575 9.9?0 C.I JP7
750 IS
AH T.«A 0.7*4 O.TA7 0.0 T-l-l 40°
0 3»
KUN STflC.
M* RATIf
313
RP
314
MR
315
MR
314
RP
317
PP
311
PP
31*
PR
380
PP
381
PR
1.40
811
1.84
SIX
1.32
82X
1.40
111
1.18
381
i.ce
4ft
I.S8
I4X
1.18
45X
1.8*
811
caz
X
10.4
8.1
1C. 7
4.0
II.*
8.8
II. 1
1*4
IS>*
3.9
15.1
4.0
11-9
!*
14.1
4.*
18.8
8.4
8
X
7.4
17.4
4.4
14.3
4.4
17.4
4.5
19-0
8.4
14.8
0.4
14.4
4.4
11.5
8.4
14.*
4.*
17.4
Cl
PPM
3O9
90
I4«
41
110
30
40
10
4O
80
13(7
117*
40
II
too
40
10
80
N<
PHI
*
4
10
4
4
8
87
8
13
8
l
II
8
81
8
*
8
II
8
UHC
PPM
8V
81
7
13
8
1
1
0
18
14
3
8
18
3
4
4
m UHC B»CH.
GM/rtGM CM/KM
Cl
P1/«C
8V 4.81 0*804 0.338 O.O 8*9
40
81 8.43 0.187 C.I97 0.0 8*C
44
1.93 O.IOI 0.130 0.0 88*
»4
0.93 0.448 0*011 O.O 8*4
37
0.74 O-8II O.OC* 0.0 8*8
18 1C.74 0-149 O.O93 0.0 8*4
1.04 0«3*S O.O3O O.O 8*8
3*
1.41 0*148 O.IOI. 0.0 8*1
74
1.34 0.818 0.04* 0.0 874
' M
40°
-------
TABLE E-3. STEADY-STATE FLUE GAS RECIRCULATION DATA FROM A 0.22 M I.D.,
INSULATED, TUNNEL-FIRED COMBUSTOR WITH THE 1.0-60°-A NOZZLE FGR BURNER
0.051m Choke L = 0.50m
RUN STtlC.
M». RA1H
3J2
fc*
323
RR
384
RR
32ft
RR
J24
RR
327
RR
32*
RR
32*
RR
330
RR
331
RR
332
RR
1.12
III
1.03
321
t.ei
221
1.14
sex
I. 09
301
1-10
221
|.0<
ESX
1.20
P21
I.O7
831
1.22
I3X
1.13
161
caz
x
13.*
i.e
14.1
3.9
I2.(
e.t
13.4
?.«
14.1
3.4
14. O
2. 1
14.2
3.0
12.7
8.4
14.3
8.4
12.4
1. ft
13.1
02
X
2.1
l*.0
0.7
16.2
3.*
17.1
3-8
17. S
1-9
16. S
8.1
17. S
i.e
14.7
3.7
17. ft
1. ft
17.4
4.O
11.7
8.4
C»
rm
113
40
k!400
tot
177
40
110
80
148
»O
101
80
90
30
40
ft
8*0
*0
4O
ft
45
PPM
86
S
10
8
4
8
I
8
10
8
13
8
88
5
83
8
81
8
31
4
35
UHC
2
0
14
S
84
4
*
4
1
3
O
0
O
0
O
o
0
0
0
0
335
RR
If.
535.
RR
556
FR
??7
ff
53"
RF
559
RK
5.0
Rfi
! 30°
I «7
I ff
0 0.50 0.<7* 0.000 O.C 50?
I 'I
0 0.36 0*671 O.OPO O.C Wt
0 0.57 0.771 P. OPO P.P 50'
0 5«
0 0.77 P. *0y 0. POP P.P 7* ^
0 **
p p. 4* 0*57^ p.rpo o.r T?
o «T
0 0.5? 0. AM9 O.OPO 0.0 791
0 ?>
0 0.35 O.»1* O.OPP 0.0 P»»
P 7*
P 0.«4 0.4AB P.OOP 0.0 ?9»
0 *P
o 0.49 o.«*» O.POP c.r T7
0 «'
198
-------
TABLE E-3. (Continued)
0.042m Choke L = 0.50
Fiw Moir.. ro? r? ro NH i«r r« N? IMP Fern. I«T. Swirl
NO. F»ilfl T J f-Fn im >FI r.i/KC.1 ri/^r.i ri/ ?.« 17.5 10 * T
3** l.*0 ll.l A.* 70 3? P 0. .17 0.719 0. OPP P. P '>" 10"
FF l?» 1-7 IR.R 10 7 0 77
3*7 1.07 \J.f I./ IR7 J3- O P.^^ n.^6 O.PPO I.P 7« I 30°
FF r>9J ?.* 16.* 6C * 0 *6
3*R 1.07 I *. S l.£ Al 71 P O.R7 P.MR P. POP 0.0 ^PP
FF 3U 3." 1».3 ?0 f 0 ?*
3*9 1.1? 1?.R 7.6 30 71 0 0.** O.?3f 0. PPP 0.0 T7
FR 7B1 3.S l«-7 1071 5'
350 1.71 l?.0 3.* 71 73 P 0.35 P. *P? P. Of-P 0.0 ?Q9
FF 771 7.R 17.; 10 7 I
351 l.H I?.! **.?> 77 7R P 0. ** 0« *R 0 0*001 0.? TCR
RF 77T P./. 17.f 10 3 I '1
3i7 I.I? 1^.* }* IfO 31 I 1.78 0.*-)* O.005 0-5 FH
FF ?3T ?.7 17.* ? r 1 *3
3S3 l.*9 10.* 7.* 70 ?! I P.*0 0.7<« 0. PP9 O. » 9ft 30°
.FF III 0.9 19.1 IP P I 3*
I*.I I.R 130 79 « 1.11 0."-1 P.P7f O.n 7»R 20°
3SS ?.?» 1.3 7.* *7 71 O !>» P.73? P.PPP 0.0 ««<
FF PM 3.1 17.* I* - P 'f
3V. l.r? 17.7 *.O *0 71 0 0. f* 0. >*» n. POO I.P 777
pp . |v? 7.1 ^.5 7 7 O 3?
3S7 1./5 IP. t- 7.P MA9 7A 0 If.flP 0.^*P 0.000 *. f 7P7 20°
FF 171 1.5 IR.V 40 f C 79
1.0-90°-A 011 Nozzle
0.051m Choke L = 0.50m
Fiw STOIC, re? 07 ra w IMP ro NO ivr *ern. ire Sirlrl
w0. F»7io » 5 em FF» FFI G«/K6« PH/I i *
359 1.77 |7.P .1.7 35 PI 0 O.S9 P, 37R P.PPP P. T P9»
FP 371 3.9 16.? 70 * 0
360 I.?? 17.7 *.0 77 7? P O.*i 0.3«O P.PPP P. O
FF 3IT 1.7 I*.* l"> » I *<
3AI 1.39 ll.l A.3 *0 If 0 0.7* 0.33! 0. TOP O.P 3f 40°
FP 311 .i.P IA.J 7P * 0 ««
199
-------
TABLE E-3. (Continued)
zo8
0.064m Choke L = 0.50m
MIM JTBIC. TOP OP re u» IMP n NO IMP torn. irr. Swirl
i*p I.P« IP.O *.n *i* 10 IT T.I? P.19« I.P«» P.O IPP 20°
FP » 31» 3.'' 1*.* 13O * ?S
3A? 1-1? 11.* P.* "0 15 II 1.35 P.P«S P.PH I.C
FF 1O7 .1.7 IA.5 Pf * I
1 0 0-1« 0. «PP P. POP P.* W
FF 3P» 3.4 I*.* 10 A P "
If4 1.IP I?.* P."1 IP '* P P. *4 P.7P^ O. POP *.P '<'
K> ?>! ?.' 1'.' IP 4 P 'P
-"» 1.19 17.7 1.5 PO ?0 0 C.?l 0.«"> O.PPO ' 0 P77
PP . ui i.j is.* 10 5 n ?»
170 1.76 l.t 9.5 7P P* 1 l.»* P.**? O.C1J C.O ?P1
PF l?> 1.' 1^.« 10 * 0 f>9
171 1.^7 10. f 7.P U5 IP 1* !« O.F" P.179 0.0 1IH 30"
PF 311 3.2 IA.J ?0 « t A«
17? 1.55 «.» 7.9 139 10 1 ?.19 p.?17 0.0*11 P. O ?1H 40°
RP . p{>J r.d 17.5 *0 7 t 57
173 1.41 | l.O «.5 '1 71 ? P.79 p.*?.l O.OP* P-P ??1
SF. ?SI ?.' IA.7 tO i 9 66
17< I.3H II.I (..I 17 M 0 O.31 0.7?& 0.000 C.O 710
FB ?PT ?.7 17.* 7*0 «
.175 I.I/. 11.1 1.1 70 31 P 0.31 0.511 0.000 P.P 79 *
RK 301 3. * 16.5 7 9 0 «9
37« 1.3» 11.4 5.9 PO 35 0 0.3* O.«9« O.OOO P. C 1?^
F> 311 3.A I*. 1 10 I' 0 7'
377 1.17 13.3 3.3 PP 7« 1.10 13.9 ?.0 P5 5' P O.JC 0."5» O.OOP 3.C ff
C.O m ffl ?.5 17.* 10 10 I 39
170 I.JO II.S A.« |« 7* P 0.35 l.«->l 0.000 0.5 30'
FH III 1.5 I«.R 591 P»
1RO 1.07 11.5 l.J AO *0 n 0.«5 O.fl7 0.000 *.O ni
(f 3PT 3.* |A.» ?S |0 I
.v *.i: in.K r.i IP *o o o.'< o. *> P. opr r.* p*i
i-r « 3U ?.* I/.? IP IP i *?
200
-------
TABLE E-3. (Concluded)
0.051m Choke L = 0.50m
M« STOIC, ro? at co NB fnr m NO i«c t»rv. TFC- Swirl
NO. PATIO
3«? l.?0
RP
r«n
H
3X4
F.F
385
RP
316
RR
387
FF
311
1.40
P91
1.37
191
I.W
Ml
I.?S
- 311
1.18
311
t
!?.»
3.1
II. 1
3. A
1 1.3
3.6
l?-0
3.7
12.4
3.9
13.3
4.O
7
>.*
1».4
A. 4
1*.*
6.O
If. 6
i.O
16. &
4.5
IA.4
3.4
IA.4
PP1 PHI F
PI ?#
10 10
:>5. i»
10 «
to ?s
10 5
PO ft
10 5
JO ?6
10 S
?o ??
7 5
pi oi/ .<«<< 0-000 P.O Pof 4
| «0
201
-------
TABLE E-4. FLUE GAS RECIRCULATION DATA OBTAINED IN INSULATED, TUNNEL-FIRED,
0.50 M LENGTH COMBUSTION CHAMBERS WITH THE 0.051 M CHOKE, 40° SWIRL ANGLE FGR BURNER
H
O
4J
I/I
3
1
U
c
M
1
cs
c
i
t
+
0
r
C.
]
(
1
i
e*
c
c
Cycle-Averaged Data
^ Side-Fired *-j
0)
4J
rt c
4-> 00 C
CO I-H O
1
£
rt
* m
> M H-l O
r>
3
Cycle- Averaged Data
RUN ST61C.
NO. RATIO
383 1-13
RR = 33Z
389
RR
390
RR
391
RR
392
RR
393
RR
394
RR
395
RR
396
RR
397
RR
3****
* O
RR
1.36
= 2?Z
1*13
- 31*
1.35
= 3G*
1.21
- 3U
1-2S
= 33X
1.29
- 33X
1-10
= 37t
1. 16
= 3U
1.32
= 31*
1.15
= 31*
C62
X
13. C
4.2
1 1*4
3-4
13.7
3.6
11. 5
3.4
12.6
3.5
12.1
3.9
12. C
3.9
13*6
4.4
13.1
3.7
1 I-C
3.9
13.3
3.9
G2
Z
2.5
16.1
6.0
16.6
2.5
16.4
5.3
16.5
3.9
16.3
5>C
16.1
5.C
16.1
1.9
15.7
3*1
16.4
5.4
16.4
2.9
16.4
CG
PPM
25
10
49
10
1C
7
44
10
23
5
20
10
£0
1C
1043
1C7
33
1C
25
15
1C7
CC
»i f+
IVK/
PPM
37
11
26
10
33
10
32
10
35
10
2G
9
23
9
33
6
35
G
3C
10
35
10
UKC CC NG UHC
PPM GM/KGM CM/KGM GM/KCM
3 C.37 0.597 O.C65
C
16 C.9C C.5C7 0.162
2
6 C.'£C 0.613 C.043
C
30 C.SC C.616 C.311
0
10 C.3G 0.620 C.CS6
C
0 0.34 C.522 G.CCC
0
0 0.34 C.523 C.CCC
0
1 15.11 0.513 C.C12
0
1 0.52 C.573 C.CC9
0
6 0.44 C.573 O.C65
C
2 2.36 0.567 0*017
C
BACH. TFG
SM6KE C
C.O 266
77
C.O 316
62
C.O 307
66
O.C 316
77
C.C 313
71
C.O
68
C.C 2SC
63
C.2 249
54
C.I 232
49
0.1 304
35
0.0 241
6C
202
-------
TABLE E-5. CYCLE-AVERAGED, FLUE GAS RECIRCULATION DATA WITH THE 40° SWIRL,
0.051 M CHOKE FGR BURNER IN THE 0.22 M I.D. INSULATED COMBUSTOR
LO
O
II
<
1
O
-a c
1u «
5" 1
r^ C
U_
1 *~
RR
CO CO _
- ^° °1 402
-»-> O
(O
RR
403
' RR
. <^£ 404
""**-* t/)
, ^ ^ RR
> 5 oo ^05
J Ln r
-^ i
) £
r^ RR
> n ro
> o
CM 408
22
r 31X
1*01
= 32Z
1.04
- 30Z
1.15
- 33Z
1.12
32Z
1.20
- 31Z
1.02
« 33Z
C02
Z
11 .1
3.0
11 .6
3.8
14.1
4.4
12.7
3.8
14.8
3.8
14.6
3.5
13.3
3.6
13.8
4.0
12.5
3-7
14.4
3.4
02
Z
17.3
5.5
16.2
1.9
16.1
4.0
16.4
0.4
16.2
0.9
16.5
2.8
16.1
2.3
16.2
3.6
16.3
0.6
16.1
C0
PPM
40
10
51
15
23
10
20
10
-1600
1499
85
41
28
10
23
10
48
7
-1600
-1600
N0
PPM
40
9
32
10
38
10
35
9
42
5
35
7
31
5
40
1 1
34
10
44
11
UHC CO NO UHC BACH. TFC
PPM GM/KGM GM/KGM GM/KGM SMOKE C
12 0.75 0.808 0.129 0-0 349
0 79
14 0.92 0*622 0.146 0-0 343
0 88
2 i0.35 0-595 0.018 0-0 332
0 79
8 0.34 0*613 0.073 0.0 327
0 82
I -21.51 0.616 0.008 2.3 313
0 66
2 1.19 0.533 0.016 1.2 301
0 49
8 0.44 0.514 0.067 0.3 277
0 57
1 0.35 0.638 0.008 0-0 261
2 57
10 0.78 0.583 0.091 0.6 263
0 52
1 -21.71 0-649 0.012 1.2 227
2 35
203
-------
TABLE E-5. (Continued)
J
C
s
C
C
C
j
0
1-
£
1
O
^
C
CM
^
2
^
r-
U
L
i
j
3
O
n
U
L
>>
X
n. '
3
T
3
ff
£ <
- O
> c
> *J
C
r-
«=C
1 (/)
0 ^
S e
i
in oo
r-v us
o o
i
oo x
*O ^"
O i
Z T-
o o
0
CM
RUN
NO.
409
RR
410
RR
41 1
RR
412
RR
413
RR
414
RR
415
RR
416
RR
417
RR
STOIC.
RATIO
1 .27
- 32%
1.16
= 29%
1.08
= 29Z
1.13
- 30Z
1.10
- 32Z
1.10
= 34%
1.09
- 31%
1.38
- 28%
1.13
= 37%
CG2
Z
12.1
3.4
13.0
3.4
14.7
3.2
13.7
3.2
14.0
4.2
14.1
4.4
14.0
4.0
11 .1
3.6
13-2
4.2
02
%
4.8
16.2
3.2
16.6
1 .7
16.6
2.5
16.5
2*1
16*2
2*1
16.0
1 -9
16.4
6.2
16.7
2-5
15.7
C0
PPM
30
5
-1600
535
1 12
90
43
10
28.
10
27
10
73
5
354
10
80
10
NO
PPM
29
9
11
5
44
10
36
6
34
10
34
11
146
44
95
26
134
43
UHC C0 N0 UHC BACH. TFG
PPM GM/KGM GM/KGM GM/KGM SMOKE C
11 0.52 0.5-40 0.106 0.0 248
0 38
60 -24.64 0.184 0.532 7.4 227
20 35
8 1*61 0*683 0.062 0>3 248
0 38
14 0*66 0*582 0.117 0.1 262
0 46
8 0.42 0.532 0.064 0.0 256
0 46
8 0.41 0.532 0.064 0.0 253
0 46
6 1.07 2.268 0.048 3.9 278
0 46
187 6.53 1*877 1.968 2.6 292
0 57
18 1.21 2*156 0.158 2.8 282
0 57
204
-------
TABLt E-5. (Concluded)
\
*»-
o
1
C
(O
' cf
0
o
JC
-M
-o
*|_
u_
a
t
to
i
CD
S.
r
u_
1
'oJ
C
C
3
1
RUN
N0.
00
^
i-
LJ
Lj
0)
M
S i
O CM
^
2
-i
O
oc
u:
i
CM
-------
APPENDIX F
PRE-PROTOTYPE AIR-COOLED FINNED COMBUSTOR TESTS
In order to ascertain that a heavily air-cooled firebox could indeed
duplicate the emission behavior of the water-cooled fireboxes dis-
cussed in Section IV, before committing the prototype conceptual
warm air furnace design to such a combustor, one was built and tested.
Figure F-l is a layout drawing and Figure F-2 is a photograph of the
combustor, showing the side-fire burner orientation and twenty-four
heavy cooling fins spaced around the combustion chamber perimeter.
The finned section was wrapped with a sheet metal shroud for contain-
ment of the forced air coolant flow. Although air-cooled side-fire
chambers were fired earlier in this task, the chambers used were
simple cylindrical surfaces with only free convection air-cooling
resulting in an inside wall temperature of ~815 C (1500 F). This
side-fired, finned combustor was designed to stay considerably
cooler (by about 500-600 C) during the burner-on time but also to
stay quite warm (-60-100 C) during the burner-off time.
Hot firings in the finned research combustor were conducted with the
same 1.0 ml/s optimum burner described in Section IV. The results are
presented in Table F-l. The parameters that were varied during these
tests were: (1) cycle time, (2) chamber length, and (3) firing rate.
Runs 433 to 446 were made with a 0.75m chamber length and with varying
cycle times (33%-on constant). The 10 minutes-on/20 minutes-off cycles
were typical of earlier data taken, while the 4 minutes-on/8 minutes-off
207
-------
NOTE: Detail dimensions are called out in inches to maintain commonality with material stock.
ro
o
oo
*****>
Figure F-l. Layout Drawing of the Finned Air-Cooled Research Combustor
-------
FLAME
OBSERVATION
PORT
COOLING
FINS (24)
SIDE-FIRE
BURNER
PORT
SHEET
METAL
AIR SHROUD
Figure F-2.Photograph of the Finned Air-Cooled Research Combustor(Shown
Without the Lower Ai> Coolant Inlet Duct)
FORM 608-B-13 REV. 10-73
209
-------
TABLE F-l. CYCLE-AVERAGED FLUE GAS ANALYSIS DATA, OIL BURNER/CHAMBER
MATCHING EXPERIMENTS
Burner: 1.0 ml/s Optimum
Nozzle: 1.0-60-A
MIL
1.000 ML/5
MAY 02, 1975
COMBUSTOR:
0.25m 'x 0.75
STOIC
C0P
7
0
5
T0 NO
PPM PPM
t'Hr C0 N0
PPM GM/KGM GM/KGM
I.I/ 13.5 P. 6
l.PR 1P.1 />.9
1.P3 IP. 5 4.1
17
50
30
8
4?6 1.01 14.5 O.P »1600
! 4J37 1 .08 M.I 1.6 P6
1.11 13.9 P.P P7
; 4»9 1.13 13.6 P.5 3P
p/
35
/?
39
34
4 6?
& 64
LIP 13.7 P.3 PO
1.1P 13.7 P.4 30
1.17 13.1 3.P 3!
3 l.PO IP.6 3-7 37^
I.00 14.5 O.P »1600
1.04 ]4.4 0.9 65
43
41
jI 0.99 '0.64&
:
0*44 jO.ftA/ >:0
4fl6 1.09 13.9 l.B
P5
210
-------
TABLE F-l. (Concluded)
Burner: 1.0 ml/s Optimum
Nozzle: 1.0-60-A, 1.25-60-A, 0.75-60-A
u.
8
S
ui o
IVl
i
c
o
0 U.
^fe
00
t
^
-------
cycles were more representative of current practice. Examination
of the results revealed that the burner/combustor operated well
(low carbonaceous pollutant concentrations) at very low stoichiometric
ratio conditions, but the NO emissions ranged between 0.4 and 0.6
g/kg, and averaged slightly above the target value of 0.5 g/kg. The
heat extracted from the combustion chamber section was on the order
of 8205 J/sec (28,000 Btu/hr), which is nominally equivalent to the
heat extracted from the water-cooled combustor that resulted in
<0.5 g/kg of NO. The difference in the NO concentrations was probably
due to higher temperature gradients (hot spots), that are often en-
countered in air-cooled systems, resulting in slightly higher local
peaks in flame temperature. The finned combustor was of heavy wall
design to reduce hot spotting by increasing heat conduction, however
the temperature gradients were probably still greater than in a water-
cooled system.
The effect of cycle time variation was most noticeable in the carbon-
aceous pollutant concentrations, primarily due to the averaging of the
start "spikes" over varying time spans. As can be seen in Table F-l,
the shorter 4-on/8-off cycle is generally higher in the carbonaceous
pollutant concentrations, due to the shorter time-span averaging.
Runs 447 to 455 were tests with the shorter 0.50m length chamber. As
anticipated, they showed a slight increase in carbonaceous pollutants
and a corresponding decrease ( 5-8%) in nitric oxide emissions. The
shifting of the levels of carbonaceous pollutants resulted in
212
-------
unfavorable combinations (exceeding target specifications) of emission
levels, suggesting that longer (>.50m) chambers should be used.
Variations of firing rate were made to investigate the versatility of
this combustor/burner combination. Runs 456 to 458 were made with a
1.25-60°-A (-1.3 ml/s) oil nozzle and the results showed a corresponding
rise in all pollutant concentrations. The smoke emissions showed the
greatest increase, probably due to oil droplets contacting the cooled
wall at this higher flowrate. Runs 459 to 462 were fired with a lower
flowrate, 0.75-60°-A nozzle (-0.8 ml/s). This reduction in firing rate
did result in a significant reduction in nitric oxide emissions. How-
ever, the carbonaceous pollutant levels increased, but that was probably
due to the off-optimum choke plate rather than to the finned combustor.
It has been noted before that underflring an optimized burner head
results in sensitivity to production of carbonaceous emissions,
especially smoke. The lower nitric oxide levels were interpreted to
mean that a slightly larger diameter combustor would be more suitable
for the 1.0 ml/s firing rate. Examination of the combustor after the
30 test series showed no sooting on the cooled combustor walls.
213
-------
TECHNICAL REPORT DATA
/Please read InUructions on the reverse before completing)
1. RLPORT NO. 2.
EPA-600/2-76-038 j
4. TITLE ANDSU8TITLE
Residential Oil Furnace System Optimization--
Phase I
3. RECIPIENT'S ACCESSIONING.
5. REPORT DATE
February 1976
6. PERFORMING ORGANIZATION CODE
7. AUTHOR(S)
L.P. Combs and A.S. Okuda
8. PERFORMING ORGANIZATION REPORT NO.
R-9815
9. PERFORMING ORSANIZATION NAME AND ADDRESS
Rocketdyne Division
Rockwell International
6633 Canoga Avenue
Canoga Park, CA 91304
10. PROGRAM ELEMENT NO.
1AB014; ROAP 21BCC-27
11. CONTRACT/GRANT NO.
68-02-1819
12. SPONSORING AGENCY NAME AND ADDRESS
EPA, Office of Research and Development
Industrial Environmental Research Laboratory
Research Triangle Park, NC 27711
13. TYPE OF REPORT AND PERIOD COVERED
Final; 6/74-8/75
14. SPONSORING AGENCY CODE
EPA-ORD
15. SUPPLEMENTARY NOTES
EPA project officer for this report is B. Martin, 919/549-8411, Ext 2235.
16. ABSTRACT
The report gives results of an analytical and experimental investigation of
technology for improving pollutant emission characteristics and thermal efficiency of
residential oil furnaces. A digital computer model was programmed for cyclical
(transient) thermal analyses of typical warm air oil furnaces; design features and
operating conditions were varied parametrically to discern influences on thermal
efficiency. Heat in the exhaust flue gases, found to be the major source of ineffi-
ciency , can best be reduced by burning the fuel with minimum excess air and by
reducing the flue gas temperature, both within the practical constraints of emission
production and cost. Furnace operability and pollutant emissions were studied
experimentally by testing three burner types in several combustor sizes, configu-
rations, and wall constructions (cooling methods). Tests showed that the optimized
conventional burner in a larger-than-usual cooled combustor has the best potential
for minimizing emissions and maximizing efficiency. Test results were incorpo-
rated into two conceptual designs for prototype low-emission units capable of
satisfying program goals. One design, a warm air unit, was selected to be built and
tested in Phase U.
7.
KEY WORDS AND DOCUMENT ANALYSIS
DESCRIPTORS
Air Pollution
Furnaces
Fuel Oil
Residential Buildings
Optimization
Thermal Efficiency
Design
Hot Water Heating
Warm Air Heating
b.lDENTIFIERS/OPEN ENDED TERMS
Air Pollution Control
Stationary Sources
Emission Control
Recirculation
c. COSATI Field/Group
13B
13A
21D
13M
12A
20M
13. DISTRIBUTION STATEMENT
Unlimited
19. SECURITY CLASS (ThisReport)
Unclassified
21. NO. OF PAGES
223
20. SECURITY CLASS (Thispage)
Unclassified
22. PRICE
EPA Form 2220-1 (9-73)
214
------- | |