Radial Swirler Designs for Ultra-Low NOx Gas Turbine

Combustors

Gordon.E. ANDREWS, Nick. ESCOTT and Michael.C. MKPADI

Energy and Resources Research Institute School of Process, Environment and Materials Engineering

The University of Leeds, LEEDS, LS2 9JT, UK

Tel. 0113 2332493 e-mail g.e.andrews@leeds.ac.uk profgeandrews@hotmail.com

All correspondence to Prof. Gordon E. Andrews

Abstract Low NOx lean well mixed combustion for gas

turbine primary zone applications were investigated using four radial swirler designs. Each swirler had eight vane passages and the four passage designs were curved, aerodynamic tapered flat bladed, rectangular and circular passage cross sections. The swirler exit diameter d to the combustor diameter D expansion ratio D/d of 1.84 was used with a swirler exit diameter of 76mm and combustor diameter of 140mm. Three methods of fuel injection were investigated: single hole radial vane passage injection, swirler outlet throat 8 hole wall injection and 8 hole radially outward central injection. The results showed that the different radial swirler designs did not have a strong influence on the NOx emissions. All radial vane swirler passage designs gave minimum NOx emissions in the 1-2 ppm range for lean combustion at 600K and 740K inlet temperature and atmospheric pressure. There was evidence that the radial swirler design influenced which fuel injection location, passage or exit throat wall, gave the lowest NOx emissions. Introduction

Alkabie and Andrews (1988, 1989, 1990) were the first to show that radial swirlers with passage fuel injection followed by a dump expansion could achieve ultra-low NOx emissions for lean combustion. Low NOx combustor designs based on this work were put into production by Siemens in their 1-15 MW range of gas turbines and demonstrated very similar NOx emissions at pressure, to those first demonstrated by Alkabie and Andrews at atmospheric pressure (Andrews et al., 1999).

In this work the original one fuel hole per radial vane passage was used to investigate different ways of manufacturing the radial passage: curved passage (RS1), parallel passage (RS5), circular passage (RS6), aerodynamic tapered flat blade passage (RS2). The swirler exit diameter d to the combustor diameter D expansion ratio D/d of 1.84 was used with a swirler exit diameter of 76mm and combustor

diameter of 140mm. All injection locations were suitable for liquid fuels, but only natural gas was used in the present work. Alkabie and Andrews (1988, 1989, 1990, 1991a) showed that radial swirlers with passage and outlet wall fuel injection give a NOx performance close to that of premixed gases, but with a better flame stability margin. They also showed that the same basic design gave very lo w NOx emissions for kerosene and gas oil (Alkabie and Andrews, 1989, 1990; Andrews et al., 1993). This initial work was carried out at relatively low swirler air flow capacities, corresponding to only 40% of the combustor air flow in the primary zone, which would only occur at low power conditions. Larger flow capacity swirlers have been investigated and shown to give very similar low NOx results (Escott et al. 1993; Andrews et al., 1999). The present work is a continuation of this research into high flow capacity radial swirlers.

The work of Alkabie and Andrews is the basis of the Siemens low NOx gas turbine designs for the Tornado, Typhoon, Cyclone and Tempest series of industrial gas turbines. The design principles adopted in these industrial designs are those developed by Alkabie and Andrews, as outlined by Alkabie (1998). The Siemens designs are set out in a series of papers by Boyns and Patel (1997), Gallimore et al., (1997), Hornsby and Norster (1997), Scholtz and Depetri (1997). The main change in their design from the original Alkabie and Andrews design was to change from a curved bladed radial vane passage to a parallel passage blade, following suggestions to them by Andrews (1991). The performance of this type of blade passage is the subject of the present work, but with single hole fuel injection per passage.

The second change made by Siemens was to increase the number of fuel injection holes and to locate them in the side walls of the swirler vane passage, so that there was no flow blockage by the radial fuel injector. This took a lot of development to get the fuel/air mixing optimised and used several fuel holes per passage. However, the NOx results were no better than the original Alkabie and Andrews (1988) results at the same flame

temperature. Thermal NOx formation was small at these temperatures (<1850K) and any influence of pressure on the results was also small. Andrews et al. (1999) showed that the two sets of NOx results at the same inlet temperature and flame temperature condition were very similar. Thus the increased number of vane passage fuel injector holes had no great influence on the NOx emissions.

Alkabie and Andrews (1989, 1990) and Andrews et al. (1993) were the first to show that this type of single fuel injector per radial passage or swirler outlet plane fuel injection could achieve low NOx with low CO and UHC emissions for liquid fuels kerosene and gas oil. The Siemens design was also shown to capable of dry low NOx performance with liquid fuels using similar vane passage wall fuel injection as for gaseous fuels (Alkabie et al., 2001).

The authors have continued with the use of a single fuel injector hole per passage as this was the simplest system for liquid fuels and had been shown to have similar NOx and flame stability as gaseous fuels. Also the authors have used gaseous and liquid fuel injection through the wall of the swirler outlet throat (Alkabie and Andrews,1989), which is a location that prevents flashback or autoignition with liquid fuels. Both of these fuel injection locations are used in the present work. Radial Vane Swirler Designs

Four radial vane swirler designs were investigated and their design details are given in Figs. 1-4. The four designs were: curved blade radial swirler (RS1); straight sided aerodynamic profile radial swirler (RS2); parallel passage radial swirler (RS5) and circular passage (RS6). The advantage of the curved passages and the flat aerodynamic bladed passage was that the radial vane passage inlet had a greater area than the outlet and this allowed a fuel injection pipe to be inserted into the inlet air for the single point centre-line fuel injection to be achieved without any increase in the pressure loss. The flat bladed aerodynamic swirler had the additional advantage that by making the blade moveable around a pivot, the radial swirler could be made into a variable geometry swirler which would be useful in variable geometry combustor applications.

Fig. 1 Curved blade radial swirler (RS1)

Both RS1 and RS2 swirlers were designed to

have a high discharge coefficient. The aim with RS1 was to minimise flow separation due to the flow interacting with the blades. For RS2 the

aerodynamic inlet with no sharp edges was expected to improve the discharge coefficient and to reduce turbulence in the vane passage. Discharge coefficient measurements were made for each radial swirler using the combustion test equipment without combustion. The measurements were made using an upstream axial flow rather than the reverse flow used in the work of Boyns and Patel (1977). These measurements were made with the 140mm downstream combustor in place.

Discharge coefficients were also made with the 76mm diameter wall fuel injector in place. This was 40mm long and effectively added a throat to the swirler exit. A very similar swirler exit throat was used by Boyns and Patel (1977), but as an expansion joint for the flame tube.

The discharge coefficient results are shown in Table 1, which also lists the minimum flow areas of each swirler. The product of the discharge coefficient and the minimum flow area gives the effective flow area of the swirler, which controls the pressure loss. The geometric swirl number is also given (Alkabie and Andrews, 1991b), which was similar for all swirlers. For all swirlers this was high enough to generate a central reverse flow.

Fig. 2 Aerodynamic profile flat bladed radial swirler (RS2) Table 1 45o Radial Swirler Discharge Coefficients. Radial Swirl

Flow Area mm2

Cd dump expansion 140mm combustor

Cd with 40mm long 76mm discharge throat

Geometric Swirl Number

RS1 2162

0.62 0.60 1.30

RS2 2213

0.63 0.59 1.32

RS5 1700

0.81 0.74 1.27

RS6 1838

0.73 0.72 1.31

The discharge coefficient results show that the curved passage of RS1 and smooth inlet profiles to RS2 did not result in the anticipated high discharge coefficient. Both of these swirlers basically behaved as sharp edged orifices. It should be noted that the swirler 76mm diameter outlet diameter had an area of 4536 mm2 , which is greater than the minimum flow area in the vane passages. Hence, the discharge

coefficient is controlled by the aerodynamics in the vane passages. Flow visualisation work was undertaken on RS1 and this showed flow separation from both the back face of the swirler and from the vane surface. This produced a vena contraction in the passage, which was not sufficiently long to achieve flow reattachment to the vane walls. The discharge coefficients for RS5 and RS6 were significantly higher than for RS1 and RS2, even though the inlet geometry to both radial passages was sharp edged. The improved discharge coefficient was due to flow re-attachment within the vane passage. This occurs in short tubes at a tube length to diameter ratio of 0.8 and this was exceeded in the RS5 and RS6 designs. A tube with a sharp edged inlet and an L/D of greater than 0.8 but shorter than about 2 has a discharge coefficient in the 0.8-0.84 region (Andrews and Mkpadi, 1984). RS5 behaved like this and this is the passage shape used by Boyns and Patel (1997). For circular holes the diameter had to increase relative to the cross sectional width of RS5 and the passage L/D was shorter. There was a reduction in Cd as a consequence.

For all the swirlers except RS5 there was little change in the discharge coefficient when the downstream 76mm diameter wall fuel injector throat was added. For RS5 there was a large reduction in Cd, indicating a greater aerodynamic influence of the 76mm throat discharge duct. This implies an aerodynamic interaction between the vane passage flow and the throat , which is only present for the parallel passage RS 5 radial swirler. The reasons for this effect are not known. The RS2 blade type has been used by Willis et al (1993) and Scarinco and Halpin (1999) in the RB211 and Trent industrial gas turbines. In these applications two radial swirlers of this design were mounted in a counter-rotating radial swirler format in the primary zone of an axially staged combustor. The performance of the RS2 swirler in a counter-rotating format has been investigated by Andrews et al. (1995) and Escott et al. (1995). This work showed that there was no advantage in terms of the NOx emissions or flame stability of the counter-rotating swirl configuration, compared with a single radial swirler with passage fuel injection. There were disadvantages in that the flame stability was inferior, the NOx higher and there were associated acoustic problems. Andrews and Mkpadi (2001) have also investigated the addition of a central injection pilot to this configuration to enhance the flame stability and to increase the simulated engine power turndown. A central fuel injector pilot is also used by Willis et al. (1993) and Scarino and Halpin (1999).

The parallel passage rectangular cross section vane passage, RS5, had the disadvantage that the insertion of the fuel injection tube decreased the passage flow area and increased the pressure loss. Also the maximum velocity in the passage was in the plane of the fuel injector and the induced turbulence in the passage, although good for improving fuel/air mixing, would increase the risk of flashback into the passage. In the Boyns and Patel (1997) radial swirler design, which has a parallel rectangular passage design, this problem

was overcome by using fuel injection through the vane passage wall.

The circular passage design, RS6, was investigated because of its potential to be made into a variable geometry swirler by inserting a butterfly valve at the inlet to the passage. It has the same

Fig. 3 Parallel passage radial swirler (RS5) problem over the increase in the pressure loss due to the presence of the fuel injector, as for the rectangular passage design. However, it is considered that the radial passage fuel injector does not have to be located inside the passage and a fuel injector could simply be located in the plenum chamber to direct fuel at the centre-line of the radial passage inlet. This would then cause no blockage and no increase in pressure loss. Also with an inlet butterfly valve it would act to improve the fuel/air mixing in a similar way to port fuel injection in a spark ignition engine. It could also provide a simple method of adapting radial swirlers to operate on liquid fuels.

Fig. 4 Circular passage radial swirler (RS6) EXPERIMENTAL EQUIPMENT The radial vane passage fuel injection was a single hole on the centreline of the vane inlet directed in the same direction as the vane airflow. The injection hole size was 3mm diameter and the fuel spokes were 3.2mm outer diameter. The 76mm diameter swirler outlet plane wall fuel inject or consisted of a 40mm long 76mm diameter cylindrical tube, which formed a throat to the outlet of the swirler. The natural gas fuel was injected through eight equispaced 3mm diameter holes, located 20mm from the throat inlet and inclined 30o towards the upstream flow.

The atmospheric pressure combustion test facility consisted of an air supply fan, venturi flow metering, electrical preheaters, 250mm diameter air plenum chamber, flame stabiliser, 330mm long

140mm diameter uncooled combustor followed by a bend in the water cooled exhaust pipe with an observation window on the combustor centre line. The combustor length had no film cooling, as this increases NOx emissions (Andrews and Kim, 2001). Also the length was the minimum primary zone length used in industrial gas turbines. Any dilution air used in a full combustor design would be after the 330mm length of the present work. The combustor was instrumented with wall static pressure tappings and thermocouples so that axial profiles of temperature and static pressure could be measured. These were used to determine the recirculation zone size and the axial location of the heat release. Mean exhaust plane gas samples were obtained using a 40 hole water cooled 'X' configuration probe. The sample gases were passed into a heated sample line and on through a heated filter and pump to a heated gas analysis system. The gas analysis results were processed to provide a computed air to fuel ratio, combustion efficiency, mean adiabatic flame temperature and various pollution parameters. A chemiluminescence NOx analyser was used with a minimum scale of 1-4 ppm with a 0.05ppm resolution. The thermoelectrically cooled photomultiplier tube had a 40ppb noise level. Natural gas was supplied from the mains via a boost pump and was analysed for use in the gas analysis computations. WEAK EXTINCTION RESULTS The weak extinction results for the four radial swirlers were determined for 600 and 740K inlet temperatures and a 0.03 reference Mach number. This Mach number represents 60% of the combustor total air flow in the primary zone and is typical of the primary zone air flow requirements at maximum power in industrial gas turbines. This was also the flow proportion used by Boyns and Patel (1997). The results are shown in Table 2. Table 2 Weak Extinction at M=0.03. Fuel Injection Mode

Inlet Temp.

W.E. EQR

W.E. AFR

Pressure Loss %

Passage RS1

600K 740K

0.37 0.32

44.6 51.6

7.1% 7.2%

Wall RS1

600K 740K

0.29

57.0

6.2%

Passage RS2

600K 740K

0.37 0.32

44.6 51.6

6.9% 6.9%

Wall RS2

600K 740K

0.29

57.0

6.1%

Passage RS5

600K 740K

0.38 0.33

43.2 50.7

8.7% 9.1%

Wall RS5

600K 740K

0.41 0.34

40.9 48.9

8.9% 9.3%

Passage RS6

600K 740K

0.38 0.33

43.3 50.2

8.6% 9.2%

Wall RS6

600K 740K

0.38 0.34

44.0 48.9

8.7% 9.4%

These results show that there was little influence of the radial swirler design on the weak extinction

for the same mode of fuel injection. At 600K with passage fuel injection the least stable swirler was RS5 and RS6 at 0.38 equivalence ratio and the most stable was RS1 and RS2 at 0.37. At 740 K the least stable was RS5 and RS6 at 0.33 and the most stable was RS1 and RS2 at 0.32. For wall fuel injection at 740K RS5 and RS6 were the least stable at 0.34 and RS1 and RS2 were the most stable at 0.29.

These results show only a small difference in the weak extinction between the swirlers and for the different modes of fuel injection. This was because each swirler created a similar dump expansion shear layer with similar turbulence intensity.

Table 2 also shows a similar pressure loss for each design and this determines the energy from the flow that is used to create turbulence. The pressure loss was higher than usual for a gas turbine. This was because the swirlers were designed with an assumed Cd of 0.9. Escott et al. (1993) have shown ,using radial swirlers with a greater depth and fkiw area, that this higher pressure loss did not influence the weak extinction or the NOx emissions. Future work will investigate variable geometry versions of these radial swirlers, which will allow the flow area to be varied and lower pressure losses achieved. All of the weak extinction results were outside but close to the fundamental flammability limit for hydrocarbon-air mixtures at these inlet temperatures. At 600K this is 0.46 and at 740K it is 0.40. The worst flame stability in the present work was 0.41 at 600K and 0.34 at 740K. Thus the direct fuel injection in the present work using one fuel hole per injector produced an enhanced flame stability relative to the premixed case. This indicates some unmixedness in the shear layer to produce the enhanced stability. The results will show that this enhanced flame stability allowed a minimum NOx to be achieved at a primary zone condition well away from the flame stability. This led to no acoustic flame instability problems in any of the radial swirler designs tested, provided the primary zone was lean enough to achieve low NOx emissions. AXIAL WALL TEMPERATURE PROFILES The axial wall temperature profiles for RS1, RS5 and RS6 are shown in Figs. 5 and 6 for vane passage and wall fuel injection. These results show a difference, in that for RS1 wall injection has lower temperatures in the near swirler region. If this was due to fuel/air mixing effects then there would be a significant effect on flame stability, which there is not. The wall temperature profiles indicate a slower development of the flame with wall injection and the NOx results, below, show lower NOx as a result.

Fig. 6 shows that for RS5 and RS6 the wall injection has higher temperatures in the recirculation zone near the swirler than does passage injection. The weak extinction results again show no significant difference and hence there is unlikely to be any major fuel/air mixing differences. These temperature profiles indicate a slower flam e development with passage fuel injection, which results in lower NOx, as shown below.

Fig. 5 Axial wall temperature profiles for RS1

Fig. 6 Axial wall temperature profiles for RS5 &6 RS1 RADIAL SWIRLER EMISSIONS The combustion inefficiency and NOx emissions results for RS1 are shown in Figs. 7-9 for 600K and 740K inlet temperature. The results show slightly lower NOx emissions for wall fuel injection at 740K but there were slightly higher emissions at 600K. Both methods of fuel injection gave excellent fuel/air mixing and low NOx emissions. At 740K wall injection achieved 1ppm NOx at 15% oxygen with a 0.41 equivalence ratio and a combustion inefficiency of 0.1%. Passage injection achieved 1.8 ppm with the same low inefficiency at 0.4 equivalence ratio. NOx emissions at 740K were below 10 ppm at 15% oxygen with passage injection, up to an equivalence ratio of 0.55 (1980K). This would give a main burner power turndown on fuel flow of 27% before fuel staging to a central burner was required. The high er NOx emissions for central injection are clearly shown in Fig. 8, but the weak extinction was 0.27, which will give a higher power turndown.

Fig. 7 Combustion inefficiency as a function of equivalence ratio for RS1 at 740K.

Fig.8 NOx emissions as a function of adiabatic flame temperature for RS1 at 740K.

Fig. 9.NOx emissions corrected to 15% oxygen as a function of the combustion inefficiency for RS1 at 740K. RS2 RADIAL SWIRLER EMISSIONS The RS2 radial swirler combustion inefficiency and NOx results are shown in Figs. 10-12 for 600K and 740K. At 740K wall injection had slightly lower NOx emissions than for vane passage fuel injection. NOx emissions as low as 1.1 ppm at 15% oxygen were achieved with wall injection at 0.405 equivalence ratio with a combustion inefficiency of 0.06%. For passage injection very similar results occurred with 1.6 ppm at 0.395 equivalence ratio with 0.08 combustion inefficiency. The inferior flame stability at 600K increased the minimum achievable NOx emissions to 2.0 ppm at 0.48 equivalence ratio. However, the NOx emissions at 600K and 740K for passage injection were very similar indicating good fuel and air mixing, as shown in Fig. 11. These are very similar optimum conditions to those for RS1 and the superior performance of wall fuel injection was similar.

Fig. 10 Combustion inefficiency as a function of the equivalence ratio for RS2 at 600K and 740K.

Fig. 11 NOx emissions as a function of the adiabatic flame temperature for RS2 at 600K and 740K.

Fig. 12 NOx emissions corrected to 15% oxygen as a function of the combustion inefficiency for RS2.

Fig. 13 Combustion inefficiency as a function of equivalence ratio for RS5 and RS6 at 600K.

Fig. 14 NOx emissions as a function of the adiabatic flame temperature for RS5 and RS6 at 600K.

Fig. 15 NOx emissions corrected to 15% oxygen as a function of the combustion inefficiency for RS5 and RS6 at 600K. RS5 AND RS6 RADIAL SWIRLER EMISSIONS AT 600K and 740K The combustion inefficiency and NOx emissions results for RS5 and RS6 are compared at 600K in Figs.13-15 and at 740K in Figs. 16-18. The 600K results in Figs. 13-15 show that vane passage fuel injection had significantly lower NOx emissions than for wall fuel injection for both RS5 and RS6 radial swirlers. This was the opposite trend found for the RS1 and RS2 swirlers. As discussed above in relation to the axial wall temperature profiles, Fig. 6 indicates lower temperatures in the outer dump expansion recirculation zone than for wall injection. This slower heat release rate would produce lower residence time at the highest temperatures and thus result in the observed lower NOx emissions. However, there was sufficient residence to time yield a very low combustion inefficiency of 0.1% or below at most test conditions.

Fig. 15 shows that for RS6 with vane passage fuel injection NOx corrected to 15% oxygen was 1.15 ppm at 0.483 equivalence ratio with a combustion inefficiency of 0.05%. For RS5 the equivalent results were 1.6 ppm at 0.525 equivalence ratio. This shows a significant benefit of the circular radial hole RS6 design. This also had a wide equivalence ratio range over which low NOx emissions could be achieved. At 0.57 equivalence ratio (1880K) the NOx was only 2.3 ppm at 15% oxygen. For the swirler outlet throat wall fuel injection the minimum NOx condition for both swirlers was about 2.6 ppm at 15% oxygen at 0.5 equivalence ratio (1770K).

Fig. 16 Combustion inefficiency as a function of equivalence ratio at 740K for RS5 and RS6.

Fig, 17 NOx emissions as a function of the adiabatic flame temperature for RS5 and RS6 at 740K.

Fig. 18 NOx corrected to 15% oxygen as a function of the combustion inefficiency for RS5 and RS6 at 740K. The 740K results in Figs. 16-18 also show significantly lower NOx emissions for passage fuel injection with both the RS5 and RS6 swirlers, as was found at 600K. However, the performance of the two swirlers is almost identical at 740K. The only advantage that RS6 has over RS5 is the slightly leaner primary zone condition at which it can maintain a low combustion inefficiency. At the same flame temperature the two swirlers have identical NOx emissions.

Comparison of the NOX emissions at a primary zone condition of 1700K with RS1 and RS2 in Figs 8 and 11 shows significantly lower NOx emissions with passage injection for RS5 and RS6. These swirlers give NOx emissions of 1.5 ppm compared with 3 ppm and 2.5 ppm for RS1 and RS2. Both of these swirlers had NOx emissions with wall injection that was very close to RS5 and RS6 with passage in jection. The minimum NOx emissions at 15% oxygen that could be achieved with passage fuel injection, with a combustion inefficiency below 0.1%, was 1 ppm for RS6 at 0.41 equivalence ratio with 0.03 combustion inefficiency. For RS5 the equivalent results were 1.3 ppm at 0.45 equivalence ratio and a combustion inefficiency of 0.04%. However, at richer mixtures the NOx performance for RS5 and RS6 was exceptional and at 0.55 equivalence ratio (1980K) the NOx emissions corrected to 15% oxygen were only 4 ppm. This is a fully premixed performance and would allow a 25% power turn down on a main burner with NOx emissions of 4ppm at full power conditions. As RS6 lends itself to a simple variable geometry mode with butterfly valves in each circular radial hole, then power variation by air staging could easily be achieved

over a wide power range, with low NOx and low CO and UHC at all power conditions. For 76mm wall fuel injection the RS5 and RS6 NOx results are not as good as for passage fuel injection and there is no flame stability advantage. Minimum achievable NOx was 2.5 ppm for both swirlers, more than double that for passage fuel injection. This was achieved at 0.43 equivalence ratio, roughly the same primary zone equivalence ratio for the minimum NOx with passage fuel injection. As there was no significant difference in the weak extinction for the two modes of fuel injection, it is unlikely that the NOx differences were due to fuel/air mixing differences. The axial wall temperature profiles indicate that the NOx differences are likely to be due to the slower development of the axial temperature in the flame with passage fuel injection. COMPARISON OF ALL FOUR SWIRLERS All four radial swirlers are compared in Fig. 19 at 600K and in Fig. 20 at 740K for both methods of fuel injection.

Fig. 19 NOx corrected to 15% oxygen as a function of the combustion inefficiency at 600K for all four swirlers.

Fig. 20 NOx corrected to 15% oxygen as a function of the combustion inefficiency at 740K for all 4 swirlers. Figs. 19 and 20 show the superior NOx performance of the circular radial passage RS6 with passage fuel injection. A simple way to rank the swirlers is on the basis of the minimum corrected NOx that can be achieved and the equivalence ratio that this occurs at with the corresponding inefficiency. This is done from the above Figures in Table 3.

Table 3 Minimum NOx ppm corrected to 15% oxygen conditions for combustion inefficiencies below 0.1%. RS 1 1 2 2 5 5 6 6 P W P W P W P W 600K

NOx 1.4

2.3

2.0

1.6

2.5

1.1

2.7

EQR .47

.48

.48

.52

.51

.48

.50

740K

NOx 1.7

1.1

1.5

1.1

1.2

1.7

1.0

1.8

EQR .39

.41

.39

.41

.43

.42

.41

.42

The results in Table 3 show a very similar performance for all four radial swirlers. The leanest equivalence ratio at which the lowest NOx emissions occurred with a combustion inefficiency better than 0.1% was very similar for all the swirlers and both methods of fuel injection and ranged from 0.47 to 0.52 at 600K and 0.39 to 0.43 at 740K. Comparison with the weak extinction equivalence ratio in Table 2 shows that at 600K the range was 0.37 – 0.41 and at 740K 0.29 – 0.34. There is thus, at both inlet temperatures, a clear stability margin. In the worst case this is better than 20% at both inlet temperatures. Modern gas turbine control systems can operate a primary zone well within 20% of the stability limit. Hence, the conditions in Table 1 are practical operating conditions for the primary zone. If power variation was done using variable air flow into the primary zone then these NOx emissions could be achieved through the power range. If power variation was by fuel flow reduction then at maximum power a higher NOx condition would have to be used, as discussed above. The NOx emissions show that the mode of fuel injection influenced the results significantly. Wall fuel injection gave lower NOx emissions than passage injection for RS1 and RS2 and the reverse occurred for RS5 and RS6. The excellent NOx performance with vane passage fuel injection for RS5 and RS6 is considered to be due to the passages being longer and this allowed the flow separation at the inlet to reattach to the walls of the passage before the exit. This was seen in the discharge coefficient results. The circular hole radial passage swirler had the longest holes and the best performance with passage fuel injection. The jet reattachment aerodynamics are considered to produce improved fuel/air mixing in the vane passages. CONCLUSIONS 1 . The results showed that the different techniques used to manufactur e the passages did not have a great influence on the results as all passages with passage fuel injection gave NOx emissions in the 1-2 ppm range for lean combustion at 600K and 740K inlet temperature and atmospheric pressure.

2. There was no advantage in the curved blade passage design of RS1 or the aerodynamic profiled blade of RS2, both in terms of discharge coefficient or NOx emissions. 3. Although the NOx performance of the swirlers was similar there were significant differences. RS1 and RS2 had lower NOx emissions with 76mm swirler outlet wall injection and RS5 and RS6 had lower NOx emissions with swirler vane passage. 4. The best NOx emissions were for RS6 with circular vane holes that were longer in terms of their L/D than for any of the other designs. This allowed flow reattachment within the vanes and this is considered to assist the fuel/air mixing within the vanes and hence produce lower NOx. 5. The circular hole radial vane passages are ideally suited to use with a throttle valve to give a variable geometry air staged combustor.

ACKNOWLEDGEMENTS

We would like to thank the Science and Engineering Research Council for research grants in support of parts of this work (GR/J/07976, GR/J/10969, GR/L/68100). The test facilities were developed and operated by R.A. Boreham. REFERENCES

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