A Performance Monitoring of Gas for Failure Prevention Turbines

A
THE AMERICAN SOCIETY OF MECHANICAL, ENGINEERS
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Copyright © 1992 by ASME
Performance Monitoring of Gas Turbines
for Failure Prevention
ROBERT E. DUNDAS
Factory Mutual Engineering & Research
Norwood, Massachusetts
DANIEL A. SULLIVAN
Technical Innovations, Inc.
Sandwich, Massachusetts
FRANK ABEGG
Golden Valley Electric Association, Inc.
Fairbanks, Alaska
ABSTRACT
Special Nomenclature
The concept of performance monitoring for
prevention of certain serious failures in gas turbines
is described. The use of compressor mapping as the key
to avoiding surge is developed, and an example is
presented showing how the compressor in a steaminjected gas turbine can be much closer to surge in one
of two nearly-identical operating points on a steaminjection control envelope than the compressor in the
other. The technique of monitoring blade-path
temperature spread in the exhaust of a gas turbine is
then described, and examples of its value in preventing
combustor burnout and turbine blade failures in highfrequency fatigue are given. Next, a concept of
diagnosing internal deterioration by recognizing
patterns of deviation of operating parameters from
baseline data is described, and illustrated for a
single-shaft generator-drive gas turbine. Finally, the
use of a modern computer-controlled data acquisition
system to perform the above monitoring functions in
real time is demonstrated.
CDP = Compressor Discharge Pressure (P 3 )
EGT = Exhaust Gas Temperature (T6)
HR = Heat Rate-(W f /kW in any units)
NOMENCLATURE
c = specific heat at constant pressure-Btu/lb-°R
(kJ/kg- K)
h s = enthalpy of injected steam-Btu/lb (kJ/kg)
H v = Heating value of fuel-Btu/lb (kJ/kg)
N = rotor speed-rpm
P = total absolute pressure-psia (kg/m2)
P R = compressor pressure ratio
Q = generator output at terminals-kW
T = total absolute temperature- °R (°K)
W a inlet airflow-lb/s (kg/s)
Wf = fuel flow-lb/s (kg/s)
W s = water or steam flow-lb/s (kg/s)
d = pressure correction for compressor inlet flow
y = ratio of specific heats
n = efficiency
n L = assumed mechanical and electrical efficiency of
the load
e = temperature correction for compressor inlet flow
Subscripts
c = compressor
R = Standard day (reference) conditions
It = turbine
Numerical subscripts on P, T and c refer to the
stations in Figures 1 and 2.
Superscript
* denotes corrected value.
INTRODUCTION
The concept of performance monitoring as an
essential element of predictive maintenance and failure
prevention for gas turbines is rapidly gaining
acceptance throughout industry. This acceptance has
been accelerated by the adoption of computer-controlled
data-acquisition and control systems. These systems
permit the on-line acquisition and reduction of a very
large amount of performance information, as well as its
manipulation for presentation to an operator in any
manner considered useful.
Performance monitoring has an important role in
preventing certain types of failure; these are:
compressor surge, compressor and turbine blade failure
in high-cycle fatigue, burnout of combustors, and
thrust-bearing wiping. This paper concentrates on the
use of performance monitoring to prevent these very
serious failures, and shows how modern control systems
can be programmed to recognize and diagnose performance
discrepancies that could result in those types of gasturbine failure.
PERFORMANCE MONITORING
Performance monitoring consists of using the supervisory instrumentation installed on a gas turbine to
Presented at the International Gas Turbine and Aeroengine Congress and Exposition
Cologne, Germany June 1-4, 1992
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CALCULATION OF ENGINE AIRFLOW--SINGLE-SHAFT GAS TURBINE
recognize trends that indicate some internal discrepancy. It does not involve elaborate instrumentation to
assess the exact performance of the machine for comparison with manufacturer's guarantees, or for similar
purposes. That type of monitoring requires extensive
traversing instrumentation to measure temperatures and
pressures throughout the flowpath at various axial
stations along the machine, as well as, possibly, a
system to measure the airflow through the machine at
either the inlet or the exhaust. By means of such
instrumentation, the average temperatures and pressures
at each station can be determined. However, such
elaboration is not necessary for trending purposes to
enable the operator to recognize and correct deterioration that might lead to surge or component failure,
such as high-cycle fatigue of blades, or burnout of the
combustor.
All the instrumentation required for performance
monitoring of a gas turbine for failure prevention
purposes is usually in place. Figure 1 shows the
stations of a single-shaft industrial gas turbine,
indicating the instrumentation required for monitoring.
This instrumentation measures static temperature and
static pressure at key locations, and it should be
installed in the casing at axial locations where the
throughflow velocity is lowest--for example, the
instrumentation at Station 3 should be installed
downstream of the diffuser; in this way, the measured
values will approximate the thermodynamic (total)
parameters most closely. Fuel flow (and, if appropriate, injected steam or water flow) should also be
measured.
The following discussion outlines suitable
algorithms for calculation of the airflow, as well as
the turbine inlet temperature for the unit shown
schematically in Figure 1. For simplicity a number of
assumptions are required:
1. The injection of steam does not affect the
thermodynamic properties of the gas flowing through
the turbine significantly.
2. There is a 5 percent pressure drop through the
combustor.
3. The thermodynamic properties of the working fluid
are constant through the various components, and
constant regardless of the operating condition; it
is suggested that the values listed in Table 1 be
used.
Table 1 Assumed Properties of the Working
Fluid for Performance Monitoring
Station
Specific Heat at
Constant Pressure
Btu/(lb-°R)
0.24
0.25
0.28
0.25
Ratio of Specific
Heats
kJ/(kg-°K)
1.005
1.047
1.172
1.047
1.40
1.40
1.33
1.33
(ye)
(yc)
(yt)
(yt)
The equation for airflow Wa is
Stations:
Wa=
T
P
T
2
P
2
T
3
P
3
(
HWf + hsWs - ( Cp6Te) (Wf +WS
)
)
-
KQ /11 z,
(1)
C16 T6 — Cp2 T2
6
where K = 0.948 for Wa in lb/s,
K = 0.239 for Wa in kg/s.
6
and the equation for the turbine inlet temperature T4
is
+HvWf +h s Ws
T = Cp3T3Wa
f +W.)
Cp4
(W.+W
°
Load
a
(2)
CALCULATION OF ENGINE AIRFLOW--AERODERIVATIVE GAS
TURBINES
Compressor
Combustor
Turbine
Figure 1. Instrumentation required for performance
monitoring on a single-shaft industrial gas
turbine.
No instrumentation is shown at the turbine inlet
(Station 4), because the temperature at that location
is such that the durability of the installed
instrumentation would be in question. Instead, the
exhaust gas temperature, measured by the ring of
thermocouples at Station 6, is used to control the
turbine inlet temperature to the desired limit. Some
control and monitoring systems have algorithms in their
software to calculate the turbine inlet temperature, as
well as the airflow through the machine. These
algorithms vary in complexity; this is of little
consequence as long as they are used consistently in a
program of performance monitoring.
The typical aeroderivative is a two-shaft gas
turbine. The following discussion on calculation of
engine airflow, therefore, also applies to two-shaft
heavy-duty gas turbines. It also applies to aeroderivatives having dual-rotor gas generators; the lowpressure and high-pressure compressors, as well as the
low-pressure and high-pressure gas-generator turbines,
are treated as a single unit. It must be noted, however, that this statement does not apply to compressor
mapping, as will be discussed under that heading.
The instrumentation required for an adequate
program of performance monitoring of a two-shaft engine
is shown in Figure 2. As in the case of a single-shaft
machine, the parameters measured by the instrumentation
of the figure are used to calculate values for the
inlet airflow W a and the turbine inlet temperature T4.
The values of Q, W f , and W s are measured. In contrast
to the single-shaft gas turbine, the value of the
efficiency n L of the load need not be assumed if all
the indicated parameters are measured.
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incidence. Figure 4 is a map of the compressor
pressure-ratio/flow characteristics. It consists of a
series of lines of constant corrected speed, each
showing the pressure-ratio characteristic of the
compressor vs corrected flow at that corrected speed.
The concepts of corrected speed and corrected flow
allow the effects of compressor inlet temperature and
pressure to be taken into account, and, in fact, these
concepts represent the tangential Mach number at the
tips of the first stage blades, and the axial Mach
number of the airflow entering the compressor. These
parameters will be defined below in the section on
compressor mapping for single-shaft gas turbines.
Stations
5
7
8
13
T5
T T8
P3
P6
P7
P8
2
2.5
3
12
T3
P3
P 2
4
Figure 2. Instrumentation required for performance
monitoring on a two-shaft aeroderivative
gas turbine.
The assumptions used for the single-shaft calculation are applicable. Table 2 gives the thermodynamic
properties at the stations shown in Figure 2.
Table 2. Assumed Properties of the Working
Fluid for Performance Monitoring
Station
Specific Heat at
Constant Pressure kJ/(kg-°K)
Btu/(lb-°R)
1.005
1.047
1.172
1.11
1.089
1.047
0.24
0.25
0.28
0.265
0.26
0.25
2
3
4
5
7
8
Ratio of Specific
Heats
1.40 (y c )
1.40 (y c )
1.33 (Y t )
1.33 (Yt)
1.33 (Y t )
1.33 (Yt)
Figure 3. Damage to compressor of a heavy-duty
industrial gas turbine due to surge.
As corrected flow is reduced along a line of
constant corrected speed, the pressure ratio increases
until the stall of the blades is imminent. This point
is considered the surge point, and the line joining
these surge points at all corrected speeds is the surge
line. It is the boundary line between surge and
aerodynamic stability. Figure 4 also shows a typical
curve of the operating points of a compressor in a gas
turbine. This operating line is designed to assure a
reasonable flow (or pressure- ratio) margin on surge,
allowing for possible variations in installation
features and engine deterioration or damage, each of
which could lower the surge line and reduce the margin.
The equation for airflow W a is
we-
(H3,Wf + h s Ws )
-
cp ,T7 (Wf +Ws)
(3)
cp7 T7 -cp2 T2
The equation for the turbine inlet temperature T4 is
T4
cp3 T3 Wa +Hv Wf +h s Ws(4
)
Cp4 (W.+Wf +W.)
COMPRESSOR MAPPING
Compressor surge is one of the most severe hazards
in the operation of a gas turbine. When a compressor
goes into "hard stall" or "deep surge," a series of
shock waves passing back and forth through it is
generated. These shock waves cause the blades to
deflect, with two possible types of damage:
1) Clashing, or the deflection of the blades forward
into the preceding row of stationary vanes, with
possible fracture of blades and/or vanes, and
extensive subsequent damage throughout the
compressor, and
2) Clanging, or the differential tangential deflection
of blades in the same row, with impact at the tips,
resulting in bent or cracked tips.
Figure 3 shows damage to a compressor in a heavyduty industrial gas turbine as a result of a surge.
Compressor surge occurs when the airflow through the
compressor is restricted in some fashion, and the
compressor blades stall because of the resulting high
Surge Line
Operating Line
d
Lines of Constant Corrected Speed
Corrected Airflow
Figure 4. Compressor map showing surge and operating
lines.
3
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One of the more important protective functions of
the control system for a gas turbine is to avoid surge
In single-shaft machines, the control system schedules
fuel flow so that, at full throttle, a straight-line
relationship between compressor discharge pressure
(CDP) and exhaust gas temperature (EGT) is maintained
as illustrated in Figure 5. This schedule is designed
to keep the turbine inlet temperature (TIT) constant,
but, in the absence of internal deterioration, the
compressor cannot get into surge as long as this
schedule is maintained.
Compressor Discharge Pressure--Kg/cm
7
8
9
10
11
12
13
1100
2
14580
U
rn
1080
560 0
a)
0 1060
1 1040
0)
1020
540 2
0
0
a-
d
a-
1000
520
E 980
E
N
N
960
3
a
500
940
X 920
w
900 130
0
L
w
480
140
150
160
170
180
190 200 210
220
Compressor Discharge Pressure--Psia
Figure 5. Typical schedule of exhaust gas temperature
(EGT) vs compressor discharge pressure (CDP)
for a single-shaft gas turbine.
As noted above, the compressor of a single-shaft,
industrial gas turbine can only get into surge if there
has been some internal deterioration; this deterioration may result in moving the operating line toward
surge, or it may lower the surge line, reducing the
surge margin (Dundas, 1986). Two types of deterioration
move the operating line toward surge: fouling of the
compressor blades and closing down of the 1st stage
turbine nozzles, restricting the airflow through the
machine. Erosion of the compressor blades, particularly
at the tips, lowers the surge line by lowering the
pressure ratio at which the tips will stall, and,
consequently, at which generalized stall of the
compressor occurs. The rounded leading edges of the
blades, and the locally increased clearances, result in
poorly-defined incidences at the tips, and this condition degenerates into stall. Clogging or fouling of the
inlet filters by dirt, or by icing, if nonuniform, may
also trigger local pockets of rotating stall at pressure ratios below the surge line, providing a more
ready transition to premature surge.
It is well-established that compressors must be
washed or cleaned before fouling becomes so severe that
the compressor is in danger of surge. The need for
washing is signaled by some deterioration of performance, such as an increase in heat-rate, or a reduction
in CDP at full throttle. It is expected that washing
will restore the performance: sometimes a series of online washings are performed to check and slow the rate
of deterioration of the performance parameters, with
periodic shutdown for a slow-speed cleaning to bring
the performance back up to baseline (Thames, Stegmaier,
and Ford, 1989). If this washing and cleaning process
is not successful, the cause of the reduced performance
must lie elsewhere; later in this paper, a scheme for
determining the causes of performance deterioration
will be described.
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It is clear, however, that the performance deterioration that is really being addressed is a movement of
the compressor operating line to the left-toward the
surge line. It is obvious, then, that a determination
of where the compressor is on that map is of more value
than a simple review of the heat-rate. Changes in heatrate are caused by other factors, and the possibility
of surge, or the proximity thereto, may be masked by
other factors. A procedure of locating the operating
points of the compressor on a compressor map would
provide a much more reliable indication of the need for
compressor washing.
Water or steam is often injected into the combustors of gas turbines for NOx emission control.
Additional steam may also be injected into the compressor discharge section for power augmentation. This
additional mass flow may backpressure the compressor
(since the first stage turbine nozzles run choked and
control the total flow through the engine). Because
such backpressure pushes the operating point on the
compressor map toward surge, the total quantity of
steam injected into the high-pressure airstream (HP
steam) must be restricted. This is done by a schedule
of HP steam flow as a function of CDP.
Therefore, in the case of such engines a more
persuasive case for compressor mapping can be made.
Aeroderivatives have more parameters that vary with
power setting, inlet conditions, steam flow, including,
in the case of aeroderivatives with dual-spool gas
generators, two variable rotor speeds. It is considerably more difficult to determine whether or not the
compressor section needs cleaning, than in the case of
the single-shaft constant speed gas turbine. An illustration will clarify this point.
Figure 6 shows a plot of CDP vs HP steam flow
(steam injection into the high-pressure (HP) section of
the gas generator) for an aeroderivative gas turbine.
The limiting steam-flow line specified by the manufacturer is shown. A number of operating points taken over
a period of a few months are also shown; in particular,
two almost identical points (Points 1 and 2) at part
power are highlighted. Shortly after the filled-in
point (Point 1) was recorded the gas turbine experienced a deep surge. However, there was no surge when
Point 2 was taken. Further, the heat rates at the two
points were identical.
450
•°
a
31
440
430
v
C)
30
KPPH = 453.6 kg7i
420
410
29
0' Point 2
400
u
0 390
380
E
♦
HP Steam Flow
Limit
26
u
^o
a
E
360
350
t
0
27
♦
y
0'
28
Incident
(Point 1)
370
E
i
a'
0
25
30
40
50
60
70
HP Steam Flow—KPPH
80
90
Figure 6. Plot of HP steam flow vs CDP for an aeroderivative gas turbine. Points 1 and 2 are
nearly-identical operating points at the
steam control limit at part power. The
control limit is shown.
The operating points shown on Figure 6 were then
plotted on a compressor map. The monitoring system used
with the gas turbine has an algorithm to calculate an
estimated airflow from the supervisory instrumentation
readings; this is necessary for compressor mapping.
The result of this mapping is shown in Figure 7.
Point 1 is clearly to the left of the body of successful operating experience, and, therefore, is at least
trending toward surge, while Point 2 is well within the
probable stable region. The difference between the
points having otherwise identical supervisory parameters was in the gas-generator exhaust temperature: at
Point 1, the temperature was up against the limit,
while at Point 2, the temperature was 60°F (33°C)
lower. The possible effect of this difference on surge
margin would not have been detected, however, without
mapping the operating points.
The pressure ratio is
P3
Incident
(Point 1)y,
0
rr 28
, 26
T2 )
The value of Y c from Table 1 or 2 may be used in (10).
The above parameters of corrected airflow, pressure
ratio, corrected speed, and compressor efficiency can
be plotted on a compressor map for trending purposes.
Figure 8 shows a composite map of efficiency vs corrected flow and pressure ratio vs corrected flow for three
values of corrected speed. After a number of operating
points are plotted in this fashion, the internal condition of the compressor can be evaluated readily. The
figure shows the trends that can be detected. A reduction in airflow through the compressor indicates the
need for washing. If washing does not correct the flow
restriction, and, in particular, if the efficiency is
distinctly lowered, the compressor should be opened,
and the blades examined for erosion or other extensive
damage. A baseline map should be developed over time to
indicate where the compressor can operate successfully.
The envelope of successful operating points should then
be treated as a surge line. If surge or stall has been
experienced, points on the actual installed surge line
can be plotted.
It can be shown that the corrected flow represents
the Mach number of the flow through the compressor and
that corrected speed represents the Mach number of the
blades as they rotate. In a high-speed machine, such as
an axial-flow compressor, these parameters have more
significance with respect to performance than axial
velocity and rpm.
4
♦
/^1
d 25
24
T
23
22
260
270
sec
= 0.454 k
5r
310
280
290
300
Corrected Airflow—Lb/Sec
320
1
(10)
Point 2
•
27
Yc
Yc
P2
^ Z
o 29
(9)
P2
The adiabatic efficiency of the compressor is given by
Assumed
Surge Line
30
= 3
R
32
31
P
P
330
Figure 7. Compressor map for aeroderivative gas turbine
showing typical operating points, as well as
the two nearly-identical operating points at
the steam control limit in Figure 6. An
assumed surge line is also shown.
COMPRESSOR MAPPING--SINGLE-SHAFT GAS TURBINE
Corrected flow and corrected speed are the actual
airflow and actual speed multiplied by correction
factors that account for variations in actual inlet
temperature and pressure. The temperature correction is
IE
99
0
T
T(5)
S
I
.82
and the pressure correction 6 is given by
81
16
P
PZ(6)
Ns=107%
,5
N4=100%
o is
S
Nx =97.5%
13
12
where T 2 and P 2 are the actual inlet conditions, and T s
and P are the selected standard conditions, such as
21° C 1294 °K) at sea level (14.7 psia or 76 cm Hg).
The corrected airflow W* is then given by
W4
°
W/
g
vi 1,
i
a ,0
D O Good Com p ra
♦ ■ Im aired Com
6
7
(7)
1 Ib/s = 0.454• k5/s
20
21
22
23
24
25
26
27
Corrected FIow--W 4 Ib/s
28
29
30
and the corrected speed N* is
N'— N
Figure 8. Performance trending on a compressor map.
,/U (8)
-
5
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COMPRESSOR MAPPING--AERODERIVATIVE GAS TURBINES
It is even more important to include compressor
mapping in a program of performance monitoring of aeroderivative gas turbines than it is for single-shaft
machines. The former are much more complex aerothermodynamically, especially with the full steam injection
systems currently employed, and simple approaches to
assessing their health may not detect a compressor
operating point that is toward the surge side of
previous successful operating experience. The case
history of the steam-injected aeroderivative gas
turbine is a dramatic example of this possibility.
Compressor mapping can be performed as described, using
the data obtained from the instrumentation of Figure 2,
together with Eqs. 3 and 4.
In the case history referred to above, the gas
turbine was a dual-spool aeroderivative, but it was not
possible to map the LP and HP compressors individually,
because P 2was not printed out in the performance
logs, even €hough it was sensed. It is clearly important that this parameter be included in any printout of
dual-spool aeroderivative parameters, so that any
compressor deterioration that might lead to surge can
be traced to the appropriate spool.
BLADE-PATH TEMPERATURE SPREAD
Monitoring the circumferential temperature spread
in the plane of the engine or gas-generator exhaust is
another important procedure for avoiding serious
internal damage. Figure 9 illustrates the principle for
a gas turbine having individual combustor baskets. A
ring of thermocouples is placed in the exhaust, with a
thermocouple in line with each of the baskets. This
harness and the associated data reduction hardware may
do double duty--measuring and averaging the temperatures to determine the control value of EGT, as well as
recording the difference between the minimum and
maximum readings (the blade-path temperature spread)-or there may be two harnesses--one for measurement of
EGT, and the other for determining the spread.
Blade-Path
Temperature Spread
may signify a problem with the fuel nozzles or with one
or more of the combustors. Manufacturers often set
operating limits on the spread. Following is a typical
set of such limits:
If spread is greater than 90°F (50°C), operator
alert;
If spread is greater than 108°F (60°C), shut down
engine;
If more than 5 thermocouples have failed, shut down
engine;
If more than 3 adjacent thermocouples have failed,
shut down engine.
The use of blade-path spread monitoring helps avoid
three main types of internal problem, as discussed in
the following sections.
FUEL-NOZZLE MALFUNCTION
Blade-path spread instrumentation can be used with
pressure-measuring instrumentation in the fuel system
to detect malfunctioning fuel nozzles. In many heavyduty industrial gas turbines operating on fuel oil,
fuel is metered to each fuel nozzle by pressuresensitive valves known as flow-dividers. Pressure
sensors can be located downstream of each divider, and
upstream of each nozzle. If there is an excessive
blade-path spread and a simultaneous fuel-pressure
maldistribution, the problem can be traced to an
individual fuel nozzle.
COMBUSTOR BURNTHROUGH OR STRUCTURAL DAMAGE
There are various causes of combustor damage, such
as local burnthrough, structural failure, and heat
distortion. These modes of damage can result in dangerous hotspots in the basket. These problems can be
detected from the blade-path temperature spread, and
corrective action may be possible before widespread
combustor section damage occurs.
The No. 4 combustor basket in a large industrial
gas turbine ruptured, with the broken pieces passing
into the turbine section and producing extensive impact
damage. The blade-path spread had increased from 3040°F (17-22°C) to well over 100°F (55 ° C) during the
previous week. The operator's logbook for two weeks was
examined, and the individual thermocouple readings were
plotted as shown in Figure 10. The plots show that at
least one of the combustors must have had severe structural damage for at least a week prior to the incident.
There was clearly sufficient warning of this, and the
gas turbine should have been shut down at a much
earlier stage to replace the damaged basket or baskets.
The severe dip in the circumferential temperature
pattern in Figure 10 does not correspond to the angular
location of the No. 4 combustor. This is because the
gas flow develops a circumferential swirl as it expands
through the turbine, and the distortion pattern is
rotated from the turbine inlet to the turbine exhaust.
TURBINE BLADE FAILURE
Combustors
Temperature
in Exhaust Plane
Figure 9. Definition of blade-path temperature spread.
Figure 9 shows the use of blade-path spread thermocouples for individual combustor baskets. The principle
can also be applied to gas turbines having single
annular combustors. In such configurations, a harness
of from 6 to 18 equally-spaced thermocouples is used.
Under ideal conditions, all thermocouples register
the same temperature; conditions, however, are never
ideal, and a pattern of circumferential temperature
distortion exists. The degree of this distortion
Control of blade-path temperature spread is also of
importance in preventing turbine blade failures due to
high-cycle fatigue as a result of resonant vibration.
Uneven combustion results in a circumferential distortion of the gas flow into the turbine section. The
distortion takes the form of circumferential variations
of temperature, and, therefore, of pressure and velocity, with consequent circumferential variations of
aerodynamic loading on the turbine blades as they
rotate around the flowpath annulus. These distorted
patterns of loading are comprised of sine waves, or
harmonics, of all orders (numbers of complete sine
waves seen by a blade as it passes through one revolution). Normally, the lower orders (1,2,3, and 4) would
be strongest in such a pattern of distortion.
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Thermocouples
1
1020
aa)
980
d
940
900
E
860
820
X
w780
2
3
4
5
6
7
89
10 11 12
iiiuiuu
1Ii1iii
•UUUhlii 1I•1
UtiáIUI
0
30
60
90
U
540
a)
a'
U
0
520
Q)
E
500
480
460
a)
0
Q
E
Cr
-
N
440
0
L
x
420
W
2
120 150 180 210 240 270 300 330 360
Figure 10. Deterioration of blade-path temperature
spread in ten days prior to a severe combustor failure.
3
4
5
Harmonic Order
Polar Angle--deg
Figure 11. Relative strength of harmonics in the bladepath temperature spread pattern in an
industrial gas turbine.
MONITORING OF CYCLE PARAMETERS
If a turbine blade has a low-order resonance, it
will be excited by the corresponding harmonic. When the
spread is low, the strength of the excitation may not
be sufficient to initiate a fatigue crack in the blade;
however, if the spread is allowed to increase significantly, the strength of the excitation and, therefore,
the magnitude of the vibratory stresses in the blade,
may result in fatigue failure. The operator of the gas
turbine may never know whether such a low-order
resonance (which would be a design fault) exists until
the blade fails. It is necessary, then, to control the
spread to the low level associated with good combustion
to avoid such a possibility. This does not guarantee
against blade failure, since the strength of the
excitation may be excessive even at normal levels of
spread (the response of the blade depends on a number
of factors, among which is the damping available), but
it guards against the possibility of aggravating an
existing low level of resonant vibration.
A certain model of industrial generator-drive gas
turbine experienced a series of fatigue failures of
turbine blades. It was determined that the natural
frequency of that blade in its 1st bending mode was
just about 180 cps, that is, it had a 3rd order
resonance almost exactly at the operating speed of 3600
rpm. Operators of the machine were advised to maintain
the blade-path temperature spread at a low level, using
alarm and trip settings established by the manufacturer, until the appropriate redesign could be effected. A
Fourier analysis was made of the temperature spread in
one installation, with the spread at a fairly high
value. The relative strengths of the low-order harmonics result are shown in Figure 11. The 3rd order,
corresponding to the blade resonant order, proved to be
the strongest; this indicated that the blade resonance
was critical and should have been avoided. The only
recourse available to an operator, however, was to
monitor the spread and keep it as low as possible.
A third technique of performance monitoring for failure
prevention involves recognizing the patterns of deviation of the measured parameters from baseline data;
this includes evaluation of the percent deviation to
determine whether the indicated internal discrepancy
has progressed to a dangerous extent.
CYCLE PARAMETERS--SINGLE-SHAFT GAS TURBINES
(Dundas, 1986) showed how the technique of deviation pattern recognition can be applied to detect
internal discrepancies that could lead to surge for a
single-shaft generator-drive gas turbine. The reference
contained a chart of deviations for four discrepancies:
1. Fouling of the inlet filter or icing of the inlet;
2. Compressor fouling;
3. Compressor damage or erosion;
4. Closing of 1st stage turbine nozzles.
Figure 12 expands on that chart to include turbine
damage or erosion and opening of the 1st stage turbine
nozzles. Each of the internal discrepancies has a distinct and unique pattern of deviation of the indicated
parameters that permits it to be recognized. Appropriate maintenance action can then be taken. The data must
be taken under conditions consistent with the baseline
data. The compressor inlet temperature should be the
same (baseline data could be taken at a number of inlet
temperatures to provide a database for use at various
later inlet temperatures). Data should also be taken at
full throttle, i. e., right up against the EGT limit.
Thus, the concept of deviation pattern recognition is
most suitable for base-loaded gas turbines that generally run at full load.
In addition to helping avoid surge, monitoring and
correction of internal discrepancies can help avoid
compressor and turbine blade failures due to resonant
vibration (Dundas, 1982). Internal discrepancies
leading to performance deterioration often disrupt the
pressure balance across the rotor, and may result in
wiping of the thrust bearing, particularly on the
inactive face.
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Operating Parameters
P 2P3
13
Output
Fuel Flow Heat Rate
EDT
+ + = + 3
+ + _ 4 4 + +
= _ 4 y + 4 _
= 4 4 + 4 4 4
Fouling of Inlet Filter
or Inlet Icing
Compressor Fouling
Compressor Damage
or Erosion
Turbine Damage
or Erosion
= 4 4 4 4.,
= 4 4 4 4
Opening of 1st Stage
Nozzles
Closing of 1st Stage
Nozzles
% CORR T
+2
0
-2
+2
% CORR N
0
4 4
4 3,
-2
+2
% COR R N
-2
+2
Figure 12. Chart of performance-parameter deviation
patterns for various internal discrepancies.
% CORR P
0
-2
0
100 200 300 400 500 600 700 800 900 1000
Engine
CYCLE PARAMETERS--AERODERIVATIVE OR TWO-SHAFT GAS
TURBINES
Figure 13a. Trends of cycle parameters characteristic
of fouled compressors in aeroderivative gas
turbines.
Figure 12 presented a method for diagnosing
internal conditions in a single-shaft gas turbine based
on trending of the various parameters read by the
supervisory instrumentation. It is very difficult to
construct a similar chart for aeroderivative gas
turbines, particularly those with dual-rotor gas
generators. One of the reasons for this difficulty is
the fact that the gas generator speed, or speeds, can
vary, depending on the type of deterioration, independently of the power turbine output speed, which may be
kept constant, as in a generator-drive application.
Manufacturers have, however, compiled curves of the
variations of the key parameters for some types of
deterioration. Figures 13a, b, and c are typical curves
for a dual-spool aeroderivative. The cycle parameters
for trending are the corrected gas-generator exit
temperature, the corrected gas generator speeds, and
the corrected compressor discharge pressure; these
parameters are expressed in percentages of maximum
values to keep the numbers close to unity for easy
comprehension of the trends. All parameters should be
recorded at maximum power setting for the inlet
temperature at which baseline data was obtained.
Referring to Figure 2, and letting the subscript 0
refer to the baseline (undeteriorated) parameters, and
with given by (5), the trending parameters are given
by the following equations:
Percent corrected gas-generator exit temperature
(T5),
S—
T5—T50
T
%T
Hours
+2
j % CORR T
5
j
0
-2
+2
% CORR N
0
-2
+2
% CORR N
0
-2
+2
% CORR P
0
Z 0
100 200 300 400 500 600 700 800 900 1000
Engine
Hours
Figure 13b. Trends of cycle parameters characteristic
of fouled turbines or worn internal seals
in aeroderivative gas turbines.
- 2% CORR T
x100
0
508
_2
(11)
+2
0
Percent corrected low-pressure compressor speed (Ni),
-2
+2
Percent corrected high-pressure compressor speed (N2),
11
=
0
-2
Nl —N1O x100
+2 7 CORK P
0
( 12)
Ni
i
-2
-4
0
100 200 300 400 500 600 700 800 900 1000
Engine
(
%N • — NZ—NZ 0
N20
(13)
Hours
Figure 13c. Trends of cycle parameters characteristic
of blocked or bowed 1st stage turbine
nozzles aeroderivative gas turbines.
8
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and percent corrected compressor discharge pressure
(P3),
P3-P30
% ^73 l
•
P30
X100 (1k)
P2
:
IMPLEMENTATION OF PERFORMANCE MONITORING FOR FAILURE
PREVENTION
The techniques of performance monitoring described
above can be implemented readily through computercontrolled data acquisition systems. The use of such
systems in gas-turbine installations is becoming widespread throughout the utility, cogeneration, pipeline,
and other industries. Selected data can be directed,
either from the supervisory instrumentation directly or
from the engine control hardware, to a separate analysis module. This separate software module can be
programmed to perform the indicated calculations and to
enter the results into pre-designed graphical formats.
When called for by the operator through a workstation,
the calculated information can be returned to the host
computer and displayed numerically or graphically in
real time on the workstation screen.
The illustrations in the following sections were
prepared using the Powerlog Process Monitoring System,
available from Powmat Ltd. of Ballston Lake, New York.
The software for performance analysis in this system
was written by Dr. Daniel Sullivan, one of the authors
of this paper. Off-line performance data from a large
heavy-duty single-shaft unit was provided by another
author, Frank Abegg. This data was used for the
illustrations of compressor mapping and blade-path
spread analysis. Since the units from which the data
was taken have been on standby in recent years, they
were not suitable subjects for deviation pattern
analysis. The data was instead based on the model used
in (Dundas, 1986). This model was idealized, but it
does provide consistent results for illustration of the
pattern analysis procedure.
e••o,
—'-
wo
Corrected Sp.sd--W 1b/s
Real—Tana Operating Point
1
.:
sao
Assumed Surge Line
Figure 14. Compressor map with operating points, as
displayed on workstation screen.
BLADE-PATH TEMPERATURE SPREAD
The depiction of blade-path temperature spread in
real-time at an operator workstation is readily
effected using the Powerlog software. Figure 15 is an
illustration of one possible display for an engine
having individual combustor baskets. The distribution
of pressures, as well as the pressure spread, to the
fuel nozzles is also shown. The diagram of baskets and
nozzles can be color-coded to show the angular locations of the highest and lowest readings. A further
refinement could be to develop an algorithm to correct
the diagram for the swirl effect.
COMPRESSOR MAPPING
The performance data indicated in Figure 2, taken
over a single day was entered into a software module,
and the calculations for airflow, corrected flow and
pressure ratio (Equations 1,5,6,7,9) were performed for
each point. The results were then entered into the
Powerlog Graphics module and displayed on a workstation
screen, as shown in Figure 14. A wide range of operating conditions is represented in the figure; such a
wide range of conditions could be compiled over a
longer period of time, and considered as representing
the stable operating regime for the compressor. A
putative surge line, representing the limit of successful operating points on the compressor map could then
be drawn as shown (again using the algebraic and
graphing capabilities of the analysis module). This
line could be saved permanently in the compressor-map
format file.
At the request of the operator at a workstation,
the compressor map, including an assumed surge line
based on previous successful experience, would be
displayed on the screen with a real-time point. The
operator could then judge whether the compressor was
operating too close to the assumed surge line, and take
the appropriate action (such as compressor washing) to
move the operating point back to the more stable
region.
Figure 15. Workstation display of blade-path temperature and fuel-nozzle pressure spreads.
DEVIATION PATTERNS
The results of the parametric calculations performed for the study of (Dundas, 1986) were used to derive
the deviation patterns for the various internal discrepancies shown in Figure 12. Figures 16a through e
show how these would be displayed on the workstation
monitor in real time, using the Powerlog analysis
software. The software was programmed to read the
indicated parameters for a given operating point (at
the EGT limit) and to calculate the percent deviation
from the baseline point at the same compressor-inlet
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(a)
(c)
(b)
(d)
Figure 16. Workstation displays of deviation patterns. (a) Compressor fouling; (b) Compressor damage
or erosion; (c) Closing of 1st stage turbine nozzles; (d) Opening of 1st stage turbine
nozzles; (e) Turbine damage or erosion.
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temperature. These deviations were then plotted on bar
charts having deviation warning limits as shown. The
software was programmed also to recognize the patterns
of deviation using logic equations, and to show the
type of internal discrepancy existing on the display.
Depending on the amount of the deviation from
baseline, the operator can take the necessary action to
correct the indicated internal discrepancy. This may be
as straightforward as washing the compressor or cleaning the inlet filter, or it may entail dismantle of the
gas turbine to inspect the compressor, combustor, or
turbine section. Additional workstation displays can be
devised to inform the operator of possible courses of
action. Among these might be a check on the compressor
operating point or blade-path spread (through the
displays described earlier).
Various additional monitoring features can readily
be devised using a system such as Powerlog. For
instance, the control system can be programmed to alarm
when a selected parameter exceeds the preset deviation
limit; the operator can then activate the deviation
pattern display to assess the reason for the alarm and
to initiate the indicated action. Maintenance efforts
can be directed with maximum efficiency by this means.
This study of implementation of deviation pattern
analysis was applied to a single-shaft generator drive
gas turbine, because a theoretical basis for it was
available in the reference. Its application to a twoshaft gas turbine, and especially to an aeroderivative
having a dual-rotor gas generator, would be much more
complex. As far as is known, there is no existing
theoretical basis for such an application, but, of
course, one could be derived using the capabilities of
current cycle analysis software. The benefits of such
an effort would be even greater in terms of early
recognition of internal discrepancy and optimization of
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maintenance efforts because of the number of components
and the probable complexity of the deviation patterns.
SUMMARY
The three most important aspects of performance
monitoring of gas turbines for failure prevention are
compressor mapping, monitoring of blade-path temperature spread, and analysis of the patterns of deviation
of measured cycle parameters.
Compressor mapping is essential to determine if the
compressor operating point is approaching the surge
line because of compressor fouling or warping of the
1st stage turbine nozzles. Blade-path spread can forewarn of serious combustor damage due to faulty combustion for whatever cause. It should also be controlled
to guard against the possibility of turbine blade
failure in resonant vibration.
Modern computer-controlled data acquisition systems
are a very powerful tool for analysis of measured
performance parameters on a real-time basis to recognize the existence of internal discrepancies rapidly,
to diagnose them accurately, and to minimize maintenance and correction time.
REFERENCES
R. E. Dundas, "A Study of the Effect of
Deterioration on Compressor Surge Margin in
Constant-Speed, Single-Shaft Gas Turbine Engines,"
ASME Paper 86 -GT -177, 1986.
J. M. Thames, J. W. Stegmaier, and J. J. Ford, Jr.,
"On-Line Washing Practices and Benefits," ASME
Paper 89- GT -91, 1989.
R. E. Dundas, "The Use of Performance-Monitoring to
Prevent Compressor and Turbine Blade Failures,"
ASME Paper 82- GT -66, 1982.