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27
Repair
sensitivity study for compressor inlet labyrinth seal using Computational
Fluid Dynamics
Allan Thomson
& David Anderton, APM Team, Wood Group Light Industrial Turbine
Ltd.
This
study examines the potential to relax the repair limits for an industrial
gas turbine compressor inlet labyrinth seal through numerical analysis,
in order to minimize unnecessary, costly and timely rework. The
study demonstrates that for steady-state running conditions there
is considerable potential to relax repair limits without sacrificing
seal performance. It recommends that these results should be discussed
for implementation and highlights areas for further work.
During
the overhaul of a 6.4MW industrial gas turbine, Wood Group Light
Industrial Turbines Ltd (WGLIT) adopt a standard practice of restoring
compressor labyrinth seal fins and abradable liners as new design
clearances. Repair of the rotating knife-edge seals requires the
compressor rotor to be disassembled which itself incurs considerable
further rework and repair cycle time that would otherwise not necessarily
be required. It therefore follows that if the relationship between
seal clearance, bearing chamber sealing efficiency and compressor
delivery pressure losses could be quantified, then the setting of
acceptable overhaul limits for the labyrinth seal clearances would
result in reduced overhaul cost and turn times whilst delivering
the required quality and performance for the intended service duration.
Objectives
The
objectives of this study were twofold:
1. To quantify the effects
of compressor inlet bearing abradable labyrinth seal clearances
on;
- Bearing chamber to compressor annulus
sealing
- Compressor delivery pressure
2.
To examine acceptable overhaul limits for the above mentioned labyrinth
seals.
Methodology
Computational
Fluid Dynamics: Analysis with Air Only One eighth (45 degrees) of
the complete labyrinth seal was modeled with axi-symmetric boundaries,
Figure 1. Four models were built comprising 755,000 hexahedral computational
cells. The labyrinth seal was modeled as a solid wall and the rotating
fins did not cut into the seal. The oil drain in the seal was neglected.
The rotor speed was taken to be 11,000 rev/min. The clearance between
the fin the seal varied from 0.05 mm to 0.250 mm in steps of 0.050
mm.
The flow was assumed to be steady state, compressible
and turbulent. Turbulence was modeled using the two equation k -
ε turbulence model with hybrid wall functions. This approach was
chosen as the value of non-dimensional wall distance, Y+, can range
from about 1 to 250 and the wall treatment function method would
still be valid. Second order differencing was used for momentum
to ensure accuracy and first order differencing on turbulence and
enthalpy.
he pressure drop across the seal was unknown therefore
two pressure drops were chosen to be representative: 2.28 bar –
no throttling had taken place after the sealing air had been bled
off from the 8th compressor stage and 1 bar – some degree
of throttling had taken place.
At a later stage one of the
seal fins was removed, the downstream fin closest to the inlet pipe,
and an analysis was carried out at clearances of 0.05 mm, 0.10 mm
and 0.15 mm. A run was also completed, assuming the fin had cut
into the abradable seal with a clearance of 0.05 mm between the
fin and the non-abradable surface.
Extended
Model (droplets - two phase flow):
The
initial model was extended (1,038,000 computational cells) to include
the chamber with the inlet journal bearing, Figure 2.
Morgan
(1) proposed that an efficient way to simulate oil flow through
an inlet labyrinth seal was to use a transient Lagrangian droplet
model coupled with a liquid film analysis. Similar types of analyses
have been carried out on crankshaft bearings in internal combustion
engines. Due to time constraints and computing resource it was decided
not to include the liquid film analysis, as the end result would
be still be relatively correct. The rotational speed was kept at
11,000 rev/min and the time step 0.5 degrees.
Inlet
Bearing Leakage:
A generic spreadsheet
was adapted to determine the static load on inlet bearing and since
the geometry and oil type (Shell Turbo Oils Type T46) were known,
an analysis was carried out by Harrison (2) to determine the oil
leakage from the bearing. He found it was 0.4 m3/s for the bearing.
In order that a fine mist of oil was produced 80 injection points
were defined, and the particle diameter was chosen to be 0.05 mm.
The injection rate was then found to be 4.33 x 106 droplets/s at
a temperature of 50°C. Using the above calculations at a maximum
bearing clearance of 0.124 mm, the velocity of the oil droplets
was found to be 13.5 m/s.
Results
and Discussion:
A plot of air
leakage against clearance is shown in Figure 3. It can be seen that
with all the seal fins present, a worst case pressure drop of 2.18
bar across the seal and a clearance of 0.25 mm, the leakage across
the seal was just over 0.07 kg/s (less than 0.5% of the total air
mass flow rate through the engine), compared to 0.014 kg/s when
the clearance was 0.05 mm.
At 1 bar pressure drop across the
seal, at a clearance of 0.25 mm the airflow was just over 0.04 kg/s
compared to 0.008 kg/s when the clearance was 0.05 mm. The
peak Mach number across the seal fins was approximately 1 when the
pressure drop was 2.18 bar, i.e. the flow was almost choked at all
seal clearances. At 1 bar the peak Mach number was approximately
0.7. When one fin was removed (first fin downstream of the inlet
pipe in the direction of the compressor exit) then the increase
in airflow across the seal remained roughly constant at about
7% at both pressure drops, at each fin clearance, Figure 3.
Comparison
of the straight through loss at a fin clearance of 0.05 mm and the
loss incurred when the fins had cut into the seal shows that a reduction
in loss of about 5% at both pressure drops when the fins had cut
into the seal, Figure 3. Due to time constraints and computer limitations
only two models were run with oil injection from the inlet bearing:
1.
At a clearance of 0.25 mm and a pressure drop of 2.18 Bar.
2.
At a clearance of 0.25 mm and the inlet flow pipe blocked.
Two
phase Lagrangian modeling can only be simulated when the second
fluid phase occupies less than 40% by volume of the computational
cell. If this was exceeded the model would become unstable, and
fail. Due to the positioning of the oil injection points, the flow
boundaries and its nature, only about 4 revolutions of the engine
were possible before the instability mentioned above occurred. However
this was enough time to show that in model 1 (Figure 5) no oil passed
across the seal even though oil had passed the lip on the rotor.
Model 2 (Figure 7) showed oil passing through the seal, which would
very quickly enter the compressor. The models show progression of
oil particles through the seal. Figures 4 and 6 show the centerline
velocities, to the same scale 0 m/s to 60 m/s. In model 1 the maximum
velocity was 60 m/s whilst in model 2 the maximum velocity was found
to be 310 m/s.
Conclusion
It
has been shown that for the clearances considered here (Figure 3)
under steady state conditions there will be negligible reduction
in compressor delivery pressure and flow rate. For the worst case
(0.25mm clearance at a pressure drop of 2.18 Bar) 70 g/s seal air
leaked across the seal (less than 0.5% of the total air mass flow
rate). For the largest clearance considered there was sufficient
mass flow across the seal to prevent oil ingress to the seal and
hence compressor. Oil is most likely to enter the compressor via
the lab seal when the seal airflow rate is at a minimum, i.e. if
the air pipe is blocked.
Recommendations
Based
on the analysis conducted in this study the compressor inlet labyrinth
seal repair tolerance should be reviewed with the intent of implementing
relaxed clearances.
Further
studies
Transient effects, which
were considered to be secondary, such as thermal growth and start
up / shut down effects, were not considered in this study.
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