THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS
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Copyright © 1994 by ASME
Printed in U.S.A.
94- GT-277
INSTALLATION OF A TURBINE STAGE IN THE
PYESTOCK ISENTROPIC LIGHT PISTON FACILITY
M. A. Hilditch
DRA Pyestock
Farnborough, UK
A. Fowler and T. V. Jones
Dept. of Engineering Science
Oxford University
K. S. Chana
DRA Pyestock
Farnborough, UK
Oxford, UK
M. L. G. Oldfield, R. W. Ainsworth,
and S. I. Hogg
Dept. of Engineering Science Oxford University
Oxford, UK
ABSTRACT
S. J. Anderson and G. C. Smith
DRA Pyestock
Farnborough, UK
The Isentropic Light Piston Facility (ILPF) at Pyestock has been
upgraded to include a single stage, high pressure turbine. All major
non-dimensional parameters are accurately scaled during the 0.4s run
time, enabling heat transfer and aerodynamic measurements to be
made at engine representative conditions. The ILPF was previously
an annular cascade facility. This paper describes the design and
integration of the rotor module and the results of the commissioning
tests. An important feature is a novel, patented, turbobrake which is
shown to maintain the turbine at a constant speed during the run.
The recent advances in affordable computing power have hastened
progress towards the goal of a three dimensional (3D) unsteady
prediction of heat transfer in a full turbine stage. 3D inviscid steady
flow calculations are now a routine part of the design process and
have resulted in a significant improvement in turbine performance.
There are two approaches to improving the prediction of unsteady
turbine flows, either to account for the unsteadiness by improving the
loss and heat load correlations in steady codes or by developing
accurate models of the unsteady effects (Sharma et al., 1990).
NOMENCLATURE
Many experimental programmes are underway to provide data for
correlations or to aid interpretation of this complex flowfield so that
better models may be generated. Much of the experimental work has
been carried out in short duration facilities as these enable engine
conditions to be correctly modelled at a fraction of the cost of
continuous facilities and simplify the instrumentation needed for heat
transfer measurements. A recent comprehensive review of these
transient facilities is given in Jones et al. (1993). One type of
transient facility is the Isentropic Light Piston Tunnel which was
devised at Oxford University specifically for turbine testing (Jones et
al., 1973). A number of these facilities have been built in several
countries. An ILPT at Oxford University has been used for several
years to test a full turbine stage (Ainsworth et al., 1988) and a large
annular facility has recently been commissioned at the Von Karman
Institute, Brussels (Sieverding and Arts, 1992). Another such facility
is the Isentropic Light Piston Facility (ILPF) at DRA Pyestock
(Brooks et al., 1985) which has recently been upgraded to incorporate
a full rotating turbine stage.
True chord
C
CpSpecific heat
Pressure
p
Mass flow
m
Rotational speed
N
Temperature
T
Velocity
U
p
Density
µ
Viscosity
Rotational speed
w
Subscript
Inlet conditions
1
NOV exit conditions
2
Rotor exit conditions
3
Stagnation conditions
0
INTRODUCTION
The high operating temperature of a modem gas turbine dictates that
both stator and rotor blades in the first turbine stage must be cooled.
Excess cooling lowers the overall efficiency of the engine whilst too
little is detrimental to blade life. Code developers are still working
towards an accurate prediction of turbine heat transfer which is of
quantitative benefit to the designers. This is a considerable challenge
as a 15K increase in blade metal temperature has been estimated to
reduce blade life by a factor of two (Hennecke, 1984).
Work in the Isentropic Light Piston Facility (ILPF) at Pyestock has
concentrated on the influence of 3D flow on turbine heat transfer
(Harvey and Jones, 1990, Chana, 1992 and Harasgama et al., 1992).
In general it has been found that the secondary flow reduces the root
and tip heat transfer on the vane suction side, but leads to high heat
transfer near the pressure surface trailing edge. Also the horseshoe
vortex entrains the endwall boundary layer causing the growth of a
new thin boundary layer and a consequent region of high heat
transfer.
Presented at the International Gas Turbine and Aeroengine Congress and Exposition
The Hague, Netherlands — June 13-16, 1994
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A rotor blade is subject to the same secondary flow mechanisms
within the passage, but the flow is further complicated by rotation
which introduces overtip flows, centrifugal effects and interactions
with both upstream and downstream blade rows. The effects of wake
passing and shock interaction on aerofoil heat transfer have been
studied in a linear cascade by Doorly et al. (1985) and Rigby et al.
(1990) amongst others. It was found that wake impingement caused
a periodic tripping of the boundary layer from a laminar to turbulent
state and a consequent rise in the mean heat transfer rate. Shock
wave passing caused a more rapid change in heat transfer rate which
has been linked to a transient vortex bubble shed from the leading
edge (Johnson et al., 1989).
There are several transient facilities in which heat transfer
measurements have been made on rotor blades at engine
representative conditions (Guenette et al. 1989, Dunn et al. 1989 and
Hilditch and Ainsworth, 1990). In each case large fluctuations in heat
transfer rate were seen at blade passing frequencies at certain
locations on the blade surface. Comparisons of rotor data with cascade
data and state of the art predictions show similarities in regions where
the flow is 2D, but also reveal the complex nature of the rotor flow.
The current experiment in the Pyestock ILPF is aimed at a better
understanding of aerodynamic loss due to film cooling and the
provision of data sets for code validation.
H P Reservoir
Working
(\ o ;:©^
Fast Acthe Vdve
Pump Tube
Figure 1: Isentropic Light Piston Facility
THE FACILITY
The isentropic light piston facility has been described in detail by
Brooks et al., 1985. The addition of a rotor has involved changes to
the working section but not to the operation of the tunnel. The main
components, Figure 1, are a large diameter pump tube connected
through a fast acting valve to the working section of the rotor module.
Prior to a run the piston is moved to the far end of the pump tube and
the volume between the piston and the fast acting valve is filled to a
predetermined pressure. The working section is evacuated and the
rotor spun to its design speed. To start a run a flow of high pressure
air is admitted behind the piston forcing it down the tube compressing
isentropically the air in front. When the correct pressure and hence
temperature are reached the fast acting valve opens allowing a steady
stream of air to pass over the turbine for approximately 400 ms.
Figure 2: Rotor module
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The rotating assembly (Figure 2) is supported by six spokes and
suitable adjustment is incorporated to allow concentricity to be
achieved between rotating and stationary components. The complete
assembly may be removed sideways and the axis rotated to the
vertical to facilitate fitting and adjustment.
The rotating system was designed as a separate module allowing it
to be built whilst the final series of annular cascade tests were being
undertaken. Details of the rotor module are shown in Figure 2. A
novel feature of this design is the patented, aerodynamic turbobrake,
described later in the paper, which absorbs the energy produced by
the turbine so that the turbine speed is held constant throughout the
run.
The facility has been designed so that all non-dimensional groups
important to turbine fluid mechanics and heat transfer are closely
matched. The first series of measurements will be carried out at a
single operating point and temperature ratio. The expected operating
point has been calculated using a throughflow program and is listed
in Table 1. Details of the turbine geometry and running conditions
corresponding to the design operating point are given in Table 2.
Also included in Table 2 are the upper and lower operating limits of
this turbine facility.
The turbine disc and turbobrake are mounted at either end of the
shaft. The shaft is supported on two sets of bearings which are
lubricated using an oil flow system. The bearing assembly is similar
to that used in the transient turbine facility at Oxford University
(Ainsworth et al., 1988) and consists of a matched pair and a single
row of deep groove, annular contact ball bearings. An axial pre-load
is applied to the single row and the rotor thrust during the run is
taken by the matched pair. The shaft is hollow to enable electronic
circuit boards to be mounted in the shaft. These condition the signals
from instrumentation mounted on the rotor blades before transmission
through a slip-ring.
Table 1: Turbine operating point
The NGVs have previously been tested as an annular cascade
(Harvey and Jones, 1990) and the inlet contraction and vanes from
those tests are being reused. Twelve of the vanes are mounted in
removable cassettes allowing NOV instrumentation to be readily
changed. The rotor blades can be viewed by removing a cassette and
access to the rotor area of a sector of four blades can be gained by
unscrewing a small plate.
p2UAC
Reynolds Number
2.6 E 6
µ2
N
Tot
Specific speed
The rotor blades are manufactured with a lip on the front of the root
and a pressed aluminium tab is located in a groove on the underside
of the root fixing. The blade is slid in from the front and held axially
by bending the tab. This system enables individual blades to be
changed without requiring access to the whole face of the disc. Trim
balancing of the rotor is carried out by changing the appropriate
aluminium tabs for heavier ones, usually made of brass.
m Tol
Mass flow number
Downstream of the rotor the flow passes through a sudden
expansion and then enters the second throat. This variable area
device is adjusted to set the blade exit pressure and hence the stage
pressure ratio. A pressure ratio of greater than 2:1 across the throat
ensures that it is choked, and isolates the turbine from disturbances
emanating from downstream. The throat is fully annular to avoid
periodic disturbances to the rotor exit flow. A traverse mechanism for
rotor exit flow surveys is mounted in a cassette just upstream of the
second throat.
439.2 rpm K"o.s
0.792 E 3 ms K"o.s
Poi
Pot
Pressure ratio
3.2
P3
Table 2: Conditions at design and facility operating range
Parameter
At design
Operating range
Inlet total pressure
NGV exit Mach number (isentropic)
Steady rotational speed
Total temperature
Temperature ratio (NGV)
Rotor relative temperature ratio
Run time
Rotor blade mid-height radius
Rotor blade axial chord
4.6 bar
0.94
9500 rpm
467 K
1.624
1.420
400 ms
0.275 m
27.1 mm
10 bar
4000
293
1.0
0.9
300
-
-
-
10000 rpm
530 K
1.8
1.6
1700 ms
-
3
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CONTROL SYSTEM
The control system for the rotor module operates the air motor drive
system and monitors the mechanical performance of the rotating
CS8A
SG
COinEONst
AWUFr-AigN
assembly. It was designed to be separate from the system used to
control the ILPF so that the rotor module could be built and
commissioned independently.
1 GB HD/64 MB RAM
TAPE DRNE 150 MB
16
W-SKAFT ELEC1RONCSWGF15PEm
r^
ROTOR BLADE
HEAT TRANSFER GAUGES
aaESVir^ laA►aouc^as
P RING Qa CNMINELq
i
it
zpC N
16 BR
1 MHZ/CHAN
-
H
2 CPU 2 USER
UNUC
Figure 4: ILPF data acquisition system
Total pressure is measured far upstream to check for circumferential
variations and also at NOV inlet. A traversable total pressure probe
A modem TI 405 PLC (Texas Instruments) was chosen to run the
rotor module. This is far more sophisticated than the 5 TI system and
can be programmed so that several activities occur in parallel. The
block diagram shown in Figure 3 illustrates how control passes from
one stage to another. The stages which monitor switch positions, oil
and bearing parameters, speed and abort status are active throughout
the run. If an unsafe condition is detected the normal flow of the
program is halted and the drive system shut down.
will be used to check for variations across the span. The pressure
signals are recorded using a Scanivalve HyScan 2000 system which
is set to take data on 128 channels with an interval of 15ps between
channels. The ZOC 14 electronic pressure scanning modules are
multiplexed to a Digiquartz TX transducer (model No. 2200-AS-010)
which is calibrated before each run and measures to an accuracy of
0.1%. Some pressure tappings are also measured with National
Semiconductor pressure transducers (type LX 1620D) and recorded on
the slow speed A/D with a 95% confidence limit of 0.1%.
^ I n niw
-
14 7DDD
H
Bx1620^W
The operation of the facility is controlled through a 5 TI
programmable logic controller (PLC) made by Texas Instruments.
This carries out operations requested by the operator (eg opening
valves) unless the requested sequence of events is unsafe. This
system has been in place since the ILPF was commissioned as an
annular cascade facility and only the minimum changes have been
made to incorporate the rotor module.
!ti'
MASSCOMP
128 gUN5
Heat transfer measurements on the NGV are made using vanes
manufactured from machinable glass ceramic on which platinum thin
film heat transfer gauges are painted (Oldfield et al., 1978). A total
of 165 gauges have been painted on four vanes at five radial heights
and on the inner and outer platforms. The signals are passed through
an electrical analogue which exactly models the one-dimensional heat
transfer equation giving an output directly proportional to heat transfer
rate (Oldfield et al., 1984).
4 Set-up I ! Start-up Spin up I Finng^—^I I--^I
-
Oil Bad
beng
motor
Figure 3: Block diagram of rotor PLC program
The routine measurements made on the rotating assembly are rotor
and air motor speed, oil inlet and outlet temperatures, bearing metal
temperatures, bearing cage speeds, air motor driving pressure and
vibration of the bearing housing. Most of these are monitored by the
PLC and in addition the parameters which indicate the mechanical
performance of the system are recorded so that any change with time
can be analysed retrospectively. This dedicated data acquisition
system is PC based and uses the Labtech Notebook software package.
AERODYNAMIC AND HEAT TRANSFER MEASUREMENTS
The current experiment in the facility is aimed at providing detailed
heat transfer and aerodynamic measurements of a full turbine stage.
The instrumentation and data acquisition systems used for annular
cascade tests in the facility have been extended to incorporate
measurements on the rotor blade. The present data acquisition system
is shown in Figure 4. Static pressure measurements will be made
around the NGV aerofoil at five radial heights and on the inner and
outer platforms. Upstream of the NGVs and downstream of the rotor
there are static pressure tappings at several circumferential locations.
The technique of painting thin film gauges directly on to a ceramic
substrate could not be used in the rotating frame because of the high
stress levels involved. An alternative technique of coating a metal
blade with a layer of vitreous enamel on which the thin film gauges
are then mounted, has been developed at Oxford University (Doorly
and Oldfield, 1986 and Ainsworth et al., 1989a) and proved in their
turbine facility (Hilditch and Ainsworth, 1990). Six blades are being
instrumented at three radial heights. One of the blades to be used in
the ILPF is shown in Figure 5. In this case the gauges have been
manufactured by depositing a layer of gold and then photoetching
gauges and leads of the required shape.
Electronic circuit boards (Ainsworth et al., 1989a) are mounted
inside the rotor shaft to preferentially amplify the signals before
transmission through a slipring. The heat transfer signals are then
conditioned using fast response AMP-05 amplifiers and finally
recorded on either the low or high speed A/D. The electrical
analogue is not used because it does not model exactly the response
of a thin film gauge on a two-layered substrate. Instead the signals
have to be digitally processed to recover the heat transfer rate
(Doorly, 1987).
Unsteady pressure measurements using miniature high response
pressure transducers mounted in pockets on the blade surface have
been pioneered at Oxford University (Ainsworth et al., 1989b). This
technique will be used in the ILPF and four blades are under
manufacture with a total of 20 transducers.
2
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pressure rise through the turbobrake. The turbobrake blade profile
(Figure 6) is somewhat unusual, resembling that used in a transonic
impulse turbine but rotating in the opposite direction! It was designed
to maximise added whirl, not for efficiency and is a constant section
for ease of manufacture.
Earlier tests (Goodisman et al., 1992) using a 17% scale model
running at the correct Reynolds number and dimensionless speed
showed that the concept worked, and that the turbobrake stage loading
of — 2 was more than enough to absorb the ILPF turbine power. It
was also demonstrated that the turbobrake power could be adjusted by
a combination of bypass and downstream blockage. Figure 2 shows
the by-pass shutters and downstream blockage rings used on the full
size brake. Since the turbobrake flow turns on and off with the test
turbine flow during a tunnel run, there is no need for any rapidly
acting (possibly failure prone) control system and this feature
considerable enhances the safety of the turbobrake system. The
turbobrake tracks the turbine and automatically keeps the speed under
control.
Figure 5: Rotor blade heat transfer gauges
Figure 7 shows the full size (690.2 mm diameter) turbobrake blisk,
with disc and integral blades machined from forged, heat treated
HDA81 aluminium alloy. This was spin-tested to 12,000 rpm (25%
overspeed) to confirm the structural integrity of the design.
Containment rings fabricated from high-yield, stainless-steel sheet,
surround the turbine and turbobrake discs during the tunnel run. This
is designed to absorb the kinetic energy associated with an overspeed
disc failure.
Other instrumentation on the facility will include measurement of
unsteady rotor relative total pressure, inlet traverses with a hotwire,
thermocouple and pitot tube and traverses downstream of the rotor for
total pressure, flow angle and total temperature.
TURBOBRAKE
There are advantages in being able to keep the turbine rotational
speed constant during the 400 ms tunnel run, although it is not
necessary for a successful experiment (Ainsworth et al., 1988). The
use of some type of brake allows longer run times without
overspeeding the rotor, both time and ensemble averaging are more
accurate and extensive probe traversing may be performed during the
run. The advantage of not employing matched braking is that a range
of off design conditions may be produced during one run. The
braked system is, of course, more complex than the unbraked rotor.
Brakeless, inertial containment has been successfully used
(Ainsworth et al., 1988), but the rotor speed then increases during the
run and for this facility a massive flywheel would be required to keep
the increase in check. Goodisman, et al. (1992) surveyed the braking
systems (dynamometers) then available and concluded that the most
suitable was the new, patented, axial turbobrake.
The principles of the axial turbobrake have been previously
described (Goodisman et al., 1992; Goodisman, 1991) and are only
outlined here. The turbobrake consists of a second disc mounted on
the turbine shaft (Figure 2), with blading in the turbine exit flow.
The volume upstream of the turbobrake is aerodynamically isolated
from the turbine exit by the choked second throat. The turbobrake
absorbs work from the turbine shaft by adding considerable whirl to
the exit flow and this energy is dissipated in the low pressure dump
tank downstream. Unlike a conventional compressor there is no static
29 Blades
Tip Radius = 345.1 nun
Hub Radius = 246.7 mm
Axial Chord = R5_Qmm
x2,^
roc
^l1 =0.45
a o ,=r (yR7
xt e. = by
0
L.E.
T.E.
Figure 6: Mid-height section of axial turbobrake blades,
with velocity triangles (from Goodisman et al., 1992)
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This has a simple exponential solution
(Co — Cod) = ((a t — (ad)e
_e
where w, is the initial speed, and c = I/B is the characteristic time
constant of the system.
For the design, hot case (9500 rpm) this gives a time constant ti =
1.4 sec. Thus any initial speed mismatch will be reduced by only
25% during the 400 ms run. For a cold run at the same Reynolds
number and dimensionless speed the equilibrium speed is reduced to
7600 rpm and t = 1.9 sec. As this is comparable to the - 2 second
cold run time, the speed will approach the equilibrium value during
a cold run.
COMMISSIONING
The rotating system was built up systematically. The drive system
and bearing assembly were proved by spinning the shaft alone up to
the design speed. Then the turbobrake was added and spun
successfully before the rotor disc was installed. The power required
to spin the assembly in bare shaft configuration and with both discs
present is shown in Figure 9. A considerable amount of power (50
kW) is required to spin the shaft alone and this is attributed to the
high pre-load needed to prevent the bearings from skidding. The
windage on the turbobrake without through flow is proportional to the
Figure 7: Photograph of the turbobrake blisk
Although, when the brake is correctly adjusted, the turbine speed is
held constant, it is useful to be able to predict the rate of change of
speed of the system when out of adjustment. Figure 8 shows
predictions of turbine and turbobrake torque:speed curves at design
hot conditions, for a design speed of 9500 rpm. They were calculated
by using the simple two-dimensional lossless method given in
Goodisman (1991). They suggest that, theoretically, the turbobrake
should be adjusted to absorb 52% of the maximum power it can
absorb. The model tests indicate that losses already reduce the
theoretical power by 12%, and so a 59% setting is appropriate in
80
practice.
70
60
3 50
(Brake max
......
....
V
40
30
Turbine
z
Q
..
2000
_
...
20
....
10
Brake matched for 9500 rpm
•
0
1000
0
Excess Torque
100 0
2000
4000
6000
8000
Rotor speed (rpm)
1
2
3
4
5
Thoasandt
6
7
8
9
10
Shaft Speed (RPM)
Both discs • Both discs , Both discs . Shaft only
10000
12000
Figure 9: Power required to spin rotor assembly at steady
speed
Figure 8: Predicted turbine and turbobrake torque/speed
curves. Design speed = 9500 rpm.
density of the surrounding fluid and to the cube of the speed. The
power produced by the air motor is limited to 70 kW by the available
air supply and the rating of the drive-belt. In order to reach design
speed within this limit a vacuum level of lower than 20 milli-bar (mb)
is required. A typical speed-time plot is shown in Figure 10 and the
temperature of the oil and middle bearing during this run is shown in
Figure 11. It takes approximately three minutes to spin the turbine
assembly to design speed and two minutes to stop in a good vacuum.
After a tunnel firing the turbine will decelerate more quickly because
of the increased pressure in the working section.
It is clear that the excess torque TT ( = turbine torque - turbobrake
torque) can be modelled as a linear function of speed T = B(co d - co)
where wd is the design speed at which TT = 0 and B is the slope of
the excess torque:speed curve. The dynamic equation for the shaft
then simplifies to
Iw = TQ = B(w d - c^)
where 1(3.3 kgm 2) is the moment of inertia of the rotating assembly.
6
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temperature ratio commissioning run is shown in Figure 12. The
pressures in front of and behind the piston were well matched
resulting in steady mean pressures upstream of the second throat. The
small oscillations seen on all traces are a well documented
phenomenon (Jones et al., 1993) caused by oscillations of the piston
in the compression tube. This is taken into account during data
processing and has negligible effect on the high frequency phenonema
encountered in the turbine.
disc160
10
9
0
0
8
6-
s
4
RUN 1436
3
2
mr
1
00
100
200
300
400
500
A(
pressure
In
Time (secs)
w u.
Figure 10: Speed during rotor assembly commissioning
run
N
disc160
I
--- Oil outlet temp
- - Bearing 2 temp
1
U
a
m
6
E
F
Kit hub
Kit casing
L
^ O
iroat in
100
200
300
Time (secs)
400
500
iroat out
600
Figure 11: Temperature readings during rotor assembly
commissioning run
To preserve the balance of the rotor assembly each blade was
weighed and a moment weighing programme used to calculate how
they should be distributed to give the minimum out of balance force.
The vibration level with and without blades was less than 5 mm/s at
both the front and rear of the bearing housing. Published literature
and experience with the Oxford rotor (Sheard et al., 1992) suggest
that vibration levels of more than 4 mm/s are unacceptable. In-situ
balancing was carried out using a Schenck Vibroport 30 machine
which was set to perform tracking measurements. Traces of vibration
level and phase angle with respect to a once per revolution reference
signal were recorded whilst the rotor was spinning down. Two sets
of data were recorded - the datum level and a second set with a trial
weight in place. From the differences between the two sets of traces
the position and magnitude of the weight required to balance the rotor
was calculated. This procedure was successful in reducing the
maximum vibrations to less than 3.5 mm/s.
The first runs of the enhanced facility in which air was passed
through the turbine occurred in September 1993. The initial tests
were conducted with 0.5 b pressure in the pump tube and no
compression. The tube and reservoir pressures were steadily
increased until design Reynolds Number and a temperature ratio of
1:1.1 were reached. Running at a lower inlet temperature enabled the
turbine to be tuned to the correct operating point whilst running at a
lower rotational speed of 7800 rpm. A number of runs were carried
out at this condition before the second throat and turbobrake were set
to the desired values. Aerodynamic data recorded during a 1.1
°0.00 0.20 0.40 0.60 0.80 1.00 1.20 1.40 1.60 1.80 2.00
time (sec)
Figure 12: Conditions during 1.1 temperature ratio
commissioning run
Initially the turbobrake was set at maximum power and as predicted
the speed decayed rapidly during the run, Figure 13. The downstream
throat was then set to give the correct turbine pressure ratio at the
start of the run The turbobrake bypass shutters were opened in
stages, and the speed decay reduced. As predicted by Goodisman et
al. (1992), it proved necessary to install a 12.5% downstream
blockage ring together with some bypass flow to match the turbobrake
to the turbine. The adjustment of the bypass to reduce the speed
change to less than 1% during the run proved to be simple and
progressive. Once the matching point had been found, the tunnel was
then run at progressively hotter temperatures until the design
conditions (Tables 1 & 2) were reached. The non-dimensional
performance of the turbobrake scales in a similar manner to that of
the turbine, so the turbobrake settings did not need further changing,
,
as the reduced operating speed N//Tot had been kept constant.
The speed histories of two runs at design conditions with the
turbobrake correctly set are shown in Figure 14. The excellent
performance of the turbobrake is illustrated by run #1464 when the
7
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CONCLUSIONS
8000
A single stage, high pressure turbine has been successfully
integrated with the Pyestock isentropic light piston facility.
Commissioning tests are almost complete and the collection of
aerodynamic and heat transfer data at full engine representative
conditions will begin in January 1994.
7500
7000
The choice of the previously untried axial turbobrake concept has
been vindicated. The full size turbobrake performed as predicted by
the early 17% scale model tests, and proved simple and safe to use
and adjust.
6500
6000
\431
5500
ACKNOWLEDGEMENTS
5000
45000
1
Time (s)
0.5
The authors would like to thank Mr. K.J. Walton for his help in
maintaining the test facility and installing the rotor module. Thanks
are also due to the workshop staff at the Oxford University
Engineering Department for their manufacturing expertise and to
Foxley Design Associates Ltd.
2
1.5
Figure 13: Speed during cold commissioning run turbobrake absorbing maximum power
9600
REFERENCES
Ainsworth, R.W., Schultz, D.L., Davies, M.R.D., Forth, C.J.P.,
Hilditch, M.A., Oldfield, M.L.G. and Sheard, A.G., 1988, "A transient
flow facility for the study of thermofluid-dynamics under engine
representative conditions", ASME paper 88-GT-144
Run#1463
9550
9500
Ainsworth, R.W., Allen, J.L., Davies, M.R.D., Doorly, J.E., Forth,
C.J.P, Hilditch, M.A., Oldfield, M.L.G. and Sheard, A.G., 1989a,
"Developments in heat transfer and processing for transient heat
transfer measurement in a full stage model turbine", ASME Journal
of Turbomachinery, Vol. 111, pp. 20-27
Run#ta6a
9450
9400
93500
0.1
0.2
0.3
Jm
0.4
0.5
0.6
Ainsworth, R.W., Allen, J.L. and Dietz, A.J., 1989b, "Methods for
making unsteady aerodynamic pressure measurements in a rotating
turbine stage", AGARD CP-468
0.
Brooks, A.J., Colbourne, D.E., Wedlake, T.E., Jones, T.V., Oldfield,
M.L.G., Schultz, D.L. and Loftus, P.J., 1985, "The isentropic light
piston cascade at RAE Pyestock", AGARD-CP-390
Figure 14: Runs at design temperature ratio with
turbobrake correctly set
speed was maintained to within 5 rpm of the design value. (The
fluctuations are due to uncertainties in the time discretization of the
speed sensor square wave signal). Also shown on Figure 14 is a +/0.5% error bar which shows that run #1463, where the initial speed
was slightly too high and the turbobrake acted to slow the rotor, was
close to the acceptable limit. The deceleration seen at the start of
each trace occurs because the supply to the air motor is cut
immediately prior to a run and the starting transient, during which the
rotor accelerates by approximately 50 rpm, is a consequence of the
time it takes for the volume between the turbine and turbobrake to
fill.
Once the full design conditions have been reached static pressure
surveys and measurements of heat transfer distribution will be made
on the NGV. This will give further confidence in the performance of
the enhanced facility before the instrumented rotor blades are tested.
The first data from rotor mounted heat transfer blades is expected in
early 1994 with rotor mounted pressure transducers being operational
soon after.
Chana, K.S., 1992, "Heat transfer and aerodynamics of a 3D design
nozzle guide vane tested in the Pyestock isentropic light piston
facility", AGARD CP-527
Doorly, D.J., Oldfield, M.L.G. and Scrivener, C.T.J., 1985, "Wake
passing in a turbine rotor cascade", AGARD CP-390
Doorly, J.E. and Oldfield, M.L.G., 1986, "New heat transfer gauges
for use on multi-layered substrates", ASME paper 86-GT-96
Doorly, J.E., 1987, "Procedures for determining surface heat flux
using thin film gauges on a coated metal model in a transient test
facility", ASME paper 87-GT-95
Dunn, M.G., Seymour, P.J., Woodward, S.H., George, W.K. and
Chupp, R.E., 1989, "Phase resolved heat flux measurements on the
blade of a full scale rotating turbine", ASME Journal of
Turbomachinery, Vol. 111, pp. 8-19
8
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Goodisman, M.I., Oldfield, M.L.G., Kingcombe, R.C., Jones, T.V.,
Ainsworth, R.W. and Brooks, A.J., 1992, "An axial turbobrake",
ASME Journal of Turbomachinery, Vol. 114, pp. 419-425
Goodisman, M.I., 1991, "Turbine turbobrake systems", D. Phil. thesis,
Oxford University, OUEL report 1931/92
Guenette, G.R., Epstein, A.H., Giles, M.B., Haines, R. and Norton,
R.J.G., 1989, "Fully scaled transonic turbine rotor heat transfer
measurements", ASME Journal of Turbomachinery, Vol. 111, pp.1-7
Harasgama, S.P., Burton, C.D. and Chana, K.S., 1991, "Measurements
and computations of heat transfer and film cooling in turbines",
ISABE, Nottingham, UK
Harvey, N.W. and Jones, T.V., 1990, "Measurement and calculation
of endwall heat transfer and aerodynamics on a nozzle guide vane in
an annular cascade", ASME paper 90-GT-301
Hennecke, D.K., 1984, "Heat transfer problems in aero-engines",
Metzger, D.E. and Afgan, N.H., eds. Heat and Mass Transfer in
Rotating Machinery, Hemisphere Publ. Corp., N.Y.
Hilditch, M.A. and Ainsworth, R.W., 1990, "Unsteady heat transfer
measurements on a rotating gas turbine blade", ASME paper 90-GT175
Johnson, A.B., Rigby, M.J., Oldfield, M.L.G., Ainsworth, R.W. and
Oliver, M.J., 1989, "Surface heat transfer fluctuations on a turbine
rotor blade due to upstream shock wave passing", ASME Journal of
Turbomachinery, Vol 111, pp. 105-115
Jones, T.V., Schultz, D.L. and Hendley, A.D., 1973, "On the flow in
an isentropic light piston tunnel", MoD (Proc Exec), Aeronautical
Research Council, R. & M. No. 3731
Jones, T.V., Oldfield, M.L.G., Ainsworth, R.W. and Arts. T, 1993,
"Transient cascade testing", Advanced Methods for Cascade Testing,
ed. C Hirsch, AGARDograph 328
Oldfield, M.L.G., Burd, H.L. and Doe, N.G., 1984, "Design of wide
bandwidth analogue circuits for heat transfer in transient tunnels",
Metzger, D.E. and Afgan, N.H. eds. Heat and Mass Transfer in
Rotating Machinery, Hemisphere Publishing Corp., N.Y., 1984
Oldfield, M.L.G., Jones, T.V. and Schultz, D.L., 1978, "On-line
computer for transient turbine cascade instrumentation", IEEE
Transactions on Aerospace and Electronic Systems, Vol AES-14, No.
5, pp. 738-749
Rigby, M.J., Johnson, A.B. and Oldfield, M.L.G., 1990, "Gas turbine
rotor blade film cooling with and without simulated NGV shock
waves and wakes", ASME paper 90-GT-78
Sharma, O.P., G.F. Pickett and Ni, R.H., 1990, "Assessment of
unsteady flows in turbines", ASME paper 90-GT-150
Sheard, A.G., Dietz, A.J, and Ainsworth, R.W., 1992, "The dynamic
characteristics of a high pressure turbine stage in a transient wind
tunnel", ASME paper 92-GT-166
Sieverding, C.H. and Arts, T, 1992, "The VKI compression tube
annular cascade facility CT3", ASME paper 92-GT-336
9
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