Proceedings of ASME Turbo Expo 2004
Proceedings of ASME
Expo
Power Turbo
for Land,
Sea,2004
and Air
Power
for Land,
and Austria
Air
June
14-17,
2004,Sea
Vienna,
June 14-17, 2004, Vienna, Austria
GT2004-54247
GT2004-54247
MICRO GAS TURBINE COMBUSTION CHAMBER DESIGN AND CFD ANALYSIS
Joao Parente
Thermochemical Power Group
DIMSET, Univ. di Genova
Genova 16147
Italy
Giulio Mori
Viatcheslav V. Anisimov
Ansaldo Ricerche
Genova 16152
Italy
ABSTRACT
In the framework of the non-standard fuel combustion
research in micro-small turbomachinery, a newly designed
micro gas turbine combustor for a 100-kWe power plant in CHP
configuration is under development at the Ansaldo Ricerche
facilities. Combustor design starts from a single silo chamber
shape with two fuel lines, and is associated with a radial swirler
flame stabiliser. Lean premix technique is adopted to control
both flame temperature and NO x production.
Combustor design process envisages two major steps, i.e.
diagnostics-focussed design for methane only and
experimentally validated design optimisation with suitable
burner adaptation to non-standard fuels. The former step is
over, as the first prototype design is ready for experimental
testing. Step two is now beginning with a preliminary analysis
of the burner adaptation to non-standard fuels.
The present paper focuses on the first step of the combustor
development. In particular, main design criteria for both burner
and liner cooling system development are presented.
Besides, design process control invoked both 2D and 3D
CFD analysis. Two turbulence models, FLUENT standard k-ε
model and Reynolds Stress Model (RSM), are refereed and the
results compared.
Here both a detailed analysis of CFD results and a
preliminary analysis of main chemical kinetic phenomena are
discussed.
INTRODUCTION
In the framework of distributed CHP, the utilisation of nonstandard fuel coming form biomass or waste-industrial byproduct seem to be promising in the CO2 reduction issues. The
application of MTG shares different conversion systems that
comes from stand alone unit to more complex integrated
systems based on pressurised fuel cells or gasificator reactor.
This wide variety of both process and fuel requires a great
flexibility design especially for what combustion chamber
concerns.
Giulio Croce
Dipartimento di Energetica e
Macchine
Università di Udine
Udine 33100
Italy
A newly designed micro gas turbine combustor, for a 100
kWe power plant in CHP configuration, is being developed at
Ansaldo Ricerche facilities aiming at the possibility of burning
non-standard fuels in micro-small turbomachinery.
Recuperated micro gas turbines combustor operating
conditions are far from standard applications. In particular, high
inlet temperature leads to higher mixture reactivity, which
significantly affect flame stability, especially from the
autoignition and flashback point of view.
The work here presented is mainly addressed to the
combustor design optimisation and verification, trough
chemical kinetics and CFD analysis, before the incoming
experimental testing.
So, using 0D-2D simplified models, a numerical framework
has been carried out in order to verify and optimise the
combustor design. In particular, the following parameter were
analysed: combustor stability; solid structures maximum
temperature, combustor outlet maximal temperature, and
emissions.
Later, full 3D CFD simulation was performed to verify the
simplified approach.
Furthermore, two different turbulence models are
discussed, the standard k-ε model and the Reynolds Stress
Model (RSM), to assess their influence on flow field prediction.
NOMENCLATURE
A
section [m2]
CHP Combined Heat and Power
D
characteristic diameter [m]
LBO Lean Blow Out
U
velocity [m/s]
RSM Reynolds Stress Model
S
flame velocity [m/s]
t
time [s]
T
temperature [K]
TIT
Turbine Inlet Temperature
u’
turbulent velocity fluctuation [m/s]
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Combustor symmetric air feeding layout was adopted for
minor flow disturbance near the burner and passage inlet
sections.
Combustor outlet section was not yet optimized from the
temperature profile point of view, since turbine connection is
not yet clearly defined. This problem will be focussed on the
second phase of the optimization process.
2s
two step
Greek letters
thermal diffusivity [m2/s]
α
equivalence ratio [-]
φ
Subscripts
avg
average
h
hydraulic
L
laminar
ref
reference
T
turbulent
th
thermal
ARI100 CONCEPTUAL DESIGN
The combustion system (ARI100) was designed for a 100kWe power plant in CHP configuration with an expected overall
efficiency of 80%. Upstream from the combustion chamber, a
recuperative heat exchanger heats compressed air in order to
provide an electrical efficiency value about 30%. An uncooled
single-stage radial expander is used. Then, a 950-°C upper
bound exists on the gas stream temperature at the turbine inlet
section (TIT). This constrain allowed to assess design operative
conditions for the combustor with the help of a thermodynamic
code for optimal electrical efficiency. The main thermodynamic
quantities of the combustion chamber are listed in Table 1:
Table 1: Combustor main thermodynamic parameters.
Air inlet temperature
Air inlet pressure
Overall air-to-fuel ratio
Combustion efficiency
Thermal power
630°C
3.8 bar
6.61
0.98
337 kWth
Figure 1: ARI100 combustor (1: Liner; 2: Burner; 3: Casing; 4: Air
inlets; 5: Boroscope window; 6: Diffusion fuel-line; 7: Premix fuelline; 8: Ignitior).
This first prototype basic design requirements are as
follows:
a) component simplicity
b) easy access to each component
c) variable geometry for further design optimisation.
Accordingly, was adopted the following conceptual design:
(i) single silo chamber shape;
(ii) natural gas feed burner;
(iii) swirled premixed burner stabilised by diffusion pilot flame
to achieve low NOx production;
(iv) liner conventional film cooling.
Figure 1 shows a partial section of the ARI100, and Figure
2 for a detailed view of both liner and burner components.
Considering further optimisation, and the possibility of
testing different configurations, the ARI100 lay-out included
the following features: axial position adjustment of the
diffusion-line fuel injectors (4 in Figure 2); fine tuning of
primary combustion air admitted through the swirl and diffusion
channels (1 and 2 in Figure 3); secondary combustion air
control trough a set of grids.
Figure 2: ARI100 combustor, burner and liner details (1: Film
cooling holes and deviators; 2: Dilution holes; 3: Flameholder; 4:
Diffusion-line fuel injectors; 5: Ignitior; 6: Swirl vanes; 7: Premixline fuel injectors).
1) Burner
ARI100 burner design started from the concept of lean
premix combustion.
Two fuel lines are included in the burner. The main line
(premix line, 7 in Figure 2) was designed for 50%-100% load.
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The secondary line (diffusion line, 4 in Figure 2) was designed
for start-up and partial load operations. The latter line should
also provide a stabilising pilot flame for high-load operating
conditions.
The main components in the burner are: swirl channel (1
in Figure 3); diffusion channel (2 in Figure 3); flameholder (3 in
Figure 2). Both the radial swirler and the flameholder were
included in the burner layout in order to anchor and stabilise the
flame.
Figure 3: ARI100 prototype, burner details (1: Swirl channel; 2:
Diffusion chanel).
Flow swirling arises thanks to 18 flat vanes. The rated
solidity factor of the latter is 1.1, in order to provide both 0.8
swirl number and 46° swirl angle at the channel throat section,
as suggested in [1][2][3]. Premix fuel is injected trough 18
pipes, with 4 nozzles each, at the middle of each swirl vane pair
in the outlet section.
As for further developments towards non-standard fuels,
enough space has been left in the centre axis for one additional
fuel line. This was a main constraint on the flameholder size in
the present burner configuration.
A 0.524 equivalence ratio was assumed for the primary
combustion zone. Air and fuel flow distribution inside burner at
design conditions is 90% and 10%, for the premix and diffusion
channels respectively.
2) Liner
Preliminary liner design assumed 30 m/s average flow
velocity and an area-to-volume ratio close to 3 m-1. CFD
analysis allowed further design of both the dome and the
convergent end section in order to obtain the actual recirculation and hot-product dilution patterns.
In particular, both the film cooling system and the dilution
system were designed with the help of an in-house code, which
has been developed in MathCAD environment.
Various approaches to liner cooling have been analysed in
[1][2] for gas turbine combustion chamber applications. Film
cooling system with splash ring has been selected as a
reasonable compromise between cooling efficiency and
manufactory easiness. Detailed information about film cooling
system is reported in [2].
Conceptually, liner has been subdivided into different
sections (or panels). Each panel starts from the position where
one splash ring equipped with a “band” of cooling holes is
positioned and finishes before the next down-flow splash ring.
A band of dilution holes were taken into account inside a panel.
The in-house code was developed to calculate the set of main
parameters for each panel, using simple engineering relations.
Within the main design criteria, liner maximum temperature of
1000K and 5% of maximum pressure losses across the liner,
this approach allowed evaluating various configurations by
changing input data and automatic updating of the calculation.
In addition, changes during project development could be
managed rapidly and efficiently. This approach to calculation
has been checked by comparing it with example of film cooling
system analysis [1][2].
Globally, the liner consists of tree parts: dome (truncated
divergent cone), straight tube, and truncated convergent cone
(see Figure 2). External casing consists of two main parts: both
of them are truncated divergent cones (see Figure 1).
Liner preliminary design included an overall number of 7
panels, i.e. 2 in the liner dome (“Panel 1 Dome” and “Panel 2
Dome”) and 5 in the straight tube zone (Panel 1 to Panel 5).
The last 3 panels have also a stripe of dilution holes each.
Dome 2nd panel, absent in Figure 2, was excluded from the liner
layout after CFD analysis. Refer to the CFD Section for further
discussion.
CHEMICAL KINETIC ANALYSIS
A main characteristic of recuperated microturbines is the
quite high value (630°C) of design inlet air temperature.
Accordingly, the burner design needs dedicated investigations
focussed on the most relevant kinetic parameters as far as both
safety and system stability are considered.
This analysis was focussed on the following phenomena:
(i) autoignition;
(ii) flashback;
(iii) flame stability (lean blow off).
Autoignition may occur whenever air-fuel mixture
residence time inside the burner is higher than ignition delay.
The ignition delay for the mixture in the premix channel
was evaluated with the help of the SENKIN module of
CHEMKIN II code, a GRI Mech 3.0 complete mechanism and
the corresponding thermodynamic database [4].
Figs. 4 and 5 display the dependence of ignition delay on
the equivalence ratio and temperature, respectively. The values
905 K and 0.5 were assumed as reference values for the inlet air
temperature and the equivalence ratio respectively.
The following is observed:
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(i) ignition delay is just slightly dependent of the equivalence
ratio (this fact is also confirmed in [1] although with
opposite curve slope);
(ii) ignition delay is strongly dependent of temperature.
(iii) ignition delay is about to 1 s in this case; it reduces to 0.1 s
with a 100 K increase of the inlet air temperature.
The mixture residence time inside the swirl channel has
been estimated in 10-3s, which is quite lower than the calculated
ignition delay. Furthermore, with the injection system directed
towards the center axis of the burner and placed downstream the
swirl vanes in the middle of each vane pair, rich spots and
recirculation paths should be avoided. Also, continuous mixture
acceleration along the channel should prevent autoignition.
Refer to [5][6][7] for detailed discussion of flame
propagation trough the boundary layer. Quenching distance is
the most critical parameter. In turn, it can be correlated to
laminar flame velocity SL.
A simplified analysis, based on LAMINAR approach, leads
to the following necessary condition for flashback prevention:
u avg α
(1)
⋅
> 1/ 8
D S L2
Application of a laminar correlation to a turbulent flow
case is somehow controversial. Kroner’s flashback experiments
[6] for a swirled turbulent flow agree with the present laminar
analysis for the vortex breakdown mechanism (iv). The authors
of [6] write down a flashback limit formally equal to (1).
The laminar flame velocity was computed with the help of
the PREMIX module of CHEMKIN II and the GRI Mech 3.0
complete mechanism [4]. The hydraulic diameter of the swirl
channel throat was taken as the characteristic diameter.
Figure 6 displays the flame velocity dependence on the
equivalence ratio at 905 K.
Figure 7 displays the exponential dependence of laminar
flame velocity on temperature.
Figure 4: Ignition delay vs. Equivalence ratio (Tref=905K).
Figure 6: Laminar flame velocity vs. Equivalence ratio (T=905K).
Figure 5: Ignition delay vs. Mixture temperature (φref=0.5).
Flashback can be triggered by one of following
mechanisms:
(i) boundary layer flame propagation;
(ii) turbulent flame propagation through the core flow;
(iii) strong combustion instability;
(iv) combustion vortex breakdown.
It was assumed that flashback could only be induced by the
first (i) mechanism in ARI100. The second mechanism (ii) is of
minor interest, since core flow mean velocity inside the burner
is about 50 m/s, i.e. at least one order of magnitude higher than
the laminar flame velocity. Basically, mechanism (iv) refers to
the EV burner type. Mechanism (iii) is not discussed here.
Figure 7: Laminar flame velocity vs. Temperature (φ = 0.5).
It should be highlighted that relation (1) was developed
within the laminar case for a cylindrical burner [7]. This
relation was used in the preliminary design of the premix throat
section. Later verification/optimisation was performed, using
both CFD and analytical computations. From this optimisation
process two changes were performed in the burner premix
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channel. Firstly, premix outlet channel was passed from a
strictly diagonal channel to a diagonal-cylindrical channel (the
actual profile) to reduce premix outlet radial velocity. In the
previous version, central vortex was too thin and long. Latter
external diameter at the throat section was reduced to ensure
that flashback could be safely avoided at design conditions.
Figure 8 displays the boundary layer radial velocity profile
in the premix channel throat section and the laminar and
turbulent flame velocity. The quenching distance is also
displayed in both laminar and turbulent cases. The origin of the
r-axis was set at the swirl channel-dome edge. The radius is
non-dimensionalised with the premix channel hydraulic
diameter.
The boundary radial velocity profile was computed taking
into account the laminar sublayer, the buffer layer and the
turbulent layer profiles [8]. The shear stress values provided by
the CFD simulations were used. As discussed below, two
turbulence models were utilised, i.e. standard k-ε and RSM. The
following correlation were used to determine the turbulent
flame velocity (2), the laminar quenching distance (3), and the
turbulent quenching distance (4):
  u ′  0.7 
(2)
S T = S L ⋅ 3.5   , correlation extracted from [9]
  S L  
6 α
⋅
, correlation extracted from [9]
2 SL
α
= 10 ⋅
, correlation extracted from [1]
S T − 0.63 ⋅ u ′
d q,L =
(3)
d q, T
(4)
Laminar flame velocity was calculated with CHEMKIN [4]
(see Figure 6 and 7).
According to Figure 8, a quite large flashback safety
margin is expected for the ARI100 burner. Above the quenching
distance, the flow velocity is more than 2 times higher than the
turbulent flame velocity. It should also be noticed that k-ε and
RSM turbulence models lead to quite similar results.
Figure 8: Flashback verification: Flame velocity vs. Radial velocity.
The same approach followed for flashback analysis can be
followed for flame stability analysis. Here, there is a
competition between the reactivity time of the system and the
residence time inside the chamber. The characteristic time
model given in [2] follows the same line of reasoning. This
model was not utilised in the present analysis since it requires
combustor experimental data.
In this case, it was assumed that the characteristic diameter
is equal to the liner hydraulic diameter. Moreover both a mean
temperature (between the inlet air temperature and the adiabatic
flame temperature) and a mean flow velocity throughout the
liner was considered.
A quench value of ≈0.00034 was obtained, which is quite
lower than the limit of Equation (1) - 1/8 -. This leads to the
conclusion that LBO should not occur.
ARI100 CFD ANALYSIS
The CFD analysis invoked the commercial code FLUENT
6.1.22. Two approaches were followed. A preliminary 2D
model and a full 3D model allowed assessment of the upper
bound on liner temperature and angular non-uniformity in the
solid structure. CH4, CO and NOx at the combustor outlet were
also computed.
As for turbulence, both the “standard” k-ε turbulence
model and the Reynolds Stress Model (RSM) were used in two
approaches. The choice for the “standard” k-ε turbulence model
is threefold. Firstly, it is a well-tested and robust model: it costs
about 1/4 of RSM turbulence model in terms of CPU time.
Secondly, familiar combustion models dealing with the
interaction between chemical kinetics and turbulence refer to
the “standard” k-ε turbulence model. And third, the k-ε solution
data is a recommended [10] starting point for RSM calculations.
Alternatively, the Reynolds stress model (RSM) is one of the
most advanced models of turbulence available in FLUENT 6.1.
FLUENT manual [10] strongly recommends using RSM for
cases with swirl number more than 0.5. In the ARI100
combustor swirl number is about 0.8.
Attempt of using “renormalization group” (RNG) k-ε
turbulence model as possible alternative of RSM model has also
been done, but stable solution has not been obtained. The
results obtained with using so-called Realizable k-ε turbulence
model were close enough to results obtained with “standard”
k-ε turbulence model.
The “standard” FLUENT finite-rate/eddy-dissipation
model of combustion with methane-air-2step mixture was used
in all cases.
1) Two-dimensional model
A simplified axisymmetric geometry was used. Continuous
slots replaced liner holes, preserving actual cross-section area.
The computation domain included neither the fuel injection
system nor the swirler - see Figure 9.
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Figure 9: Computation domain for the 2D axisymmetric model.
For the flow across “Mixture Inlet 1”, the weights of the
velocity component normal to the boundary and the swirl
component are 1 and 0.8 respectively, according to swirler
design requirements. Mass flow, temperature, flow composition,
and operating pressure were set at combustor design condition.
The mesh was created with Gambit 2.0 software. The grid
had 27439 cells and 29010 nodes. A “fine” structured mesh and
an “coarser” unstructured mesh were used in the “combustion”
region and in the “dilution” region respectively.
Four cases are here reported. In the first two cases (Case 1
and Case 2) the “standard” k-ε turbulence model has been used,
in Case 3 and 3a it was used the RSM.
In Case 3a, a mesh refinement has been done based on case
Case 3 results in regions:
•
•
•
•
•
where second reaction rate values are higher than 5% of its
maximum value;
where swirl velocity gradient is higher than 10% of its
maximum value in the combustion zone;
near axis in dilution zones;
near wall regions in the swirl channel, where y+ is higher
than 60;
where 2D cells volume gradient is higher than 50% of its
maximum value.
The sell sizes have been decreased two times, except near
wall regions in the swirl channel, where sell sizes have been
decreased four times. This mesh has 38731 cells and 41116
nodes, that is about 40% more than in Cases 1-3.
Rates of reactions for Case 1 are presented in Figure 10.
The reactions considered for the methane-air-2step mixture are:
•
Reaction 1: 2CH 4 + 3O2 → 2CO + 4 H 2 O
•
Reaction 2: 2CO + O2 → 2CO 2
Figure 10 indicates that a flame front cannot be anchored
near the burner-dome edge. Temperature and velocity maps for
the Case 1 demonstrated that the combustion reactions were
quenched near the burner-dome edge, due to the excessive
cooling flow. To prevent this situation, the first liner cooling
hole stripe (“Panel 2 Dome”) was eliminated. This new
“optimised geometry” was studied in the Cases 2, 3 and 3a.
Figure 11 displays reaction rates for Case 2 (k-ε).
Figure 10: Case 1 maps of rate of reaction 1 and 2 [kmol/(m3s)].
Figure 11: Case 2 maps of rate of reaction 1 and 2 [kmol/(m3s)].
Figs.10 and 11 put in evidence the effect of the “Dome
Panel 2” film cooling on the combustion reactions. Comparison
of mass-averaged mole fractions of CH4 and CO at the
combustor outlet leads to relevant results (Table 2).
Table 2: 2D Results: Emisions in
ppmvd (@15% O2).
Case
CH4
CO
NOx
1
2
3
3a
1728
64.2
4.18
4.38
607
21.2
1.18
1.21
2.07
2.78
2.11
2.35
The values listed in Table 2 show that:
(i) As for Case 1, CH4 and CO concentration values at the
outlet section are quite high;
(ii) In the Case 2 and Case 3 CH4 and CO emissions were
substantially improved.
(iii) Differences between Case 3 and Case 3a results are not
significant.
The emissions improvement is due to the fact that
combustion takes place in the external vortex region too. This is
in agreement with the temperature map for Case 3 (Figure 12).
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(ii) As for the liner maximum temperature, the difference
between RSM and k-ε models is quite small (max.≈2[K]).
Figure 12: Case 3 (RSM) temperature map [K].
Maximum temperature values in solid structures are listed
in Table 3, which also reports the fluid mass-averaged
temperature at different combustor cross sections (see Figure 9).
Differences between Case 3 and 3a results (Table 2 and 3)
are not significant, this indicates that Case 3 is not grid
dependent.
Figure 14 shows that Case 2 flow characteristics differ from
Case 3 flow characteristics. There are significant differences in
both vortex shape and dimension. When RSM is used, then the
central vortex becomes about 30% longer and about 30% less
intensive than in Case 2, where the standard k-ε turbulence
model was used. This result explains the discrepancies shown in
Table 2 for these two models. Incoming experimental data
should allow assessing the most suitable model.
Table 3: 2D Results: Temperature [K].
Case
Ext. Liner
Surf. Max
Diff.
Channel
Max
Swirl
Channel
Max
Casing
Max
1
954
1216
904
926
2
987
1231
964
930
3
989
1197
996
928
3a
989
1188
997
928
Case
Outlet
Max
Max
Com.-Dil.
Interface
Outlet
1
1589
1978
1298
1228
2
1625
1995
1573
1250
3
1679
1958
1652
1248
3a
1687
1958
1650
1248
For cases 1-3, Figure 13 displays the temperature
distribution on the external surface of the liner.
Figure 13: Liner external temperature profile.
The values listed in the third column of Table 3, together
with the temperature profiles displayed in Figure 13, show that:
(i) The maximum value of the liner temperature is lower than
1000 K in all cases, even if, the flame is anchored at the
burner-dome edge for both cases 2 and 3;
Figure 14: Case 2 and Case 3 path line map.
Post-processing of flow, temperature and concentration
fields - reported heretofore with the thermal NOx formation
model available in Fluent 6.1 [10] - allowed to compute thermal
NOx concentrations for all cases. Both O and OH
concentrations were computed with the partial equilibrium
approach [10]. Computed NOx concentrations at the combustor
outlet are reported in Table 2. These values show the negligible
effect of reported design variations on NO x emission.
Lean-premixed combustion of methane may lead to
significant amount of non-thermal NOx [1]. A prompt NOx
formation post-processor mechanism is also available in
FLUENT 6.1 [10]. It relies on an overall equivalence ratio for
the flame. If φ=0.524 (ARI100 φ at design condition), the
mechanism gives negative rate of prompt NOx formation. For
this reason, only thermal NOx formation mechanism was
considered at the moment.
2) Three-dimensional model
The adopted geometry for the 3D model is shown in Figure
15, a 60° periodicity was used. As in the two-dimensional
analysis, liner cooling holes were substituted with continuous
slots, preserving cross section areas of holes, while discrete
dilution holes were retained, thus allowing for non uniformity in
the tangential direction; the swirler has not been included in the
computation domain.
As in the 2D model, implemented boundary conditions
reflect the ARI100 nominal condition. Here, both velocity and
mass fraction radial profiles in the premix channel throat were
defined from 2D CFD analysis.
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It should be noticed that in the 3D case diffusion fuel
injector was included in the computation domain, instead, in the
2D model, a premixed air-fuel mixture was considered.
holes, and section θ=45° crosses at the centre of the first
dilution hole.
All sections exhibit similar temperature profiles up to
~55% axial co-ordinate, near the first dilution holes row.
Moving downstream, angular non-uniformity of temperature
becomes significant. Major difference is observed at 90% liner
length, close to the second dilution holes row. A peak
temperature appears in this zone at both θ=30° and θ=45° when
k-ε is used (clearly illustrated in Figure 16). Such peaks (even if
less sharp) appear with RSM too.
Figure 15: Computation domain for the 3D model.
Mesh was created using Gambit 2.0 software. The grid has
477985 hexahedral cells and 538756 nodes. Both structured and
unstructured mesh was used in different zones of computational
domain. It is to note that 3D mesh is ≈30% coarser than 2D
mesh to reduce computational cost. Calculations with refined
mesh are foreseen in the future to ensure the mesh
independence of 3D results as it has been done for 2D results.
Here, two cases are discussed. Case 4 and 5 invoke k-ε
standard turbulence model and RSM respectively, with the same
discretisation and reaction schemes of the 2D cases.
Figure 16 displays the temperature map in an iso-rate-ofreaction-1-surface (2% of the maximum rate of reaction) for
Case 4. The 3D reaction rate of Figure 16 is consistent with the
corresponding 2D quantity of Figure 11.
Figure 17: Liner external temperature profiles at diferent crosssections (θ=15°, θ=30° and θ=45°).
Maximum metal temperature for both cases 4 and 5 are
listed in Table 4, which also reports the fluid mass-averaged
temperature at different cross sections of the combustion
chamber.
Table 4: 3D Results: Temperature [K].
Figure 16: Temperature map in a iso-rate-of-reaction-1-surface
equal to 2% of the maximum reaction rate (Case 4, k-ε).
Formally, 2D and 3D flow path are identical. The same
differences between the k-ε standard turbulence model and the
RSM observed in Figure 14 were also observed in the 3D
analysis (not shown here).
As for the effect of flow across the dilution holes on the
liner temperature, see Figure 17. It displays the liner
temperature at three different constant-θ sections for both Cases
4 and 5. Section θ=15° crosses at the centre of the second
dilution hole, section θ=30° passes midspan between dilution
Case
Ext. Liner
Surf. Max
Diff.
Channel
Max
Swirl
Channel
Max
Casing
Max
4
1056
1210
1055
965
5
1029
1232
1076
956
Case
Outlet
Max
Max
Com.-Dil.
Interface
Outlet
4
1397
2151
1480
1244
5
1384
2322
1501
1245
Comparing the 3D results of Table 4 with 2D ones (Table
3), higher liner temperature values are observed for the 3D
model. Such a difference can be justified by the fact that 2D
dilution jets act also as liner cooling air. The outlet maximum
temperatures are much lower than in 2D case, due to deeper
penetration of dilution air in the core flow, associate to the
higher jet hydraulic diameter in the 3D cases.
In both 3D and 2D model y+ values are always within the
range for the application of the wall functions except for the
region near the flameholder, where they are too low. For this
reason, values reported in Tables 3 and 4 for the diffusion
8
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channel maximum temperatures are less precise than other
values here reported.
Figures 18 and 19 illustrate the dilution flowpath lines and
liner wall and outlet section temperature map viewed from the
combustor outlet, for the Case 4 and Case 5 respectively.
design, dilution jet penetration is about 1/3 of liner diameter.
The reported differences between the results of the k-ε and
RSM model was to be expected, as the standard k-ε turbulence
model overestimate round jet dispersion [10].
Based of the results here presented, the pattern factor is
0.45 and 0.41 for Cases 4 and 5 respectively. In the case of an
actual integration with a turbomachine, an optimisation of the
combustion products dilution system will be required.
CO and CH4 emissions at combustor outlet section are
listed in Table 5 for both Case 4 and Case 5.
Table 5: 3D Results: Emissions ppmvd
Case
Figure 18: Liner wall and outlet section temperature map and
dilution phatlines: view from combustor outlet - 1: 1st dilution
holes row; 2: 2nd dilution holes row; 3rd dilution holes row - (Case
4, k-ε model).
CH4
(@15% O2).
CO
-2
4
1.09x10
5
6.59x10-2
1.59x10
NOx
-2
7.02x10-2
9.1
34.7
From the comparison between the results of Table 2 and
Table 5, it is shown a decrease in the CH4 and CO emissions for
the 3D cases. Instead, NOx emissions are higher when the 3D
model is used. Such a difference between the 2D and 3D
models can be explained by the higher local equivalence ratio,
hence higher local temperatures (see Table 3 and Table 4),
shown in the 3D case. In fact, this is a direct consequence of the
choice to use of a premixed inlet for the diffusion line in the 2D
model. This result is extremely important, since it indicates that
at full-power conditions, NOx emissions could be reduced if the
diffusion flame is turned off and, as shown by the 2D results,
without risking flame blow-off. Nevertheless, such condition
must be carefully analysed during the incoming experimental
tests.
It should be noticed however that the values here presented
for the emission should be considered with all the uncertainties
inherent to the calculation. Most of all, these values are an
indication of what it could be expected from the first
experimental testing.
CONCLUSIONS
Today, the first step towards the development of the
ARI100 combustor is over. It focussed on proper setting of the
combustor design for diagnostics purposes with methane only.
Both combustion chamber design criteria and CFD analysis
were discussed. Main conclusions are as follows:
Figure 19: Liner wall and outlet section temperature map and
dilution phatlines: view from combustor outlet – 1: 1st dilution
holes row; 2: 2nd dilution holes row; 3rd dilution holes row - (Case
5, RSM model).
These figures show that the dilution flow affects the
combustion chamber flow-pattern deeply.
When RSM is used, then a deeper penetration of the
dilution jet is obtained. Consequently, temperature gradients at
the outlet section are weaker than in k-ε results. RSM results
agree with the design criteria invoked in the dilution system
(i) Generally speaking, CFD simulation validated ARI100
design criteria. Even though, experimental testing is still
needed for further verification of the obtained results, in
particular from the temperature field, flow field and
emissions point of view.
(ii) Chemical kinetic analysis allowed foreseeing a stable and
safe behaviour of the ARI100 combustor in future
experiments. In fact, no autoignition, no flashback, and no
lean blow out were shown by the CFD simulations
discussed here;
(iii) 2D CFD analysis turned out to be an efficient and fast
approach to verify combustion chamber design. According
9
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to the above discussed results, a good agreement exists
between 2D and 3D model, although 3D simulation is
always required if angular non-uniformity are to be
investigated;
(iv) Results of the k-ε standard turbulence model differ
considerably from RSM results, especially when the central
vortex and the penetration of the dilution flow jet are to be
described. Experimental testing should allow assessing the
most suitable model for the CFD analysis of this type of
MGT combustors.
Furthermore, a Chemical Reactor Model (CRM), based on
the Chemkin modules, is now being developed for the emissions
estimate. Using CFD results to set a reactor model layout, more
complex mechanisms, as GRI Mech 3.0, can be used to evaluate
combustor emissions, particularly NOx emissions. Additionally,
incoming experiments should allow further investigation aimed
at model checking/comparison for turbulence, combustion and
NOx formation.
REFERENCES
[1] A. H. Lefebvre, “Gas Turbine Combustion”, Taylor and
Francis, 1998.
[2] A. M. Mellor, “Design of Modern gas Turbine
Combustors”, Academic Press, 1990.
[3] A. K. Gupta, D. G. Lilley, N. Syred, “Swirl Flows”, Abacus
Press,1984.
[4] R. J. Kee, F. M. Rupley, J. A. Miller, “Chemkin-II: A
Fortran Chemical Kinetics Package for the Analysis of Gas
Phase Chemical Kinetics”, SANDIA REPORT, SAND898009B, UC-706, Unlimited Release, Sandia National
Laboratories. Livermore, 1994.
[5] I. Glassman, Combustion, Academic Press, 1977.
[6] M. Kroner, J. Fritz, T. Sattelmayer, "Flashback Limits for
Combustion Induced Vortex Breakdown in a Swirl Burner"
ASME paper GT2002-30075.
[7] B. W. Lewis, G. von Elbe “Combustion Flames and
Explosion of Gases”, 2nd ed., Academic Press, 1961.
[8] J. P. Holman, "Heat Transfer" 7 th ed., McGraw Hill, 1990.
[9] S. R. Turns, “An Introduction to Combustion, Concepts and
Applications”, 2nd ed., McGraw Hill, 2000.
[10] . FLUENT 6.1 “User’s Guide Volume 2”, Fluent
Incorporated, 2003.
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