chapter
26
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Steam Turbines for
Power Generation
INTRODUCTION
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26.1
full speed units and the higher costs, cross compound units are
not currently in favor. Casing configurations for fossil units are
typically high-pressure unit, intermediate-pressure unit, and one or
more low-pressure units exhausting to the condenser. The highpressure and intermediate-pressure units are frequently combined
into one casing (HP–IP).
Steam turbines use both reaction and impulse designs. The fundamentals of these blade design types will be discussed as they are
currently applied. The selection of the type blading has a significant impact on the turbine design.
Steam turbines employ digital control systems that can provide
automatic turbine control to avoid mal-operation. These systems
can be supplied by the turbine supplier or by others. In some instances these interface directly with the overall plant and boiler
control systems.
Equipment supplied for the last 20 years have greatly increased
inspection intervals per the recommendations of the turbine suppliers. The inspection interval for major inspections can be 8 to
10 years. Sometimes the inspection interval is defined in terms of
equivalent operating hours (EOH) which includes the number of
startups in the determination.
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Used with permission of Siemens Energy Inc.: The statements, recommendations and conclusions set out in this chapter
are those of the author and do not necessarily reflect those of
Siemens Energy Inc. or its affiliates.
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Harry F. Martin
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Steam turbines have historically been the prime source of power
for electric power generation. Turbines come in a variety of types
with regard to inlet and exhaust steam conditions, casing and shaft
arrangements and flow directions. This chapter will focus on steam
turbines currently being applied to power generation. The steam
conditions will include those currently applied to fossil fired power
plants, combined cycle and nuclear power units.
Currently, fossil fired plants, whether coal fired or gas fired, typically supply steam at 1800 to 3500 psig steam pressures with 950 to
1050°F main steam and reheat temperatures. Double reheat units are
few and will not be discussed in this chapter in much detail. Combined cycle plants (CC) typically have multiple drum heat recovery
steam generator (HRSG) with inlet temperatures of 1050°F for main
steam and reheat steam (see Chapter 27). While no new nuclear units
have been installed in recent years, plans are underway to apply the
AP1000 nuclear systems (see Chapters 23 and 24). Past steam conditions were typically saturated steam (0.2% moisture) and reheat temperature of 500°F. All current applications are condensing designs
with regenerative extraction, except for CC units. Current application can be used up to 8 to 10 inHgA exhaust pressure. The application of air cooled condensers is more prevalent in CC applications.
Current designs are axial flow turbines. However, in some applications, radial flow stages are used in the inlet stages. Typical
arrangements are tandem compound. This means more than one
turbine casing’s rotors are coupled together on the same shaft. For
example a turbine train consisting of a high-pressure turbine (HP),
an intermediate-pressure turbine (IP) and two low-pressure turbines (LP) would have all of these in line and coupled to an electric
generator. Cross compound units (two or more shafts) are currently
in use. Some arrangements have the HP and IP on one shaft along
with a generator and the LP turbines on a separate shaft along with
another generator. In many instances the low-pressure turbine shaft
is running at half speed and has larger annulus areas to reduce LP
turbine exhaust velocity and associated leaving losses. However,
due to increased blade height and associated exit annulus area for
26.2
GENERAL INFORMATION
Any discussion of steam turbines should include some thermodynamic concepts. The first law leads to the writing of the general
energy equation for a steady flow processes.
(PE)1 + (KE)1 + u1 + (pv)1 + Q = (PE)2 + (KE)2 + u2 + (pv)2 + W
(26.1)
Where:
PE = the gravitational or stored mechanical energy of the fluid
KE = the mechanical kinetic energy of the fluid
u = the internal energy of the system
pv = the pressure volume product or flow work
Q = heat added
W = shaft work
Since enthalpy (h) is equal to u + pv, the work for a flow system
with no heat added, or negligible changes of potential and kinetic
energy can be written as:
26- • Chapter 26
The speed of sound can be written as:
V* = (dp / dr)1/2
(26.5)
Where:
V* = velocity at the critical point or the sound velocity
p = pressure
r = density
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dA
A
=
[1 − M2 ]
dp ρV 2
Where:
A = area
V = Velocity
Carnot cycle
Replace ","(26.2)
with ":"
W = Dh
(26.6)
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From this relationship we can deduce the flow passages for different flow velocities. These are summarized in Figure 26.3. The
significance of this figure is the differences in the behavior of a
diffusing passage between subsonic and supersonic diffusers and
the development of supersonic velocity in turbines.
From the definition of an isentropic process the following useful
relationships can be developed:
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Mach No. (M) is defined as the ratio of the local velocity to the
local sound velocity.
The relationship between area and pressure for an isentropic
flow process can be written as:
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Please move
"or" down one
line or further
out as shown.
In previous edit
request, we
asked that it be Fig. 26.1 "set off" from
the equations.
h = work / heat added
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The second law of thermodynamics dictates that all cycles must
reject heat. The efficiency of a cycle can be defined as,:
or
h = (heat added – heat rejected) / heat added
(26.3)
h = (Ta – Tb) / Ta
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A reversible cycle is one that is reversible in all aspects and
could produce the same thermodynamic state points run in either
direction. This of course is an idealization. However this concept
is useful. For example the Carnot cycle is defined as one in which
heat is added and rejected at constant temperature and the expansion and compression is accomplished with no losses. This cycle is
shown on a temperature entropy diagram in Figure 26.1.
The resulting efficiency of such a cycle is defined as:
 V 2 − V12 
h1 − h 2 =  2

 2gJ 
(26.7)
The condition when the velocity is zero is referred to as the stagnation state or total condition:
ht − h2 =
V22
2gJ Restore "as:" Please see
previous document
(26.4)
This is the maximum possible efficiency for a cycle operating
between temperatures Ta and Tb.
The steam turbine cycle is a Rankine cycle. This cycle consists
of compression of water, heating and evaporation of steam in a
boiler, expansion in turbine and heat rejection in a condenser. A
theoretical version of this cycle is shown as Figure 26.2. Most
units use a reheat cycle with regenerative heating though the use
of feedwater heaters using steam extracted at various locations
throughout the turbine (see Chapter 29). Regenerative heating
reduces the heat added and the heat rejected by the cycle. While
the output power is also reduced, the net effect is an improved
cycle efficiency.
There are a number of gas dynamic concepts that have significant impact on turbine design and application. These will be discussed in the following paragraphs:.
Please change ":" to "."
Fig. 26.2 Theoretical Rankine cycle
(26.8)
ENERGY AND POWER GENERATION HANDBOOK • 26-
W=
wU
(Vθ1 − Vθ2 )
g
(26.12)
For steady flow with uniform conditions and no leakage we may
write:
W=
Fig. 26.3 Subsonic and supersonic expansion
w
ϖ (r1Vθ1 − r2 Vθ2 )
g
(26.13)
Where:
v = rotational speed in radians/second
r = radius
w = flow rate
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Where in the English system:
J = 778 ft-lbf/Btu
g = 32.2 lbm ft/lbf sec2
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Writing this equation in terms of enthalpy produces:
From perfect gas relationships and the continuity equation
Pleasethe
change k-1
following non-dimensional flow term can be developed:
w 
V 2 − V22 
to k+1 in exponent.
(26.14)

W=
h1 − h 2 + 1
g 
2gj 
2
k−1 

Please remove "g"
mR Tt
 p2  k  p2  k  
(26.9)
= 2gJCp 
 −



A2pt
from theA turbine stage consists typically of a stationary row of blades,
 pt 
  pt 

and a rotating row of blades. The stationary blades are often redenominator
ferred to in
as this
nozzles. The stationary row and the rotating blades
act together
to allow the steam flow to produce work on the rotor.
first term.
Please
Where:
stages are included in a turbine casing to produce the shaft
m = mass flow
do notMany
change
output to drive the generator.
T = temperature in degrees Rankine
terms inside
Turbine stages, with the exception of last stages of the low R = gas constant
parenthesis.
pressure turbine, are typically classified as either an impulse stage
Cp = specific heat at constant pressure
or a reaction stage. In a pure impulse stage the entire stage pres k = specific heat ratio
sure drop is taken across the stationary row of blades. However, in
practice this is not practical and about 5% to 20% of the stage drop
A plot of non-dimensional flow and pressures ratio is shown
is taken across the rotating blade. The addition of another stationas Figure 26.4. Note the dashed portion of this curve illustrates
ary and rotating row to this stage forms a “Curtis” stage. However,
the double value of this equation. The flow reaches a maximum
this type of design is not in use in current designs for large steam
at the critical pressure ratio (p*/Pt). For real gases such as steam,
turbines.
the values of critical pressure ratio and acoustic velocity are found
The velocity triangles and general stage arrangement are shown
in steam tables. However, there are many instances where perfect
in Figure 26.5. Referring to this figure, the power developedwork done
gas analysis is very useful in steam turbine analysis.
by an imThe purpose of blading in a turbine is to transform the kinetic
pulse turbine can be written as:
energy of the incoming fluid into useful work. Newton’s second
law of motion can be applied to develop a momentum theorem.
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(26.10)
Please change
"work done" to
"power developed".
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w
Fθ = (Vθ1 − Vθ2 )
g
Where:
F = force in the tangential direction
V = velocity in the tangential direction
We can derive m the following equations:
W=
w
(U1Vθ1 − U2 Vθ2 )
g
replace
Where:
W = work power
U = turbine wheel speed
Vo = tangential velocity
w = mass flow rate
(26.11)
"work" with "power"
For a constant diameter the relationship may be written as:
Fig. 26.4 Flow vs. pressure ratio
26- • Chapter 26
W=
wU 2  cos α1   cos β2 
− 1 1 +
g  v
  cos β1 
(26.15)
replace "work" with "power"
Where
v = velocity ratio
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M
W=
(26.16)
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The typical constructions of these two types of turbine stages are
shown as Figure 26.6. Due to the higher pressure drops across the
stationary row in an impulse stage, the sealing for the stationary
row is at a lower diameter. However, a flow system design involving the leakage along the shaft, the sizing of the pressure balance
holes in the rotor and the possible inclusion of a seal at the inner
diameter between the rotating and stationary row is required in an
impulse or Rateau turbine. This is discussed in Reference [1]. In
an impulse stage the design should set the reaction at the hub to be
positive to ensure good stage performance. Therefore, some degree
of reaction is required. Considering the radial variation of station-
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Fig. 26.5 Velocity diagram for simple stage
wU
(2v1 cos a1 - U)
g
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Please capitalize "V",
per previous edit
request document.
Reviewing Equation 26.15, the advantages of small values of a1
and b2 are evident. This is a characteristic of impulse blading.
In a reaction turbine pressure drop occurs in both the stationary
and rotating blades. In a reaction turbine there is an increase in the
relative velocity leaving the passage. In a symmetrical stage, the
enthalpy drops are equal for the fixed and rotating rows. This is
a 50% reaction stage. For a 50% reaction turbine the powerwork can be
written as:
Fig. 26.6 Impulse and reaction stage
ENERGY AND POWER GENERATION HANDBOOK • 26-
ary exit pressures, the reaction is usually increased as the stage
heights get larger as you proceed through the turbine. Generally
these stages are designed such that there is a radial in flow into the
cavity between the stationary row and the rotor.
Important stage characteristics of a turbine stage application are
velocity ratio, and stage loading coefficient. The velocity ratio is
defined as the ratio of the blade wheel speed to the velocity that
would be obtained by the isentropic expansion through the stage
stagnation pressure drop.
n = U / Cis
The stage loading coefficient is defined as the work done in the
stage divided by the wheel speed squared.
ψ=
∆H
U2 (26.18)
1∂P Vθ2
∂ Vr
∂ Vr =
− Vz
− Vr
ρ∂r
r
∂ Z
∂ r
(26.19)
The first term on the right hand side shows the effect of a swirling flow such that the static pressure increases with radius. This
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Impulse turbines can have higher stage loadings and therefore,
fewer stages than can be used efficiently for a reaction turbine for
the same expansion. This generally means higher losses, as will be
discussed later.
Figure 26.7 shows the variation of efficiency for nominal reaction and impulse blading with velocity ratio. The figure illustrates
the advantages of impulse blading for lower stage velocity ratios
and higher blade loadings. In addition, the benefit of reaction blading when lower stage loadings can be utilized is illustrated.
In the first stage of a high-pressure turbine, an impulse or Rateau
stage is often utilized. This stage is commonly called the Control
Stage. Currently, single stage control stages are utilized. However, in
the past Curtis stages were used. The Curtis stage was popular when
large temperature drops were required due to rotor and blade mate-
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Where:
v = velocity ratio
U = wheel speed
Cis = isentropic stage velocity
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(26.17)
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rial limitations for high temperature inlets. The Rateau type control
stage has better performance than the Curtis type control stage.
The control stage combined with inlet valves serves as the flow
control for the turbine. As load is reduced the control valves can be
operated in a sequential valve mode with constant pressure operation meaning that a 4 valve unit could operate with between 1 and
4 valves open and the control stage operating at admissions of 25%
to 100%. Therefore, the control stage blade loading at 25% arc of
admission is significantly higher and has a higher pressure ratio than
at 100% admission. This is true since the pressure downstream of the
control stage is essentially proportional to flow and the nozzle inlet
pressure is still at 100% pressure. The nozzle is typically choked at
this condition but the absolute velocity entering the rotating blade is
supersonic. At partial admission, the losses are higher, however, the
reduced throttling loss and the higher inlet pressures provide overall
part load cycle efficiency improvements. There are many operating
strategies that exist with today’s plants and control systems. These
will be discussed in more detail in 26.6. However, a partial arc design when operating at part load will generally produce a better part
load heat rate than a typical full arc design.
In the low-pressure turbine, the steam is expanded from inlet
pressures of 60 to 230 psia to typical condenser pressures of 1 to 3
inHgA. This increase in specific volume requires rapid increases in
flow area especially in the last few stages. In these stages the design is significantly different than typical impulse or reaction blade
design mentioned in the preceding paragraphs. In these stages the
work done is quite large. The LP turbine can contribute up to 60%
of the unit output and the last stage alone can produce 10% of
the total output. Therefore, the efficient design of these stages is
important.
The stage design is driven by consequences of the radial equilibrium equation in the Meridional (axial–radial) plane.
Fig. 26.7 Effect of stage type on aerodynamic efficiency
26- • Chapter 26
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Fig. 26.8 Pressure distribution last stage
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effect is significant at the exit of stationary rows where high tangential velocities occur. The second and third terms are referred
to as the streamline curvature terms. The effect of the streamline
curvature terms become significant at end wall regions of the blade
path where wall tapers and abrupt changes occur.
Figure 26.8 shows the pressure drop across the stationary blade
row is greater at the base than at the tip. Consequently, the steam
velocity leaving the base or inner diameter is greater than at the
tip as shown in Figure 26.9. The space between the stationary and
rotating blades (see Figure 26.10) is the region with significant
change in relative inlet angle to the rotating blade. There is a radial variation of reaction. Therefore, the base sections approach an
impulse design with high turning. Therefore, blades with a large
amount of twist are required.
Computerized flow field programs both axi-symmetric and
three-dimensional are used in the design of low-pressure end blade
paths. The process usually starts with the axi-symmetric approach.
Very early versions of these methods are discussed in References
[2] and [3].
Annulus areas have increased with larger blades and high tip
speed. Tip speeds exceed Mach 2 and exit relative velocities are
supersonic. Hub-to-tip ratios have been reduced from the stan­
dard 0.5 of the past to values near 0.4. New LP flow field and stationary blade design concepts have permitted this change, while
maintaining hub reactions to acceptable levels. Many designs use
Fig. 26.9 Velocity distribution between stationary and rotating blades
ENERGY AND POWER GENERATION HANDBOOK • 26-
Condensation shock losses
Braking losses
Drag losses
Miscellaneous losses
Condensation shock losses occur when the rapidly expanding
steam, after crossing the saturation line, fails to reach equilibrium.
In other words although the moisture level has theoretically reached
a level as high as 3.5%, fog formation has not taken place. The
longer the delay in reaching equilibrium the greater the buildup of
super-saturation and the larger the loss.
Braking losses occur when moisture strikes the rotating blades
and causes a negative torque.
Drag losses are of two types. Drag losses due to fog drops and
drag loses due to large drops torn off trailing edges of stationary
rows. The loss occurs due to dragging and accelerating of the droplets by the steam.
Miscellaneous losses include centrifugal losses of water being centrifuged out by the rotating blades, boundary layer losses
caused by the waviness of the deposited water film, the losses generated by the kinetic energy of the blades being transformed into
heat and continuous under cooling losses. The under cooling is the
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converging diverging sections to limit losses. Stationary blade designs utilize blade lean and sweep [4, 5] to control the flow field
and limit losses.
The performance of the LP turbine is significantly affected by the
performance of the diffuser and the flow path directing the steam
to the condenser. Diffuser design is currently done using computational fluid dynamics (CFD) computer programs. Reference [6]
discusses effective use of such computer codes. The inclusion of
the effect of the last row blade flow distribution is emphasized in
the reference. Modeling, using uniform flow distributions or model
test data using uniform distributions, has shown to be ineffective in
predicting exhaust diffuser performance.
Figure 26.11 shows the variation in total exhaust loss as a function of volumetric flow for three LP turbine designs using a variety
of last row blade sizes and exhaust annulus. This factor is a significant influence in selecting which design should be applied to
a given cycle. However, one must also consider the variations in
condenser pressure and load demand through the year to select the
correct configuration.
Leaving loss is the kinetic energy of the absolute velocity leaving the last row (L-0R) of the LP turbine. This is generally defined
in terms of Btu’s. Actually there are leaving losses for the exit
of each turbine. However, the leaving loss of the last stage of the
LP turbine is the most significant. The leaving loss for the L-0R
for an LP turbine is obtained through integration of the exit flow
properties obtained from a CFD calculation. The leaving loss is a
part of the total exhaust loss. Figure 26.12 shows a diagram representing the expansion of the last stage of an LP turbine on an
enthalpy-entropy diagram (HS diagram). Note that the stage exit
static pressure, Ps-b, is less than the downstream condenser pressure. Therefore, the exhaust diffuser has recovered some of the exit
kinetic energy from the blade path. The total exhaust loss (T.E.L.)
is the value shown in Figure 26.11 at various volumetric flows.
At low volumetric flows, the flow at the base of the blade will be
recirculating and high hood losses are obtained.
Transonic turbine blades behave similarly to converging-diverging (CD) nozzles. This is true even if the blade is not a CD section
per se. The throat area of the blade at this point is sonic. As the
exit pressure decreases the passage shock wave will move down-
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Fig. 26.10 Blading velocity
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stream towards the exit plane and the blade loading will increase.
At a certain pressure, the flow in the passage is fully expanded and
the blade loading reaches its maximum value. Further decreases in
back pressure will not affect the flow pattern inside the blade passage downstream of the throat of the blade. Therefore, the blade
loading will not change with lowering of the exhaust pressure.
This condition is referred to as “limit load.” This is the reason that
the load–vacuum correction curves provided by turbine suppliers
reach a point where the power of the unit will not increase with
further decreases in back pressure. Figure 26.13 shows isobars for
a blade passage at the transonic point and after limit load has been
reached.
LP turbines expand into the wet region, and they usually employ moisture removal devices to reduce moisture content. This
moisture removal has a twofold benefit. First, the performance is
improved through the reduction in moisture losses in the blade
path and the reheating effect caused by the reduction of moisture.
Second, the less moisture available the less blade tip erosion will
occur.
The causes of moisture losses are:
Fig. 26.11 Total exhaust loss
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26- • Chapter 26
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Fig. 26.12 LP turbine last stage expansion
Fig. 26.13 Limit load
ENERGY AND POWER GENERATION HANDBOOK • 26-
( ρ) 2
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TURBINE CONFIGURATIONS
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The primary applications of steam turbines utilized in power
generation are used in conventional fossil fired (coal or oil) power
plants, nuclear units and combined cycle power plants. Most new
applications in the past 15 years have been combined cycle units.
In fossil plants the sizes range from 100 to 1500 MW. Typical steam conditions are 2400 psig and 1000°F main stream and
1000°F reheat. However, higher pressure and temperature units are
operating. New fossil units have been few but significant efforts
have been made in retrofitting existing plants to improve reliability
and efficiency.
Figure 26.15 is a longitudinal section of a fossil steam turbine
rated at greater than 800 MW The main steam inlet pressure is
supercritical and main steam and reheat temperatures are 1050°F.
This is single reheat cycle. The unit is a tandem compound design
with a shaft speed of 3600 rpm. There are eight feedwater heaters
in the cycle. The design utilizes shared bearings. This means that
each elements rotor does not have two bearings.
Flow enters the single flow HP turbine using side inlets through
control valves (not shown) and flow through the blade path to the
HP exhaust. The HP cylinder exhaust is in the base and steam flows
to the reheater. The HP turbine is a single flow design therefore,
requiring a dummy piston for thrust balancing. Feedwater heating
(26.20)
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1− 1
Equation 26.20 was shown to produce high flow rates for straight
through seals in Reference [9]. This reference provides useful coefficients for certain commonly used seal designs.
Finite element analysis computer codes and individual supplier
special computer codes are the dominant mechanical design tools
in use today. However, it is useful to verify design concepts with
an appropriate equation. Reference [10] provides explanations and
equations that have been used for this purpose.
N + ln ρ
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Where,
W = leakage flow, lb/hour
A = area, sq. in.
Po = Absolute inlet total pressure, psia.
Vo = inlet specific volume, ft3/lbm
r = ratio of inlet total pressure to exit static pressure
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1700KAP0
W=
P0 V0
N = number of throttling
K = kinetic energy annihilation coefficient (usually used as 0.8
for step seals
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loss associated with the rapidly expanding steam not being able to
catch up with the equilibrium moisture level.
Not only is there a loss associated with non-equilibrium but the
flow passing capability of the stage can be affected. Depending on
the wetness, the flow passing capability could be increase by 5%
to 6%. Reference [7] provides information on both efficiency and
flow passing effects of moisture in turbines.
Addressing last row erosion has become more important with
the larger tip speed designs in use today. Different blade designs
use a variety of protection. Stellite tips have been brazed onto the
upper portions of the blades. Blades have been flame and laser
hardened. The protection strategy is a function of the blade material used. Reference [8] provides both a description of the erosion
process and relative erosion resistances of various materials. The
primary source of erosion are the droplets that come from water
collected on the upstream stator blade. The water droplets are torn
off the blade trailing edge and accelerated to a speed less than
steam speed. Therefore, the droplets have a relatively low absolute velocity but a high relative velocity to the rotating blade. See
Figure 26.14 for the velocity diagram of this process. The water
droplets will vary in size up to the largest droplet size that can be
supported in the flow stream.
There are many types of sealing currently used in steam turbines
depending on the application. A good general equation for seal
flow is the “Martin’s equation.” A form of this equation can be
written as:
Fig. 26.14 Droplet impact velocity
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26-10 • Chapter 26
The LP is a four-flow 2 cylinder design. The LP inlet zone uses
a flow guide to reduce losses. The blade attachments are tee root in
the front end and side entry at the exhaust end. The last row blade
is a free-standing blade. The exhaust diffuser outer diameter section is attached to the inner casing and the inner diameter the diffuser section is part of the outer cylinder. The diffuser performance
is somewhat negatively affected by the LP outer cylinder bracing.
The unit is down flow exhaust to the condenser. The rotor again is
a no bore rotor except for test bore locations. The extraction zones
are contained in the inner cylinders. The extractions in each turbine
are symmetrical. However, the lowest pressure and the next to the
lowest pressure heaters are supplied from different LP turbines.
Most fossil units being manufactured today use no bore rotors.
However, at least one manufacturer supplies welded rotors. These
welds are usually between similar materials but welding between
dissimilar materials is also done depending on the application.
The unit in Figure 26.16 is for 50 Hz application and therefore
rotates at 3000 rpm. It is very similar to the unit in Figure 26.15 except that it is a six flow LP end (3 double flow turbines). The unit is
also for supercritical pressures. Some components are scaled from
50 to 60 Hz or from 60 to 50 Hz depending on the original design.
LP end rotating blades are typically scaled.
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steam is extracted from the HP section blade path and the HP exhaust. The turbine is a full arc design and inlet pressure exists between the inner and outer cylinder for a portion of the HP section.
The HP extraction chamber is formed between the inner and outer
cylinder. All of the rotating rows use single tee root attachment to
the rotor. The rotor does not have a bore except for a test specimen
bore located away from the main body of the rotor. The stationary
blades are attached to the inner cylinder.
Flow enters the IP turbine through side mounted reheat valves
(not shown). The IP turbine is a double flow design without the
need for a dummy piston. The IP first row is a diagonal low reaction design. This design reduces the axial space requirements of
the blade path and provides a lower relative total temperature to
the first rotating row for mechanical strength considerations. The
blade attachments are tee root. The IP flow exits to the LP turbines
through the cover using a large diameter large turning radius elbow. Extraction flow is supplied from two internal locations and
the IP exhaust. The internal extractions are unbalanced with the
higher pressure feedwater heater supplied from the generator end
and the lower pressure heater from the turbine end. The extraction
zones are contained inside the inner casing. The IP rotor is no bore
except for the test section.
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Fig. 26.15 Longitudinal section of supercritical pressure turbine 4 flow LP for 3600 RPM
Fig. 26.16 Longitudinal section of supercritical pressure turbine 6 flow LP for 3000 LP
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ENERGY AND POWER GENERATION HANDBOOK • 26-11
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The LP rotor and inner cylinder were retrofitted. The LP cylinder utilizes a single inner cylinder design, and the LP extractions are symmetrical and separated by the partitioning of the
inner cylinder. There is a no bore rotor. The blades use tee root and
side entry blade attachments. The last stationary blade (L-0C) uses
slots on the surface of the blade to remove moisture for improved
performance and reduced erosion. In addition, a low diameter seal
is used for the last stationary row (L-0C). Figure 3.2 of Reference
[11] shows the LP design of the same supplier as Figure 3.1 of this
reference. Note all low diameter sealing is used and the front end
stages are impulse design.
Figure 26.18 shows essentially the same design HP–IP as Fig­
ure 26.17, except that the HP uses a partial arc design with a control
stage. The space between the exit of the control stage and the inlet
row of the reaction blade path facilitates uniform flow into the reaction blade path during partial admission for improved perfor­mance
and reduced non-uniform blade loading effects. The nozzle block
of the control stage is a slide in design compared to previously used
bolted on nozzle blocks. The control stage rotating blade is mounted
to the rotor using a pinned root design. Note that the control stage
blade is quite wide relative to the width of the other HP stages. This
is due to high loading requirements during partial arc admission.
Current combined cycle units do not have feedwater heaters but
instead have inductions (reference Chapter 27). Typically, there
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Figure 26.17 shows a retrofitted unit. The HP is a full arc admission design. The high- and intermediate-pressure blade paths
are combined into one casing (HP–IP). The retrofitted components
are the inner cylinder and rotor. The rotor is a no bore design. The
HP–IP inner cylinder appears much thicker than in reality. This is
due to the fact that longitudinal section includes a large support
rib. The rib controls cylinder deformation. The HP uses integrally
shrouded reaction blades with tee root attachments. The IP blade
path utilizes a diagonal low reaction stage to limit axial space required for the blade path and to reduce relative total temperature on
the first rotating row for mechanical reasons. The IP rotor utilizes
reaction stages with double tee root blade attachments for the integral shrouded blading. The IP blade path exhaust flow turns and
flows between the inner and outer cylinders. This provides cooling
for the inner and outer cylinders. There are three dummy pistons
in this design (HP, IP, and LP) to balance thrust. The HP and the
IP blade paths are thrust balanced independently. There is an equilibrium pipe connecting the HP exhaust to downstream side of the
HP dummy. Spring back seals are utilized over the rotating row of
blades in the HP and IP. The dummies use spring back seals and in
some cases retractable seals (see section 26.5). Another supplier’s
HP–IP turbine design is show as Figure 3-1 of Reference [11]. This
design uses impulse type blading and has a disc diaphragm type
construction. This requires essentially no dummy piston.
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Fig. 26.17 Longitudinal section of HP–IP and double flow LP
Fig. 26.18 Longitudinal section of HP–IP turbine with control stage
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26-12 • Chapter 26
the use of airfoil shaped supports in the base of the turbine. These
supports are also used for oil flow to the bearing and instrumentation connections. The outer cylinder has two vertical joints. These
are; the connection of the IP and LP cylinders and the connection
of the LP cylinder and the exhaust diffuser section. After initial
assembly, the joints need not be broken and the cylinder can be
lifted as one outer cylinder. A large dummy leak off connection
is shown connecting the low-pressure side of the dummy to the
induction zone.
Figure 26.20 shows the HP turbine for a CC plant using an impulse blading design. Note the smaller dummy piston. The rotor is
machined to a disc diaphragm construction with axial pressure balancing holes to control leakage flow and axial thrust. The sealing
mounted in the stationary blades is on a low diameter with small
radial clearance spring back seals.
Figure 26.21 shows the longitudinal section of an HP–IP, LP
design for combined cycle application. This design is used for
higher megawatt applications than the arrangement shown in Figure 26.19. The design uses an HP–IP design similar to what is used
in typical fossil power plants. The LP is a double flow unit using a
titanium last stage and has side exhausts since the application uses
side mounted condensers.
The HP–IP turbine uses a large inner cylinder that is also the
carrier of the stationary blades and HP dummy seals. The design
utilizes dummy pistons to balance the thrust. The HP thrust is balanced with the HP dummy while the IP thrust is balanced with a
unique dummy piston arrangement at the turbine end of the shaft.
The HP is a full arc of admission design.
The latest designs applied to nuclear units have been in the retrofit market. New units will most likely reflect this approach for
design. Especially if a verified design concept is required. Nuclear
units typically run at half speed relative to fossil applications.
Therefore, in the countries using 60 Hz electrical systems the rotational speed would be 1800 rpm.
Figure 26.22 shows a retrofit application for nuclear HP turbine
that formerly had a control stage design. Nuclear units operate primarily at full power and the application of full admission units has
some advantages. In all applications the ability to pass the licensed
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is an induction from the intermediate-pressure drum downstream
of the HP turbine and a low-pressure drum induction at the LP
inlet or inside the LP turbine. Therefore, the flow increases from
the HP inlet to the exhaust of the LP turbine, as opposed to flow
decreasing in a conventional fossil steam plant with extractions
for feedwater heating. For this reason, the HP turbine typically has
smaller blade heights and sealing is more important for improved
efficiency. However, due to the increasing flow downstream of the
HP turbine, the incentive for very high HP efficiencies is less in
CC power plant.
Figure 26.19 shows a combined cycle stream turbine with 1800
psig inlet stream pressure and main steam and reheat temperatures
of 1050°F. The design is an HP–IP/LP design using shared bearings. The HP is a separate single flow turbine with a barrel construction. These types of turbines do not have a horizontal joint on
the HP outer cylinder. The inlet is volute design and inlet pressure
exists for almost the entire section of the space between the inner and outer cylinders. The unit has a bottom HP exhaust. These
applications have non-return valves in the piping going to the reheater since they are operated with bypass systems. The HP rotor is
a no bore design uses tee root integral shrouded blades.
The IP/LP design is a straight through design from the IP inlet
to the axial flow exhaust. The inlet is a bottom side entry. A very
large dummy is required at the inlet end for thrust balancing. Due
to the large exhaust area of the IP/LP design the unit is not independently thrust balanced. The region between the inlet to the first
stage and the dummy is cooled with HP exhaust steam to provide
long creep life to the rotor in this area. The rotor is welded between
the IP and LP parts of the shaft to permit the use of different materials to satisfy high temperature and high strength requirements of
the different ends of the turbine (see Section 26.4). The rotor uses
integral shrouded blades in IP and front stages of the LP sections.
Interlocked blades are used in the LP end. The last stage of this
particular turbine uses a titanium blade due the very large blade
height.
There is an induction between the IP and LP blade paths. This
induction may not be operative below 20% load. The exhaust uses
an axial diffuser. The design of this diffuser is improved through
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Fig. 26.19 Longitudinal section of an HP and IP/LP turbine for combined cycle application
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ENERGY AND POWER GENERATION HANDBOOK • 26-13
Fig. 26.20 Impulse HP turbine for combined cycle applications
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rying the stationary blades and providing extraction chambers. The
blades are tee root construction.
In the nuclear cycle, there is only an HP and LP turbine due to
the relatively low inlet pressures of nuclear cycles. Between the
HP exhaust and the LP inlet, steam usually flows through a moisture separator reheater. The separator is required since HP exhaust
steam is wet. The reheater uses either HP inlet steam or HP inlet
and HP extraction steam to reheat the HP exhaust steam after the
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power flow is a critical criterion. Non-equilibrium two phase flow
has been shown to be a significant issue in some applications [7].
This unit uses a diagonal design first stage. The unit has unbalanced
extractions in the HP. One extraction supplies the high-pressure
feedwater heater and the other supplies the first stage of a two stage
reheater for superheating the flow to the low-pressure turbine. The
inlet shows a diagonal stage with a flow guide to reduce losses.
The design features a small inner cylinder and blade rings for car-
Fig. 26.21 HP–IP, LP turbine for combined cycle application
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26-14 • Chapter 26
Fig. 26.22 HP nuclear turbine
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computational fluid dynamics (CFD) has made complicated analysis routine. This has improved the performance and reliability of
today’s steam turbines. The use of advanced analysis tools has also
improved costs through more effective material utilization.
Steam turbine materials for the current applications have not
changed significantly over the years but the processing of the materials has improved to reduce impurities. Blading materials are
typically 12%Cr stainless steels. However, Cr content can be as
high as 16%. Currently, there a number of titanium designs for
the very high tip speed blades. These are mostly being operated
in combined cycle units. For moisture protection of low-pressure
blades, either flame hardening or stellite shields brazed to the
blade are used. Some alloys, that cannot be flame hardened, can
be laser hardened. Titanium blades, in a combined cycle application, do not require shielding. Reference [8] provides useful
information on droplet erosion and the relative resistance of materials to erosion.
Most HP and IP rotor materials are made of low alloy steel.
These steels are nominally 1%Cr. However, the rotor should have
high strength and toughness at the high operating temperatures.
Most rotors are made from one forging per element. However,
some designs utilize smaller forgings welded together at the OD to
DESIGN
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26.4
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moisture is removed. Moisture removal effectiveness of today’s
separators is quite high and essentially all of the moisture is removed before reheating.
Figure 26.23 shows a nuclear LP turbine. The unit is a double
flow turbine using tee root and side entry blade attachments. This
design utilizes a shrunk on discs design. The first stage disc is
keyed, as well as, shrunk on to the rotor. Many nuclear LP turbines
today use integral rotors to mitigate high stress and stress corrosion. However, the disc design of Figure 26.22 has been verified
with service and analysis and is accepted in the industry.
The extraction zones are separated by sections of the inner cylinder. These extractions in the wet regions of the blade path also
provide moisture removal. In addition, moisture removal slots are
located at the outer boundary of the flow passage and slotted hollow stationary blades are employed at the L-0 location.
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Turbine design technology has developed to analysis based
methodology as opposed to a scaling or experienced based methodology. The application of finite element analysis (FEA) and
Fig. 26.23 LP nuclear turbine
ENERGY AND POWER GENERATION HANDBOOK • 26-15
number in steam is generally near one the use of relative total temperature is appropriate.
Trt = Ts +
(26.21)
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The control of relative velocity is often used to reduce the operating temperature of the components such as blades and rotors.
This concept for cooling IP rotors is discussed in Reference [17].
For LP rotors, solid rotors, welded rotors and shrunk on disc rotors are used in the industry. For shrunk on disc rotors the required
disc shrink fit is evaluated with the objective to preclude discs
coming loose during startup transients and overspeeds. Some discs
are keyed to the shaft intending to prevent the disc from overspeeding relative to the shaft in the event that the disc would come loose.
Fracture mechanics is also applied in the design of LP rotors.
Rotors are evaluated for high cycle fatigue. The calculations
need to address all anticipated loading on the shaft. In addition,
the effect of torsional stress for short circuits and the stress due to
misalignment are included.
Thrust balance of the rotor is usually done such that HP, IP and
LP turbines are balanced independently. This limits the number
of design cases needed to check thrust and the size of the dummy
pistons. Large dummy pistons are not required with impulse blading. This leads to lower dummy leakage rates. In some designs
independently balancing of thrust is not practical such as the HP,
IP/LP design shown in Figure 26.19. In this case, the axial thrust
of the produced in the LP end of the IP/LP is too large to contain.
Therefore, the thrust can be offset through a built in unbalance in
the HP or the thrust bearing is selected to accommodate the maximum loading.
In the past, double inner cylinder construction was typical for
LP turbines. However, with the use of FEA, the single inner cylinder design can be designed for the required operating life. However, the blade carriers may be made of a higher alloy. This is
particularly true when moist steam is present in the blade path at
that location. The increase in carbon content can help reduce the
potential for flow assisted corrosion. The casings need to provide
sufficient flow area for the inlets and extraction areas to prevent
large pressure losses in these areas. The peak stresses of these designs usually occur at full temperature. In this case, the cyclic duty
of the casing can be assessed just knowing the times and temperature ranges during operation.
Since the outer pressure boundary of the LP inner cylinder is the
condenser, thermal shields are generally used over most of the cylinder outer diameter. These help to limit the thermal stresses in the
casings. The large axial gradients seen in LP turbines and the stiffness variation in the casing, lead to the potential for ovality at the
exhaust end of the machines, prudent design changes and appropriate setting of radial clearances are used to address this situation.
HP and IP blade paths are typically designed on a unit specific
basis. This is done, in order to better match the individual cycle
requirement, while achieving high levels of efficiency. Typically,
the process is computerized to achieve optimization. The process
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Vr 2 2gJCp
Where:
Trt = the relative total temperature
Vr = relative velocity,
g = gravitational constant
J = Joules constant
Cp = specific heat at constant pressure
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form the rotor body. The design of rotors for future higher temperatures will most likely utilize more welded construction. Reference
[12] discusses design changes for higher temperature applications.
In this reference the welding of 12%Cr and 3.5NiCrMoV is discussed. Figure 26.19 illustrates a current application where the LP
and IP rotors are joined by welding. This is discussed in Reference
[13]. Some rotors of this type have used solid rotors with different
heat treatment to obtain the appropriate high strength for the LP
and the higher temperature properties for the IP end.
For HP and IP rotors, fracture mechanics evaluations are made
with the objective that the any flaw that cannot be detected in the
rotor does not propagate to critical crack size during the life of the
rotor in order to preclude a rotor burst. This evaluation includes
the material toughness and the effect of operation during the target
life of the unit.
In many applications the HP and IP cylinders are designed to carry
the stationary blades as opposed to have a special blade ring. Figures
26.15, 26.16, 26.17, 26.18, 26.19 and 26.21 show this type of construction. In this arrangement, the inner cylinder must be designed
to maintain the required degree of roundness to provide proper orientation during operation. These cylinders typically use an axial rib
to aid in this aspect. Figure 26.19 illustrates the concept of a blade
ring construction in the IP portion of the IP/LP turbine. Figure 26.22
shows the blade ring construction for a nuclear HP turbine. Blade
rings provide some advantages for control of deformation to control
seal clearances but can increase cost and outage time at inspections
and other outages requiring significant disassembly.
HP and IP casings, rotors and valves are designed to a particular
life. The life factors in operating time and cycles. The cyclic life
will vary depending on the application. For example, a nuclear unit
does not see many startup and shutdown cycles or load changes.
Therefore, duty cycles are required, as well as, startup and shutdown profiles for cyclic duty evaluation.
Please
the"
HP and IP rotors generally
areremove
the limiting"of
items
in low cycle
fatigue, with HP rotor usually
the most critical.
The rotors are deas requested
in previous
signed with the objective that the limiting location meets the duty
document.
cycle requirements for the
application. The allowable stresses must
include the effect of high temperature on the fatigue strength. Typical limiting locations are in the blade attachment areas or in some
other fillet location at the intersection of a large and small radius
on the shaft, e.g., HP dummy piston. If the limiting blade attachment location is not visible, the geometry can be altered to make
this location visible during inspections. Stress-relieving grooves
are generally used to both relieve the stress of the blade attachments and to put the limiting stress location in a more inspectable
location.
The operating temperatures and transient temperature distributions are required for this type of analysis. The temperature throughout the turbine needs to be specified since most analysis is done
using the entire component, such as: rotor, cylinder or valve body.
A heat transfer analysis is done using appropriate flow information
and convective heat transfer coefficients. In some designs radiation
heat transfer is important, since steam is an absorbing re-emitting
medium [14]. References [15] and [16] describe in great deal heat
transfer analysis that can be applied to steam turbines. Some of
these correlations have been verified from actual steam turbine data.
Reference [15] even includes data verified from actual turbine rotor
bore temperature measurements. This reference is specifically for a
reaction turbine design. Reference [16] is primarily for an impulse
design. However, many of the methods are transferable.
For heat transfer the appropriate temperature to be used is the
relative adiabatic wall temperature. However, since the Prandtl
26-16 • Chapter 26
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heating (bypass operation in HP), and transient overloads are usually addressed in blade mechanical design margins.
The LP end blading design is usually the result of an iterative
process in which axi-symmetric and three-dimensional flow field
methods are used to produce a blade path that meets aerodynamic
efficiency and mechanical limitations. CFD codes are even
capable of some degree of unsteady analysis. For example in
Reference [4] unsteady analysis shows that the stationary blade
wake effect is more pronounced at the base of L-0R blade than at
the tip.
The blade profiles are generally Mach number dependent such
that sections with subsonic, transonic and supersonic relative exit
Mach numbers would have different shapes and velocity distributions. A converging diverging section is often applied at exit Mach
number is greater than 1.4. However, LP turbines are subject to
large variations in flow conditions due the large variation in volumetric flows. Therefore, each design should address the specifics
of the intended application.
In addition to improved airfoil aerodynamics, flow field effects
to improve performance have been aided by improved CFD capability. For example, the use of lean and sweep in LP stationary
blades increases performance. Figure 26.25 shows an illustration
of a stationary blade design using improved CFD analysis. This
concept is discussed in References [4] and [5].
The LP end blades are tuned blades. This means that all modes
of vibration below some harmonic of running speed are tuned. For
example one supplier may use the 7th harmonic. Free-standing and
interlocked blade designs are currently being used. The interlocked
designs have much greater stiffness after the blades are “locked
up.” Lock up occurs when the blade interlocking features come
into contact. This is typically between 50% and 70% of running
Fig. 26.24 Bowed blade
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would have prescribed limits that might include a set of basic secPleaseetc.
change
"of"process would setions, blade root attachments,
The design
lect blade section, stage
to loading
"or". and reaction in order to achieve
the best efficiency while meeting cycle parameters such as: inlet
temperature, inlet and exit pressures and flow. The blade could be
bowed orf just taper twisted. Typically, in the HP turbine, the blades
would be bowed due to the low aspect ratio, blade height to blade
chord. Figure 26.24 depicts a bowed blade arrangement for an HP
turbine. Reference [18] discusses the concept of bowed blading.
At low aspect ratios the reduction of secondary flow losses is more
significant. Bowed blades are also applied in IP turbines and the
front stages of LP turbines. See Reference [19] for a description of
one approach to automated design and optimization. This process
can be utilized in HP, IP turbines and for the un-tuned blades of
LP turbines.
The HP, IP, and front stages of LP blades are typically integrally
shrouded, or one piece construction as compared to older designs
that may have used shrouds that were riveted to the blade. The
shrouds typically are made tight at assembly for added mechanical
strength. Some integrally shrouded blades are designed to run with
small gapes between blades. The concept is to reduce tip deflection
and add mechanical damping. Typically, these blades are un-tuned
which means they do not have speed or frequency limitations during operation to address harmonic effects.
The blade attachments are typically tee root, double tee root (for
larger blades), and side entry roots. The designs need to address
aerodynamic and centrifugal loadings. Flow disturbances, windage
Fig. 26.25 Last stage stationary row with sweep
and lean
ENERGY AND POWER GENERATION HANDBOOK • 26-17
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Fig. 26.26 Interlocked LP turbine last row
rotating blade
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suppressed. In free-standing blades mix–tuning has been quite successful in eliminating un-stalled flutter. However, in interlocked
blades mix-tuning is not possible due to the interlocking. However, mis-tuning requires a more significant change to the blade
and attempts to change the aerodynamic characteristic of the blade
to blade interaction. This is quite difficult as explained in Reference [20]. However, a computer code (TRACE) that solves the
Reynolds-Averaged Navier Stokes equations in the relative frame
has been applied to analyzing un-stalled flutter. Figure 26.27 (taken
from Reference [20]) shows the computation of aerodynamic
damping for three different Mach numbers. The aerodynamic
damping is expressed in terms of logarithmic decrement. ND represents the nodal diameter pattern of the vibrating coupled blade
structure. Negative ND indicates a backward traveling wave while
a positive ND indicates a forward traveling wave. Un-stalled flutter
is always seen in the backward traveling wave. Negative log decrement means the blade is unstable. Combined with knowledge for
the mechanical damping of the system, this approach can be used
to determine the susceptibility to un-stalled flutter.
Moisture erosion of rotating blades in the wet steam region results from liquid droplets impinging on or near the leading edge.
See Reference [8] for a detailed description of this process. Damaging droplets are not the moisture in the primary blade flow. The
primary source of the water that erodes the blades comes from
moisture deposited on the upstream stationary row. This deposition is caused by inertial and diffusion processes. The moisture
flows to the trailing edge of the blade and drops are formed in
the slow moving wake behind the trailing edge of the stator. The
droplets are broken up into smaller droplets by the drag force of the
steam. This process creates a distribution of coarse moisture that
can range up to 500 microns or more. These droplets are slowly accelerated and carried to the rotating blade. Figure 26.14 illustrates
the water droplet velocity in this process.
Erosion protection is in general twofold: blade tip protection through flame hardening, or adding stellite strips, etc., and
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speed due to the centrifugal untwist of the blade. Tuned blades are
verified on prototype designs by use of a rotating test. The blade
frequency control for subsequent units typically uses some type of
stationary frequency test. The fact that these blades require tuning for reliable operation leads to operating speed restrictions at
no load and off frequency limits on line. However, not all modes
require a speed restriction. Figure 26.26 illustrates an interlocked
last row rotating blade design that includes a mid-span snubber and
a shroud interlocking feature. This figure also illustrates a four lug
side entry root “fir tree root.”
The blade attachment for these large blades is typically a side
entry fir tree root or a pinned root. The root steeple stresses will
limit the number of speed cycles a design can tolerate and thus
could affect the cyclic duty potential for some designs. Blade roots
are sometimes rolled or shot peened to improve their HCF, LCF,
and stress corrosion capability.
For LP end blades, the start of stall flutter can be assessed and
consideration for this built into the operating guidelines (see Section 26.6). However, un-stalled flutter does not occur at off design
conditions and therefore, the protection is in the design itself. In
the past a variety of methods were used to predict un-stalled flutter
such as:
· Dynamic cascade tests and analytical predictions based on
these results
· Model turbine tests
· Using experienced based criteria such as Strouhal number or
reduced frequency.
These methods have produced moderate success.
Mix-tuning is a process of slightly modifying a nominal blade to
achieve some prescribed natural frequency variation. In un-stalled
flutter, the blades vibrate at the same frequency. If there is sufficient variation in natural frequency blade to blade, this vibration is
Fig. 26.27 Aerodynamic damping for three
different Mach numbers as a function of nodal
diameter
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26-18 • Chapter 26
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Fig. 26.28 Seal configurations
portant in the HP turbine and becomes less of a factor as the blade
height increases. For this reason, impulse blading with its disc diaphragm construction with low diameter sealing (see Figure 26.6) is
sometimes applied for low flow small blade height turbine designs.
Seals in the LP turbine in this application use tip to tip sealing
that employ many seals but can accommodate large differential
expansions.
One new seal concept is the use of abradable seal material on the
seal carrier ring. The concept of this approach is shown in Figure
26.29. The clearance is reduced. However, in the event of a rub
the abradable material wears as opposed to seal. The leakage flow
reduction with this approach has been approximately 20%.
Currently brush seals are being applied in steam turbines (see
References [21, 22]). The concerns of large pressure ratio and high
fence height requirements have been addressed and brush seals
have been successfully operated in steam turbines. Figure 26.30
shows a diagram of a brush seal application. The brush seal is
installed with a clearance and the “bristle blow down effect” resulted in as much as a 50% reduction in leakage. Brush seals can
be installed in a variety of arrangements such as a spring back or
retractable seals.
Retractable seals have been used in certain locations in HP and
IP turbines. To date, applications have been in the glands and
dummy pistons. The concept is to use spring force to hold the seal
ring segments at a larger diameter until pressure forces overcome
the spring load and move the ring segments to a lower diameter.
This approach permits smaller radial clearances during operation
while having more rub protection during startup, thus reducing the
concern for seal rubs affecting seal performance. This concept is
discussed in Reference [23].
Honeycomb seals have been applied in LP turbines. These have
mostly been applied over free-standing blades (Figure 26.31). They
provide for better sealing and greater rub protection than would be
achieved with a hard liner. Therefore, clearances are smaller with
this seal compared to a hard liner.
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moisture removal. Moisture removal can employ slotted end walls,
slotted stators and even heating of the stator blade to evaporate
the moisture. Generally, the slotted stator uses two locations for
removal of stator blade surface water. One is located on the pressure surface and one on the suction surface. If a hollow blade is
used for evaporating surface moisture, the steam is extracted from
another location, generally a LP extraction zone, and piped into the
stator with appropriate drainage provisions. This water can be sent
either to the condenser or to a low-pressure feedwater heater.
Turbine exhaust hood design can have a significant impact
on turbine performance. Exhaust hoods can be either axial (see
Fig­ure 26.19), radial or as is most common radial/axial, see Fig­
ure 26.21). Axial diffusers in general provide the best performance.
The ideal situation is to have a diffusing exhaust hood where
the turbine blade exit pressure is less than the condenser pressure. However, this cannot exist over the entire operating range.
Therefore, the design point of the diffuser is quite important. As
discussed in Reference [6], CFD can be used quite effectively
in design. However, the blade path hood interaction needs to be
included.
For units that have significant condenser pressure variations seasonally, it may be prudent to design the diffuser to a slightly offdesign condition. For example, if a unit that needs to generate peak
power in the summer time, it may be prudent to sacrifice some of
the performance in the winter months in order to maximize the
performance when the power demands are the highest.
Sealing has made a significant improvement in turbine perfor­
mance. In addition, some of the sealing concepts being applied reduce the potential for seal rubs and therefore, reduce the efficiency
degradation. Turbine efficiency degradation is inevitable but the
degree of degradation between overhauls can be affected by seal
selection as well as prudent turbine operation (see section 26.6).
Figure 26.28 shows one supplier’s seal application philosophy.
The seals in the HP and IP turbines are stepped design for reduced
leakage. Because of the reduced blade height, sealing is most im-
ENERGY AND POWER GENERATION HANDBOOK • 26-19
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Solid particle erosion (SPE) damage to turbine blade paths is
the result of iron oxide particles (magnetite) entering the steam
turbine. These particles can result from exfoliation from the boiler
and steam piping. There have been many papers discussing the
generation of these particles and ways to remedy the situation external to the turbine. Reference [24] discusses this topic in great
detail. Minimizing exfoliation is a most effective way to reduce
SPE in steam turbines.
SPE is found in the regions of the high-pressure turbine (HP)
and intermediate-pressure turbine (IP) near inlets to the turbine.
Damage has also been seen in valves, cylinders, blades and rotors
in the dummy pistons. The HP damage can be in the blade path
for units with or without a control stage. If hard particle damage is
an issue, the control stage blades would be significantly impacted
primarily due to the high velocities at partial arc operation.
In the blade path, the larger particles cannot follow the steam as
it accelerates and turns in the blade passage. This causes the particles to impact blade pressure surface of the stationary row. Reference [24] provides calculations that show that damaging particles
do not travel at steam speed. However, designs with higher nozzle
exit velocities will produce higher particle impact velocities. The
higher velocity and greater steam turning make impulse designs
more susceptible to SPE. Particles that leave the nozzle will impact
the rotating blade with more negative incidence than the steam due
to the slower particle velocity. Therefore, particles will impact on
the inlet suction or convex side of the blade. This effect cannot
only cause leading edge erosion of the rotating blade but can under
certain instance produce particle rebound that can cause damage to
upstream nozzles suction surface.
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Fig. 26.29 Abradable coating
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While the impact of transient operation (e.g., Startup) is debated
in the literature, Reference [25] supports the impact of transient
operation and concludes that bypass systems have a significant impact on reducing hard particle erosion.
In general, reducing velocity and flow turning will reduce SPE.
Typically the impulse nozzle will have greater turning and higher
velocity than a reaction stationary blade design. Erosion would be
expected to vary with the kinetic energy of the particle, thus producing a variation with velocity squared. Reference [26] shows a
power law dependence of velocity on solid particle erosion with
the exponent varying between 1.91 and 2.52. Even in a reaction
blade path the reduction of stage pressure ratio would have a significant impact on SPE reduction.
The impact of solid particle erosion can be reduced through the
application of coatings. For control stages, coatings have been in
service for more than 26 years. The boride diffusion coating can be
applied to nozzle vanes and rotating blades as needed. The success
of this coating application has been documented to extend the period of maintenance due to solid particle erosion [27]. The time interval for SPE damage to be addressed could double when coatings
are used. The nozzle design can also have a significant impact on
erosion as discussed in Reference [28]. In addition, the increased
nozzle pitch reduces SPE. For many units with control stages,
valve management features are available that will permit switch-
Fig. 26.30 Brush seal
Please make figure
smaller per
previous edit
request.
Fig. 26.31 Honeycomb seal
26-20 • Chapter 26
26.5
PERFORMANCE
IP turbine
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The efficiencies of today’s HP and IP turbines are quite good
relative to the use of parallel sided blading and straight seals. Currently efficiencies can be as high as 90% for the HP turbine element including valve losses, and 93.5% for the IP turbine including
valve losses. The performance of the LP turbine will vary significantly with the level of moisture, the condenser pressure and the
exhaust arrangement.
Performance verification is an important subject. Overall heat
rate testing has been addressed by ASME PTC 6 committee. Reference [30] documents the code for standard fossil and nuclear units
with regenerative feedwater heaters and Reference [31] is for combined cycles.
Reference [30] now combines the full-scale and the alternate
test into one report. The PTC 6 committee recommends the use of
the code for conducting acceptance tests of steam turbines and any
other performance test where performance levels are to be obtained
with a minimum of uncertainty. In fact, the code recommends the
use of the full-scale test with a condensate flow measurement for
fossil unit steam turbines. However, the alternative test uses fewer
measurements and makes greater use of correction curves for cycle adjustments. The uncertainty of the full test is reported to be
±0.25% on unit heat rate, while the alternative test reports an uncertainty of 0.33%. For nuclear units the where most of the cycle
is wet the accuracy is between ±0.375% and 0.5%. For combined
cycles the test procedure of Reference [31] reports an uncertainty
of ±0.50%.
These tests are quite expensive due to the instrumentation requirements including a flow section for flow measurement. In
fact very few “pure” code tests are conducted since other arrangements are frequently made during the negotiation, installation or
test phases. Therefore, unit by unit specific test definitions are
frequently used for unit verification. When tests are conducted,
the condenser pressure is of particular importance. The vacuum
corrections are developed by the turbine supplier by a variety of
methods, such as CFD modeling, model testing, and experience
with similar units.
For measurement of HP, IP efficiencies a standard efficiency
test can be conducted to by measuring the pressures and temperatures (enthalpy drop test). Tests on HP turbines can be conducted
at valves wide open to eliminate valve losses in the data. In some
cases pressures downstream of the valves are measured and the
efficiency definition uses this point as the inlet condition. However, it can be difficult at times to measure pressures far enough
downstream of bends, and other flow disturbances to get accurate
pressure readings.
High- and intermediate-pressure turbine enthalpy drop tests are
typically within ± 0.6% uncertainty. In many designs, the perfor­
mance levels are affected by leakages from items, such as HP, IP
and LP dummy piston leakages, and inlet seal leakages. To determine the level in some instances tests referred to “influence factor”
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HP turbine
Profile loss
Secondary flow loss
Leakage loss
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The common method of defining performance for the overall
turbine is heat rate. This is defined as the heat added to the cycle
to develop one kW of output. This is really the reciprocal of efficiency. Typically heat rate is expressed in Btu/kW-h or kJ/kW-h.
The turbine heat rate only includes the heat that is transferred to the
steam and not the heat released by the fuel. The plant heat rate will
always be higher than turbine heat rate. Heat rates are calculated
using the ASME steam tables.
The blading and sealing currently utilized in steam turbines have
reached such high levels that significant improvements in heat rate
will require higher steam pressures and temperature. The direction
of these efforts was discussed in Section 26.4.
The performance of the steam turbine is obtained by computer
models that are proprietary to the specific turbine supplier. However, most models have similar methodologies in that the essential
losses are included and that some simplification is used to reduce
the complexity of the blade path to a computer model.
Typically, losses addressed are blade profile losses, secondary flow losses, leakage losses, and moisture losses. The blade
profile losses are typically corrected for surface roughness and
Reynolds number, Mach No., trailing edge thickness, and incidence angle. Secondary flow losses are associated with the turning of the boundary layer on the end walls of the blade passage.
The losses can be evaluated from cascade testing, CFD modeling,
model turbine testing, etc. These losses are typically correlated
with ratio of blade pitch to height. It is this loss that a bowed blade
will affect.
The application of multi-axis numerically controlled machines
to manufacture machined blades has eliminated the use of parallel
sided blades. In HP, IP, and the front stages of LP turbines using
reaction blading, shapes such as “bowed” blades are typically used
for performance improvements. Bowed blades typically shaped as
shown in Figure 26.24. The improvement comes from forcing flow
toward the hub and tip of the blade passage thus suppressing the
vortex that is developed in the passage. References [18] and [19]
discuss this concept in detail. In impulse designs bowed blades are
not typically used and flow control concepts are used to improve
performance.
In a typical reaction steam turbine the distribution of the average
losses are as a percentage of the total loss are shown as follows:
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ing from partial arc to full arc during startup. This can significantly
reduce the stage velocity and therefore reduce solid particle erosion at a time when hard particles are being generated. This valve
mode transfer feature permits the return to partial arc operation to
improve performance at part loads.
Future designs will need to address significantly higher pressure
and temperature. Designs are currently being developed for 5000 psi,
1300°F steam conditions. Therefore, there will be more use of
materials that utilize higher chromium (10-12%) for rotors and
casings. In addition, the need to address significant temperature
variation in the same casing will lead to greater use of welding
of dissimilar materials. In addition, there will be a need for more
Nickel based materials for casings. In a study reported in Reference [29] in 1992, it was concluded that the next step in future optimized steam conditions for a double reheat pulverized coal cycle
would be between 4000 psi, 1100°F/1100°F/1100°F and 5000 psi,
1200°F/1100°F/1100°F.
It is interesting to note that a 325 MW rated turbine generator
with 5000 psi inlet pressure and 1150°F steam inlet temperature
and double reheat of 1050°F/1050°F went into service at The Eddystone Generating Station in 1957. The unit operated until 1975
and is now an ASME Historical Mechanical Engineering Site.
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ENERGY AND POWER GENERATION HANDBOOK • 26-21
Fig. 26.32 Heat balance diagram
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tests (see Reference [11]) are conducted to evaluate these effects.
The need for this type of testing is certainly design dependent.
Figure 26.32 shows a turbine heat balance diagram for a supercritical pressure inlet HP turbine with single reheat. The inlet temperatures are 1050°F/1050°F. The unit has separate HP, IP and 2
double flow LP turbines. The unit has 8 feedwater heaters, which are
supplied by extractions from the turbine. This unit has one heater
above the reheat pressure. This means that an extraction comes
from the blade path of the HP turbine before the HP exhaust. This
cycle makes a big improvement in heat rate since in this instance
the feedwater temperature was increased by almost 50°F. However,
most plants use the HP exhaust (cold reheat) as the highest pressure
extraction. Reasons for the lower pressure extractions are initial
cost and potential reliability issues with high-pressure heaters.
The other extraction arrangements are somewhat typical. There
are three extractions from the IP and three from the LP casing. The
extraction flows varying from 5% to 11% of the local blade path
flow. The LP inlet pressure for this design is relatively low compared to most US power plants for this size. This means larger piping from the IP to the LP turbine. However, with longer blades in
the IP back stages relative to the shorter blades in the front stages of
the LP turbine, there is a potential for performance improvement.
The rated heat rate for this unit is quite good at 6946 Btu/kWh.
HP turbines can be either full arc or partial arc of admission
units. If the unit is planned to be operated at part load for a significant period of time, a partial arc design ought to be considered. For
subcritical units sliding pressure operation is an option. Therefore,
justification can be made for either full or partial arc designs with
a control stage. Performance over the load range and load reserve
requirements need to be considered. Reference [32] discusses
many options for operation with full arc, partial arc and full arc
designs with overload valves. Reference [33] discusses comparisons between full arc and partial arc designs. These two references
illustrate that the design of the unit makes a significant difference
in the off design performance. For a partial arc design, it has been
shown that operating with full pressure partial arc admission until
50% admission and then sliding the inlet pressure for further load
reduction is an attractive operating strategy for heat rate improvement and it can also minimize cyclic stresses in the rotor relative
to pure constant pressure operation. One important consideration
is the prediction, years in advance, as to how the unit will be operated. Therefore, operating flexibility ought to be considered.
26.6
OPERATION AND MAINTENANCE
Operational concerns for steam turbines encompass a variety of
issues. Some of these are low cycle and high cycle fatigue, material
creep, vibration, steam chemistry, and water induction. Steam turbines currently in use have a large variety in the types of operating
instructions and control systems to address operation issues. However, many units today employ a digital hydraulic control system
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26-22 • Chapter 26
Fig. 26.33 Startup recommendation
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that permits starting and loading the steam turbine in a controlled
fashion as to adhere to the limits recommended by the supplier.
There are a number of limits that must be adhered to prior to
admitting steam to the turbine. The steam temperature must be
high enough to limit too large of a mismatch between steam and
metal temperatures for items such as steam chests, cylinders, and
rotors. In addition, enough superheat is required to limit internal
condensation. The inlet pressure and temperature recommendations are obtained from the controller or the instructions. Fig­
ure 26.33 shows one such recommendation in graphical form to
limit rotor stress during initial admission of steam and rolling time
to synchronous speed. For units that have bored HP–IP rotors long
heat soak times at less than full speed are required to heat the bore
to temperature near or above the Fracture Appearance Transition
Temperature (FATT). This is done to have sufficient ductility in
the rotor to prevent crack growth of any cracks in the rotor of a size
below or at the ability of inspection techniques to identify. Some
reduced acceleration to full speed is required for a cold rotor even
with today’s rotors.
The startup of a steam turbine is a complicated process of checks
if manually done. That is why controllers are now available that
perform the following functions while controlling turbine component stress, avoiding blade and rotor resonance ranges:
· Starting the turbine from turning gear speed to nominal speed
· Synchronization
· Loading and unloading between zero and rated load
Since more units are cycled today, automated control is more
critical since these units start and shutdown more frequently.
In addition, with automatic control, there are varying operating
strategies that can be implanted to reduce stress. Reference [34]
discusses this topic for a combined cycle plant. If the operation
is not controlled by a controller, the operator would need to follow a set of instructions for loading and unloading, as shown in
Figure 26.34.
Turbine control systems usually use some type of on line computation of rotor stress to control low cycle fatigue damage during startup, shutdowns and load changing. These calculations are
generally addressing the low cycle fatigue limiting location on
the rotor. These locations are either in the blade attachments areas of the HP and IP turbines. The calculations are numerical
but in some instances use mathematical relationships to determine the thermal stress. This approach can factor in the actual
rates of temperature, speed and/or load change, as opposed to the
linearization of the assumed event while using a chart. The control system output can also advise the operator regarding desired
steam conditions for reducing stresses at startup. In addition, the
amount of margin remaining or an extrapolated peak stress can
be provided for either manual or controller action. These can be
provided for a number of components such as HP and IP valves,
and HP and IP rotors.
The boundary temperatures for the models can be obtained using steam temperatures and applying appropriate heat transfer rates
to the surfaces. In some instances appropriate metal temperatures
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ENERGY AND POWER GENERATION HANDBOOK • 26-23
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Fig. 26.34 Load changing recommendation
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measured on casings are applied as the surface temperature for the
rotor. Guidelines for allowable increases in temperature are provided to the operator. In general, the rotor limits the low cycle fatigue of the turbine. However, if a stationary part, such as a steam
chest, needs additional limits, these need not be modeled. Thermocouples installed directly into the part can be used to monitor
component stress. For example, in a steam chest shallow and deep
thermocouples are routinely used to provide either a controller or
operator information relative to the stress limits.
When starting a partial arc design unit that has the capability
to transfer from full arc to partial arc operation, it could be beneficial to the transfer to governor valve control with the unit in
single valve mode. The reason is that once the unit is loaded, the
inlet temperature and the control stage exit temperature will increase rapidly and the rate of change may be more than desired. If
the unit is in single valve mode, transferring to partial arc would
reduce the control stage exit temperature, thus reducing the rotor
stress. Data showing this is situation is shown in Reference [33].
Control systems can help protect the turbine from a variety of
other issues that could affect turbine reliability. The rotor train acceleration rate can be set to provide a more rapid speed increase
and to limit operation in regions of rotor critical speeds or tuned
blade resonances for the tuned LP end blades (usually the last three
stages of the LP turbine).
Condenser pressure limits are in intended to prevent blade damage through high cycle fatigue and to avoid excessive windage
heating. One such limit curve is shown as Figure 26.35. This particular chart is for a combined cycle unit. The LP induction pressure is used as an indicator of last stage flow since there are no
extractions or inductions beyond this location. The horizontal lines
at low flows are the low flow limit line. One limit is for alarm and
the other is for automatic trip. The angled lines that rise from about
18 psia to about 26 psia represent the stall flutter limit for the blade.
Un-stalled flutter is not addressed by operation but by the design of
the blade. These lines are determined by calculation and testing.
Exhaust hood sprays provide some protection for windage heating. This cooling of the blade by an exhaust hood spray is aided
by the recirculating flow that would be present at the base of the
large last row blade during windage heating. Thermocouples in the
exhaust hood provide input to alarm messages or input to activate
hood spray directly.
There are times when feedwater heaters need to be removed
from service. In general, rules are applied on what heaters can be
removed without reducing turbine load. As heaters are removed
from service, flow is increased in the following stages of the turbine. This produces overload potential for the blades which is to
be avoided. Reheater sprays increased above the design can also
produce overloading situations throughout the turbine among other
undesirable issues.
The control of steam purity is an important aspect of turbine
operation. The presence of corrosive impurities in steam can cause
damage to turbine components. The damage is caused by corrosion,
stress corrosion, corrosion fatigue, and erosion-corrosion. Caustics, salts, and acids must be controlled. Deposition of i­mpurities
can also cause thermodynamic losses and distress by lowering the
efficiency of blades and upsetting the pressure distribution. Deposits can reduce the flow passing capability of the blading due to
deposits especially in HP turbines with their relatively small blade
passages. Stress corrosion has been a significant issue in some rotor and blade attachment designs in LP turbines.
Limits for steam chemistry are sometimes provided with different allowables for different stages of operation. For example different limits for normal operation at load and startup. The normal
operating limits are more restrictive. When operating with poorer
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26-24 • Chapter 26
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Fig. 26.35 Exhaust pressure limitations
Te = Ta + AN
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chemistry than ideal, the limit may have time restrictions. When
operating with steam from an auxiliary boiler such as supplying
gland steam during startup, more latitude is generally given for
chemistry limits.
Generally, continuous or routine monitoring of samples taken at
the HP inlet and the Hot Reheat inlet is recommended. When steam
enters, the cycle from multiple pressures from separate steam
sources such as in a combined cycle unit, each source should be
monitored. References [35], [36], [37] and [38] provide more information regarding the issue of steam chemistry in power plants.
Some retrofit designs and new unit applications since the 1990s
have recommended 8- to 10-year inspection intervals. Some of
these inspection intervals may be based on equivalent operating
hours. One definition in a published standard is defined as:,
(26.22)
Please replace
Where:
Te = equivalent operating hours
with "."
Ta = actual operating hours
N = number of total starts (cold, warm and hot)
A = multiplier on the number of starts
","
Reference [39] discusses inspection and overhaul of large steam
turbines. This reference recommends a value of A in Equation 26.22
of between 20 and 30.
Intermediate inspection intervals may also be recommended
with varying degrees of component inspections. For example if minor, medium and major inspections are defined with only a major
inspection requiring total opening of inner and outer casings. Over
a 200,000 hour equivalent operating time frame, inspections could
be recommended as follows
Minor
Medium
Minor
Major
25,000
125,000
50,000
150,000
75,000
175,000
100,000
200,000
Recommendations are provided for inspections to be made on
line in varying time intervals to identify items such as instrumentation errors, drain malfunctions, feedwater heater drain valve and
non-return valve operation.
Special attention to the oil lubricating system is warranted. Some
units use large supplemental cleaning systems that vary from partial to 100% oil flow capability.
The inspection recommendations for individual units and, if applicable, the calculation of equivalent operating hours may be different than the example given above.
The collection and tracking of data on shaft vibration levels
and phase angles is recommended. This is helpful information
to review especially going into a long turbine outage, since it
may highlight the need for additional work that is currently
unplanned.
The potential for water damage is present. Therefore, it is important to review plant design, operating practices and instrumentation against the requirement of Reference [40]. New units
should address these concerns in the initial design. Systems to
detect the potential for water damage are sold as retrofits to existing plants.
26.7
SUMMARY
The state of technology available in the current steam turbines
designs is quite advanced. Advances in Finite Element Analysis,
ENERGY AND POWER GENERATION HANDBOOK • 26-25
13. Zabrecky, J.S., Bezugly, J.A., Brown, M.K., and Martin, H.F.,
“High Power Density 60 Hz Single Flow Steam Turbine with
42 Inch Titanium Last Row Blade for Advanced Combined Cycle
Application”, PWR-VOL. 34, 1999 ASME Joint Power Generation
Conference.
14. Hottel, H.C., and Egbert, R.S., “Radiant Heat Transmission from
Water Vapor”, AIChE Transactions, Vol. 38, 1942.
15. “Cyclic—Duty Turbine Boiler Operating Practices”, Electric Power Research Institute Report EPRI CS-3800, Project 911-1 December 1984.
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16. Brilliant, M., and Tolpadi, A. K., “An Improved Analytical Approach
to Steam Turbine Heat Transfer”, Proceedings of ASME Power 2004,
March 30-April 1, 2004, Baltimore, Maryland.
17. Oeynhausen, H., Drosdziok, A., Ulm, W., and Termuehlen, H.,
“A­dvanced 1000 MW Tandem Compound Reheat Steam Turbine”,
P­resented at 1996 ASME Joint Power Generation Conference.
18. Chen, S., and Martin, H.F., “Blading Design to Improve Perfor­
mance of HP and IP Steam Turbines” PWR-Vol. 30, Proceedings of the
International Joint Power Generation Conference, 1996.
References
1. Moroz, L., and Tarasov, A. “Coupled CFD and Thermal Steady State
Analysis of Steam Turbine Secondary Flow Path,” ASME Paper
JPGC2003-40048, Presented at the 2003 International Joint Power
Generation Conference, June 16-19, Atlanta, GA.
20. Stueer, H., Schmitt, S., and Ashcroft, G., “Aerodynamic Mistuning of
Structurally Coupled Blades” GT2008-50204, Proceedings of ASME
Turbo Expo 2008: Land, Sea and Air, June 9-13, Berlin, Germany.
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2. Steltz, W.G., Evans, D.H., and Stahl, W.F., “The Aerodynamic
Design of High Performance Low Pressure Steam Turbine,” Third
Scientific Conference on Steam Turbines of Great Output, Gdansk,
Poland, Sept. 1974.
19. Simon, V., and Oeynhuasen, H., “3DV Three-Dimensional Blades –
A New Generation of Steam Turbine Blading” PWR-Vol. 33, 1998
ASME International Joint Power Conference Volume 2.
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26.8
sented at the 2000 International Joint Power Generation Conference,
Miami Beach, Florida, July 23–26.
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Computational Fluid Dynamics and Computer Aided Design have
made this possible. Some of the design methods and concepts have
been validated through the application to retrofit designs, as well
as, new units. Very large annulus areas are available on full speed
units. Significant heat rate improvements will require higher pressure and temperature steam conditions.
Extended service intervals are becoming more common for new
and retrofit units due to the increased reliability of the current designs. The use of equivalent operating hours is currently being applied to steam turbines. Automated control systems have aided in
improving reliability and operating flexibility.
This chapter attempts to provide information for use in understanding the current state of technology in steam turbines. Due to
space limitations a great deal of material could not be included.
However, the use of references has attempted to provide resources
for additional information. In addition to references cited in the
text, References [41] and [42] will also provide useful general
information.
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21. Neef, M., Sulda, E., Suerken, N., and Walkenhorst, J., “Design Features and Performance Details of Brush Seals for Turbine Applications”, GT2006-90404, Proceedings of ASME Turbo Expo 2006:
Land, Sea and Air, May 8-11-13, Barcelona, Spain.
3. Marsh, H. “A Computer Program for Through Flow Fluid Mechanics in an Arbitrary Turbomachine in Radial, Axial and Mixed Flow
Types,” Aeronautical Research Council R. and M, No. 3509.
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4. Stueer, H., Truckenmueller, F., Borthwick, D., and Denton, J., “Aerodynamic Concept for Very Large Steam Turbine Last Stage” Paper
GT2005-68746, Proceedings of ASME Turbo Expo 2005, RenoTahoe, Nevada, USA, June 6–9, 2005.
5. Lampart, P., “Numerical Optimization of Stator Blade Sweep and
Lean in an LP Turbine Stage,” ASME Paper IJPGC2002-26161, Proceedings of IJPGC, 2002 International Joint Power Generation Conference, Phoenix, AZ, USA, June 24–26, 2002.
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6. Gray, L., Hofer, D.C., and Takenaga H., “Recent Advances in the
Prediction of Exhaust Losses for Low Pressure Steam Turbines with
Downward Exhaust Hoods,” Advances in Steam Turbine Technology
for Power Generation, PWR Vol. 26. ASME 1994.
7. Leyzerovich, A., “Wet-Steam Turbines for Nuclear Power Plants”
Published by PennWell.
8. Heyman, F.J., “Liquid Impingement Erosion,” ASM Handbook, Volume 18, “Friction, Lubrication and Wear Technology” ASM International, 1992.
9. Meyer, C.A. and Lowrie, J.A., “The Leakage Through Straight and
Slant Labyrinths and Honey Comb Seals,” ASME Journal of Engineering for Power, Vol. 97, pp. 495–502.
10. Young, W.C., “Roark’s Formulas for Stress and Strain,” 6th Edition,
Published by McGraw Hill, Inc.
11 “Evaluating and Improving Steam Turbine Performance”, Second
Edition, by Cotton, K.C.
12. Magoshi, R., Nakano, T., Konishi, T., Shige, T., and Kondo, Y.
“Development and Operating Experience of Welded Rotors for High
Temperature Steam Turbines”, ASME paper IJPGC200-15007, Pre-
22. Stephen, D., and Hogg, S. J., “Development of Brush Seal Technology for Steam Turbine Retrofit Applications.” IJGPC2003-40103,
Proceeding of the IJGPC2003 International Joint Power, June 16-19,
Atlanta, Georgia.
23. Little, N., Sulda, R., and Terezakis, T. “Efficiency Improvement on
Large Mechanical Drive Steam Turbines”, Proceedings of the 30th
Turbomachinery Symposium.
24. Sumer, W.J., Vogan, J.H. and Lindinger R.J., “Reducing Solid Particle Erosion Damage in Large Steam Turbines”, Proceedings of American Power Conference, Vol 47, 1985 pp 196–212.
25. Reinhard, K.G., “Turbine Damage by Solid Particle Erosion” ASME
Publication 76-JPGC-PWR-15, 1976.
26. Sapate, S.G., and Roma Rao, A.V., “Effect of Erodent Particle Hardness on Velocity Exponent in Erosion of Steels and Cast Irons”, Materials and Manufacturing Processes, Volume 18, Issue 5, January 2003,
pages 783–802.
27. Kramer, L.D., Quereshi, J.I., Rousseau, R.A., and Ortolano, R.J.,
“Improvement of Steam Turbine Hard Particle Eroded Nozzles Using Metallurgical Coatings”, ASME Publication 83-JPGC-Pwr-29,
1983.
28. Tran, M.H., “Aerodynamic Design to Minimize Solid Particle Erosion on Control Stage Blading”, EPRI Conference on Solid Particle
Erosion, March 1989.
29. Silvestri, G.J., Bannister, R.L., Fujikawa, T., and Hizume, A. “Optimization of Advanced Steam Condition Power Plants”, Transaction of
the ASME Vol. 144, October 1992, pg. 612–620.
30. ASME PTC 6 - 2004 (Revision of ASME PTC6-1996)
31 ASME PTC 6.2 – 2004
26-26 • Chapter 26
34. Ulbrich, A., Gobrecht, E., Siegel, M. R., Schmid, E., and Armitage,
P. K., “High Steam Turbine Operating Flexibility Coupled With
Service Interval Optimization”, ASME Paper IJPGC2003-40072,
Presented at the International Joint Power Generation Conference
June 16–19, 2003, Atlanta, Georgia.
35. “The ASME Handbook on Water Technology for Thermal Power
Systems”, Paul Cohen, ed., ASME, 1989.
38. “Proceedings: Fifth International Conference on Fossil Plant Cycle
Chemistry”, EPRI, Palo Alto, CA. 1997. TR-108459.
39. “Recommendations for Inspection and Overhaul of Steam Turbines”,
VGB-R115Me, Published by VGB Technical Association of Large
Power Plant Operators”
40. “Recommended Practices for the Prevention of Water Damage to
Steam Turbines Used for Electric Power Generation”, ASME TDP1-2006 (Revision of ASME TDP-1-1998).
41. “Marks Standard Handbook for Mechanical Engineers”, Eighth Edition, McGraw Hill Book Company.
42. Sanders, W. P., “Turbine Steam Path”, Published by PennWell.
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36. “Proceedings: Eighth International Conference on Cycle Chemistry in
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37. “Proceedings: Sixth International Conference on Fossil Plant Cycle
Chemistry”, EPRI, Palo Alto, CA. 2001. 1001363.
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33. Silvestri, G.J., and Martin, H.F., “An Update On Partial-Arc Admission Turbines For Cycling Applications”, Presented at the Electric
Power Research Institute 1985 Fossil Plant Cycling Workshop, Nov­
ember 5–7, 1985, Miami Beach, FL.
ators”, June 20-22, 2006, Calgary, Alberta Canada. EPRI, Palo Alto,
CA: 2007. 1014831.
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32. Termuehlen, H., “Variable Pressure Operation and External Turbine
Bypass Systems to Improve Plant Cycling Performance”, Presented
at the Joint ASME/IEEE/ASCE Power Conference, Charlotte, N. C.
October 9-10, 1979.