Proceedings of the Second Middle East Turbomachinery Symposium

Proceedings of the Second Middle East Turbomachinery Symposium
Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
Modernization of Steam Turbine Diaphragms for the
Saudi Aramco Gas Plants
Mr. Samuel L. Golinkin
Siemens Demag Delaval Turbomachinery, Inc.
Hamilton, NJ, USA
Mr. James E. Luker
Siemens Demag Delaval Turbomachinery, Inc.
Hamilton, NJ, USA
Mr. Michael J. Lipski
Siemens Demag Delaval Turbomachinery, Inc.
Hamilton, NJ, USA
Mr. Riyadh A. Al-Jaafari
Saudi Arabian Oil Company,
Udhailiyah, KSA
Samuel Golinkin is a
Fellow Engineer with
Siemens Demag Delaval
Turbomachinery, Inc. in
Hamilton, New Jersey,
USA. His professional
career has extended over
50 years across three
countries (former USSR,
Austria and the USA) and
covers many aspects of the steam
turbomachinery field, including design,
manufacturing, troubleshooting and consulting.
His expertise includes both new turbines and
upgrades/repairs as well (up to 300 MW). Since
joining Delaval in 1991, Mr. Golinkin was
actively involved in different aspects of
mechanical design for new units and upgrades.
He has proposed and implemented numerous
design and manufacturing improvements, and
created new standards and procedures which
have resulted in efficiency and reliability
increase. He is the author/coauthor of 20
technical articles and holds 10 patents (and has
two patents pending), related to steam turbine
technology.
Mr. Golinkin has a B.S. degree and a M.S.
degree (summa cum laude) in Mechanical
Engineering from Kharkov Polytechnic Institute
in the Ukraine
Mr. Luker received his B.S. degree in
Mechanical Engineering (2009) from Drexel
University and is currently working toward his
M.S. degree from Rutgers University.
Michael J. Lipski is the
Manager of Repair and
Revamp Engineering at
Siemens Demag Delaval
Turbomachinery, Inc., in
Hamilton, New Jersey, USA.
He has 39 years of
experience in the
turbomachinery
industry. Mr. Lipski is extensively involved in
development, design, manufacturing issues, and
field troubleshooting. He has previously worked
as a Compressor Design Engineer, and his
responsibilities included the development, design
and manufacturing of major components for
centrifugal compressors. He is the
author/coauthor of 3 technical articles and
currently holds one U.S. Patent and has two
patents pending.
Mr. Lipski has a B.S. degree (Mechanical
Engineering Technology) from Trenton State
College.
ABSTRACT
James Luker is a
Mechanical Engineer with
Siemens Demag Delaval
Turbomchinery, Inc. in
Hamilton, New Jersey, USA.
He joined Siemens in 2007,
and has been involved
with steam turbine upgrades
and repairs, related R&D
projects and troubleshooting.
The critical components of the Saudi Aramco
Uthmaniyah and Shedgum gas plants are eight
trains of turbomachinery, 4 trains per plant. Each
train consists of a steam turbine and a centrifugal
compressor. At the time of their design and
manufacture, these steam turbines were at their
chronological level of technology. They operated
reliably and effectively for their original design.
However, during their 35+ years of operation,
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
some of the diaphragms developed problems as
the gas plants production gradually increased.
With increased gas production, the mechanical
strength of the heavily loaded diaphragms
became marginal. As a result of the increased
loading, these diaphragms deflected more than
they did originally and caused axial rubs with
serious diaphragm and rotor damage.
Fresh steam enters the steam chest through five
valves, then expands along the steam path and
leaves the turbine through an upper exhaust.
Over the years as the operating conditions
changed, Saudi Aramco personnel made several
requests for the manufacture of “spare”
diaphragms. Even though the operating
conditions changed, the basic Saudi Aramco
requirements did not. Specifically:
 The diaphragms had to have adequate
increased mechanical strength to withstand
the increased loads.
Figure 1 - Cross-Section of the Original Delaval
KJ-MV-NC Steam Turbine
The turbine’s steam path consists of five Rateau
stages. The first (control) stage has a partial arc
(approximately 50 percent) steam admission
while the remaining four stages operate with full
(100 percent) admission.
 They had to be more efficient.
 They had to be a “drop-in” design which
could be installed into the existing turbine
without any rework of the case and rotor.
The nozzle ring and four diaphragms are of
standard welded design. They were fabricated
and machined using the manufacturer’s proven
procedures. The nozzle ring is attached to the
case steam chest by special wedges on its outer
diameter (OD), and by bolting on its inner
diameter (ID). Four diaphragms are installed in
the case. Their assembly into the case will be
described later.
At each request, diaphragm design changes were
made to produce diaphragms that met Saudi
Aramco’s requirements. Each design step
exemplified the latest technology available at the
times of their design and manufacture.
This paper describes the design and
manufacturing evolution leading to the current
diaphragm design. The advanced analytical tools
(FEA for defining stresses and deflections) were
used for evaluating the mechanical integrity of
the various diaphragm designs. Design
improvements (vane profile, tangential vane lean,
tip and new shaft seals), and manufacturing
improvements (electron beam welding, laser and
water jet cutting, electric discharge machine wire
cutting), which have led to substantial gains in
mechanical strength, efficiency and reliability
will be described in detail.
The entire rotor is made from a solid forging and,
together with its rotating blades, is a rugged
turbine component. Rotating blades are fixed on
the rotor by axial fir-tree fasteners which mate
with the broached axial grooves in each of the
five discs. In each row, the tips of the blades are
interconnected. In the first stage the blade tips
are connected by a combination of integral and
loose shrouds. In the remaining four stages the
blade tips are connected by loose shrouds only.
BRIEF HISTORY
The rotor is supported at each end by a tilt-pad
journal bearing. A single, equalizing, tilt-pad
thrust bearing locates the rotor axially.
Steam Turbine Design
Each impulse steam turbine located at Saudi
Aramco’s Uthmaniyah and Shedgum gas plants
is, per design, a multi-stage and multi-valve,
non-condensing unit with a power output of
32239 HP at a rotor speed of 5713 RPM. The
turbine operates with inlet steam conditions of
4136 kPA, 371 °C, and a backpressure of 550
kPa, (600 PSIG, 700 °F and a backpressure of 80
PSIG) (Figure 1).
Interstage shaft seals and both end seals, front
and exhaust, are also of the manufacturer’s
standard design. Each seal ring is made from
leaded-nickel bronze and consists of four springbacked segments.
All things considered, in the mid 1970’s these
steam turbines were at the leading edge of
turbomachinery design and manufacturing
technology.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
Operational History
Two Saudi Aramco gas plants, Uthmaniyah and
Shedgum, form the largest integrated facility in
the world. This facility went into commercial
operation in 1980 and has gradually increased
production (to NGL Fractionation plants and
Sales gas to the power industry) over 30+ years.
The subject eight turbomachinery trains are a
critical component of the facility, since all LP
sweet gas is compressed through them.
Shortly after the start of the production (mid
1980s) the nozzle rings in several turbines were
found to have broken or damaged exit edges.
While the exact cause of the damage was unclear
until much later, the manufacturer addressed the
problem as described in the “Optimal Edge-toEdge Clearance” section, and the steam turbines
operated without notable incidents for more than
10 years.
Figure 2 - Severe damage of Row #4 Blades
In the late 1990’s inspections showed some
damage and wear to the stage 3, 4, and 5
diaphragms. Therefore spare diaphragms of the
original design were purchased, manufactured,
and installed.
From 2001 to 2005 it became apparent that the
original diaphragm design was no longer
adequate for the plants increased output.
Turbomachinery technology had also changed
during the intervening years, and the
manufacturer was able to offer modified
diaphragms that improved the mechanical
strength and efficiency. These modified
diaphragms were purchased and some of them
were put into operation.
Figure 3- Severe Disc #4 Damage: Rubbed
Metal Filled Equalizing Holes
In 2005 through 2009 Saudi Aramco experienced
diaphragm and rotor failures in four steam
turbines. It was clear the damage occurred due to
an axial rotor rub against the diaphragms (Figure
2 to Figure 4). The damage at stages 4 and 5 was
especially severe.
Figure 4- Complete Disintegration of Stage #4
Original Diaphragm
In another turbine with an undamaged rotor, the
vanes at the horizontal joint of the stage 5
diaphragm were loose enough that they could be
moved by hand (Figure 5). Although the
damaged rotors were repaired and placed back
into operation, it was established that both the
original diaphragm design and the modified
diaphragm design no longer had sufficient
mechanical strength for the current plant
operating conditions. This was confirmed with
Finite Element Analysis (FEA) provided by the
manufacturer.
Working with the manufacturer, it was
determined that it was possible to substantially
improve the reliability and performance of the
entire turbine steam path. However, optimizing
the steam turbines did not make financial sense
unless the entire turbine-compressor train was
likewise upgraded and coordinated with the
plans of the gas plant’s development. Therefore
Saudi Aramco postponed the entire turbine
upgrade, and requested that the manufacturer
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
create a new generation of diaphragms with
maximum mechanical strength, reliability, and
DIAPHRAGMS: GENERAL FEATURES
AND REQUIREMENTS
Main Requirements
Diaphragms are the major stationary component
of a steam turbine. Diaphragms provide the
following important functions (Bistritzkiy, 1956):
 Transform potential steam energy of the
steam into kinetic energy of the powerful
steam jet (Figure 7).
 Direct the steam jet into the rotating blades in
the most effective manner.
Figure 5 - Unreliable Stage #5 Original
Diaphragm: Vanes at Joint can be Shaken by
Hand
 Establish the pressure drop per stage and
withstand heavy pressure loads at high
temperature, especially in the first few stages
of the typical steam turbine.
efficiency for the existing steam turbines which
would be installed in a “drop-in” manner and
could be used with the existing rotors and cases.
The requested generation of diaphragms would
be designed for current operating conditions and
incorporate the latest developments in diaphragm
design and manufacturing technologies. These
new diaphragms have been manufactured,
installed, and are successfully operating in five
steam turbines (Figure 6). Saudi Aramco plans to
install these new diaphragms in their remaining
steam turbines.
 Ensure minimum deflections under heavy
loads in order to prevent any rubs.
 Contain all the stage seals (tip, root, and shaft
seals), which minimize steam leakages
beyond the main steam path and thus
improve stage efficiency (Figure 8).
 Divide the interior steam turbine space into
separate compartments for each rotating
wheel of the turbine rotor. A separate “room”
for each rotating stage minimizes damage if a
failure occurs (Figure 9).
Figure 6 - New Current Diaphragms in the
Turbine Case
Figure 7 - Steam Velocities Through an Impulse
Turbine
Basically, the joint Aramco-Manufacturer
experience in design, production, and operation
of several diaphragm generations for the subject
turbines, reflects the entire evolution of
diaphragm design and manufacturing over the
last 40 years. Below is a brief description of the
major steps of this remarkable evolution.
Figure 8 - Diaphragm: A Housing of the Stage
Seals
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
Figure 10 - Diaphragm: Overall View Of
Assembled Diaphragm
Figure 9 - Diaphragms Assembly in Case: Axial
Positioning
A typical welded diaphragm assembly consists
of three main parts (Figure 10): the outer ring,
the “squirrel cage” (or cage), and the inner ring
(or center). The “squirrel cage” is made of two
concentric thin rings known as the outer and
inner spacing strips, with the vanes between the
rings. Each spacing strip contains through-holes.
Their quantity is equal to the number of
diaphragm vanes, and they are located evenly
around the spacing strip circumference. Each
vane is inserted through these profile holes in the
outer and inner spacing strips and is affixed to
them by welding. The completed “squirrel cage”
is assembled with the outer and inner rings of the
diaphragm and welded together by four
significant circumferential welds. After welding
is complete, each diaphragm must be stressrelieved and final machined.
General Design Features
Each turbine diaphragm is a complicated 360
degree plate, which is composed of three major
components made from different materials. The
outer and inner rings are reliably connected
together with a set of vanes having complex
geometry. While vanes, operating in the main
steam flow, are made from special stainless steel,
both rings can be made from such metals as:
carbon steel, carbon-moly or chrome-moly alloys,
and even cast iron - depending upon steam
conditions (pressure and temperature) and loads
(Bistritzkiy, 1956).
To accommodate installation of the rotor, the
diaphragm is split into equal halves creating a
horizontal joint. The outer ring supports the
diaphragm in the case while the inner ring hangs
free. The solid structure can be achieved by one
of the following manufacturing methods:
The diaphragm assembly into the steam turbine
case is arranged as follows: the tongue on the
diaphragm’s outer diameter (OD) is inserted into
a mating groove in the case; this fixes the
diaphragm in the axial direction. Each diaphragm
is secured in the horizontal and vertical direction
by using a set of perpendicular special keys, also
known as the “thermal key-cross” (Figure 11).
 Welding
 Casting
 Pinning
The lower half of a diaphragm is suspended in
the case lower half at the horizontal joint by two
special supports. These supports are located on
the left and right side of the diaphragm and
provide vertical alignment of the diaphragm. The
horizontal alignment is accomplished by a
centering key or pin located at 6 o’clock position.
The “thermal key-cross” design provides the
proper diaphragm position during operation,
protecting from the inevitable thermal and
mechanical case deformations.
 Machining from a solid plate
Welding is the most common, reliable, and cost
effective method of diaphragm fabrication.
Casting a diaphragm, consisting of cast iron
rings and stainless steel vanes, is inaccurate and
does not have sufficient strength. The two
remaining methods provide accuracy and
sufficient strength, but are very expensive.
Therefore, these three fabrication methods are
not used widely in industry and will not be
discussed in this paper.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
AIRFOIL DESIGN
Figure 12 shows a comparison of the original
and the current airfoil design. The adjacent
original airfoils form a channel that accelerates
steam flow starting from the entrance; and only
after the steam travels almost through the entire
channel is it turned to the proper exit angle (or
direction). Inevitably, turning the high velocity
steam flow causes intensive steam separation
resulting in substantial profile and secondary
losses.
Figure 11 - Diaphragm Alignment in Case by
Means of the “Thermal Key-Cross”
PROGRESS IN DIAPHRAGM DESIGN
Extensive R&D, using achievements in
electronics, computation, and modeling during
the last 20+ years, has yielded a new generation
of advanced design tools such as Computational
Fluid Dynamics, (CFD) (Deckers et al. 1997,
Dawes 1988, Dawes 1990, Holmes and Tong
1985, Turner et al. 1993) and Finite Element
Analysis, (FEA) (Moaveni 2008). Use of these
calculation methods along with cascading test
data and rotating telemetry testing, resulted in a
greater fundamental understanding of the
aerodynamic and thermodynamic processes
inside a steam turbine. This new knowledge
made it possible to generate significant design
improvements in the major steam turbine
components (Paul et al. 1989, Shcheglyaev 1993,
Troyanovsky 1993, Cofer 1996).
Figure 12 - Evolution in Vane Airfoil
The current airfoil design acts exactly opposite
the original design. Adjacent current airfoils
form a channel which first turns the entering
steam flow to the proper direction and only then
accelerates it. Since the steam flow velocity
during its turn is relatively low, the overall losses
are much smaller than those in the original
design. Additionally, the current airfoil has the
improved configuration (geometry) of the inlet
edge, which allows it to accept steam flow at
different angles without significant separation.
This widens the range of effective operation.
Also the new vane airfoil is significantly
stronger than the original one which increases
the mechanical strength and reliability of the
entire diaphragm.
In the area of diaphragm design, there are the
following improvements, which resulted in a
substantial increase in diaphragm reliability and
efficiency:
 New vane airfoils
 Optimal tangential lean of the vanes
 Contoured endwalls in the HP diaphragms
References:
 Optimal edge to edge clearance between
vanes and rotating blades
Paul et al. 1989, Shcheglyaev 1993, Goel et al.
1993, Shelton et al. 1993, Dejch et al. 1994,
Cofer 1996
 Advanced stage seals
Below is a brief description of these proven
design achievements implemented in the UGP
and SGP steam turbine diaphragms.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
them. This is especially detrimental for designs
without radial tip seals. Concurrent with the area
of overload, some area at the root of the rotating
blades experiences flow starvation. Such uneven
steam flow distribution along the rotating blades
seriously affects their efficiency and reliability.
OPTIMAL TANGENTIAL LEAN OF THE
VANES
Maximum utilization of the available kinetic
energy of the steam flow depends, among other
factors, upon the smooth, unimpeded entrance of
the steam jet into the rotating blades. This
smooth entrance can only be achieved by
directing the steam jet at a certain compound
angle.
In the period between the 1970’s and 1990’s the
continuous design and technological progress,
which is described later in this paper, allowed the
improvement of the vane orientation from a
negative lean to a radial orientation of the vane
exit edges. As seen in Figure 13b, this change
resulted in a better steam flow distribution along
the rotating blades’ height, which improved the
efficiency and reliability. However, it did not
completely resolve the problem.
The axial angle component of this compound
angle was defined long ago. The first ideas
regarding the radial angle component and the
importance of proper tangential vane orientation
(lean), defining this radial angle, were published
in the 1950’s (Shcheglyaev 1993, Filippov and
Zhong-Chi Wand 1964). However, the existing
limitations in theoretical knowledge, tools, and
technology did not allow these ideas to develop
and be implemented into everyday practice.
Since the 1990’s further design and
manufacturing improvements made it possible to
develop the optimal tangential vane orientation
with positive lean of the exit edges. Such vane
positioning almost provides an even steam flow
distribution along the entire height of the rotating
blades in the high and intermediate pressure
stages (Figure 13c). However, it is not enough
for low pressure stages with long, rotating blades.
Only the latest 3-D design of the vane and blade
airfoil areas, (not shown here), resolves the
problem for these low pressure stages (Jansen
and Ulm 1995, Oeynhausen et al. 1996). Figure
14 shows another view of the different tangential
vane leans and their impact on stage efficiency,
based upon the tests made in the former USSR
and Japanese R&D (Kawagishi et al. 1991).
Figure 13 shows the possible tangential leans of
the vanes, the orientation of their exit edges and
the schematic direction of the steam flow
provided by the following leans:
a.
Negative lean
b.
Radial orientation (no lean)
c.
Positive lean
Figure 13 - Tangential Lean of the Vanes
Up until the 1970’s, due to manufacturing
limitations, the vanes were oriented with
negative lean. As evident from Figure 13a, vanes
with negative lean cannot effectively distribute
the steam flow along the full height of the
rotating blades. The peripheral area of the
rotating blades is overloaded. Moreover, a
significant portion of the steam flow is directed
above the rotating blades, totally by-passing
Figure 14 - Relative Stage Efficiency with
Different Tangential Lean of the Vanes
Figure 15 shows actual diaphragms with
negative and positive vane leans.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
Figure 15 - Diaphragms with Different
Tangential Lean of the Vanes
References:
Filippov et al. 1964, Troyanovsky 1991,
Kawagishi et al. 1991, Shcheglyaev 1993,
Troyanovsky 1993, Sakamoto et al. 1993,
Tanuma et al. 1995, Singh et al. 1995, Jansen et
al. 1995, Oeynhausen et al. 1996
CONTOURED END WALL
The first two to four stages of the steam turbine
operate with high pressure, high temperature,
and high density steam. The small volumetric
steam flow in these stages requires partial arc
steam admission and short vanes and rotating
blades. The large pressure drops across the
nozzle ring and diaphragms impose high loads,
which dictate the use of special vane profiles
with large nozzle ring/diaphragm axial widths.
Therefore the stationary vanes and the rotor
blading on the first few stages have a low aspect
ratio, (i.e. small airfoil heights and large axial
widths).
Figure 16 - Contoured End Wall Test Data
Steam flow at these stages moves through a long,
narrow passage, which results in increased
profile and secondary losses. The end wall
boundary layer in these stages plays a much
more negative role compared with the tall
blading because it:
Starting in the 1980’s, numerous studies on this
issue resumed in different countries, and they
continue to this day (McIntosh 2011). These
studies confirmed that by special profiling of the
nozzle ring and diaphragm outer wall (i.e. the
inner diameter of the outer spacing strip), it is
possible to significantly reduce the profile and
secondary losses. The benefits of this profiling
are:
 Significantly decreases the steam passing
area.
 It actively interacts with the main steam flow
forming an additional radial flow, which
causes disturbances and energy losses.
 Substantial reduction of the outer boundary
layer, especially near the exit edge of the
airfoil area.
Similar to the tangential lean of the vanes, the
idea about special profiling the outer wall of the
nozzle ring and diaphragm with a low aspect
ratio was published in the 1950’s (Shcheglyaev
1993, Dejch et al. 1960). Again, this idea could
not be materialized due to manufacturing
limitations.
 Reduction of eddies and other disturbances in
the main steam flow.
 Funneling of the main steam flow towards
the inner cylindrical wall, (i.e. the outer
diameter of the inner spacing strip).
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
 Minimization of steam leakage beyond the
designed steam flow path (above rotating
blade shrouds and below platforms).
So, the contouring of the outer end wall of the
nozzle ring and diaphragms results in an increase
in stage efficiency (Figure 16) and reliability of
the stages with the low blading aspect ratio.
 Minimization of mixture between the main
steam flow and leakage through the shaft
seals.
References:
Dejch et al. 1960, Sieverding 1975, Tran 1986,
Atkins 1987, Denton 1987, Shcheglyaev 1993,
Troyanovsky 1993, Cofer 1996, Leyzerovich
1997, McIntosh et al. 2011
The major drawbacks are as follows:
 Minimal vane setback leaves the vane exit
edges susceptible to damage during
manufacturing (sub-arc welding and final
machining) and repair.
OPTIMAL EDGE-TO-EDGE CLEARANCE
(EEC)
 The high velocity steam flow, in conjunction
with limited axial spacing between the vanes
and blades, can result in severe damage to
both components (Figure 18), reducing
reliability and efficiency.
EEC is the axial distance between the vane exit
edge and rotating blade inlet edge of each stage
(Figure 17).
As it was mentioned earlier, damage of the exit
edges of the nozzle ring vanes in the first
Aramco turbines occurred at the very beginning
of their operation. It was evident that the vane
failures occurred due to intensive excitation
forces. Without knowing the exact nature of
these forces, the problem was resolved
by redesigning the nozzle ring. The strength of
the vanes was substantially increased by
changing the airfoil and by connecting all the
vanes together by a circumferential rib located in
the middle of their height. At the same time,
similar nozzle ring damage in Russian high
pressure turbines with power capacities 25100MW was resolved by substantial increase
(approximately 1.4 times) of the control
stage EEC - without changing the nozzle ring
design.
Figure 17 - Axial Clearance/ Vane Setback in a
Turbine Stage
EEC is an important factor that affects stage
performance (both efficiency and reliability).
The required EEC for optimal performance is
dependent on whether or not the stage has
effective tip and root seals.
For a stage with effective tip and root seals,
leakage from the main steam path is kept to a
minimum, allowing the EEC to be increased.
With a higher EEC, steam flow becomes more
laminar along the pitches and blade length
because irregularities in the steam flow (eddies,
wakes, etc.) have more axial space to diminish.
This produces the following benefits:
For a stage without effective seals, steam
leakages beyond the rotating blades are a
decisive factor of performance. Tests of a stage
with 35 mm long blades showed that an increase
in EEC from 0.8 mm to 2.5 mm resulted in an
efficiency drop by more than 5 percent. For this
reason, up to the 1970’s, most turbine stages had
small EEC. To achieve a small EEC, minimal
vane setback (“vs” in Figure 17) was
implemented (.254-.762 mm) or (.010-.030”).
 Improved efficiency due to better steam path
utilization.
 Improved reliability due to decreased
excitation forces on the vanes and blades.
 Increased vane setback eliminates
manufacturing damage of the vane exit edge
and associated vane repair.
The main benefits of small EEC on a stage
without effective tip and root seals are
summarized as follows:
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
INTERSTAGE SEALS
In a steam turbine, inter-stage steam leakage
bypasses or invades the main steam flow through
the steam path (i.e.: vanes and rotating blades).
Steam leakage is a significant source of
performance loss due to:
 The missing energy of the steam which
leaves or bypasses the main flow.
 The flow path disturbances (wakes, eddies,
etc.) caused when leaked steam enters and
collides with the main steam flow, affecting
both efficiency and reliability.
Figure 18 - Damaged Vanes Due to Insufficient
Edge-to-Edge Clearances
Figure 20 shows the major causes of
deterioration, or reduction, of the turbine stage
performance based upon the research performed
in the former USSR and in the USA (“TurboCare
Retractable Packing” 1999).
In summary, the magnitude of EEC required for
optimal performance is dependent on other
design features of the stage, namely, the use of
tip and root seals. For stages with no effective tip
and root seals (most vintage turbines), it is
beneficial to maintain a smaller EEC for the best
performance, at the cost of reliability. For stages
with effective tip and root seals (used on most
modern turbines), EEC does not have to be used
as a method to minimize steam leakage. The
designer is able to increase the EEC, thereby
reducing turbulent flow and excitation forces
(Figure 19).
Figure 20 - Major Causes of Reduction in Steam
Turbine Efficiency & Performance
Despite the different timing in data acquisition,
differences in turbine design, power output and
operational modes, both pie charts illustrate the
same conclusions: steam leakage beyond the
blade shrouds and roots, and between the shaft
and the diaphragms is the main source (76.5
percent to 81.0 percent) of turbine performance
deterioration.
Figure 19 - Dependence of the NPF Steam
Stimulus upon Relative Edge-to-Edge Clearance
References:
Lakshminarayana et al. 1973, Traupel 1977,
Langston 1980, Shcheglyaev 1993, Lazzeri et al.
1996
Minimizing steam leakages and keeping them
stable during long- term operation is the most
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
effective way to achieve and maintain higher
steam turbine efficiency and reliability.
 Deviation of steam flow because of
centrifugal effect of rotary blades.
In a typical impulse stage, there are three routes
for steam leaks through the gaps between stator
and rotor components:
Similarly to all other steam turbine components,
tip seals went through a long evolution in design,
from the first primitive axial rigid seals to
current complicated multi-fin or honeycomb
designs.
 Between the diaphragm ID and shaft OD – a
so called, “shaft steam leak”.
Below is a description of the major known tip
seal designs: axial rigid seal, radial rigid seals,
radial flexible (spring-backed) seal, and multi-fin
seals. The special anti-vibration tip seal design is
described in a separate section titled “Interstage
Seals and Low Frequency Vibrations”.
 Over the outer diameter of the rotating wheel,
above the blades – a “tip steam leak”.
 Near the root of the rotating blades through
axial space between diaphragm and rotating
wheel – a “root steam leak”.
The amount of steam leaks and the possible
direction of their flows can be quite different.
The specific leak flow pattern depends upon
multiple factors, including the characteristics of
the main flow, the seal design(s), the
configuration of the axial space between the
diaphragm and the rotating wheel, the number of
pressure-balance (equalizing) holes in the disk,
their diameter and location, etc.
Axial Rigid Seal
Accordingly, in order to minimize and contain
these leaks, the modern turbine stage
incorporates three types of seals: shaft seals, tip
seals and root seals. (Figure 21)
Figure 22 - Axial Rigid Seal
This primitive seal design is a fin, formed on the
inlet edge of the rotating blades shroud which
works against a mating flat face of the
diaphragm. Figure 22 shows the blading and
diaphragm which are both made from hard metal.
The axial clearance “a” between the shroud fin
and the flat mating face controls steam leakage
to some degree, especially in the new condition.
However, the axial clearance in this type of seal
has to be large in order to prevent rubbing which
results in damage to the blading. During turbine
operation this clearance increases due to rotor
thermal expansion, thus significantly decreasing
its sealing ability. Moreover, as the rotor thermal
expansion changes, depending upon the
operational regime, the axial clearance also
changes making the performance of the entire tip
seal unstable.
Figure 21 – Interstage Seals in Modern Impulse
Turbine Stage
This goal can be realized by improving the
interstage seals.
TIP SEALS
The primary function of tip seals is to prevent
steam leak from the main steam flow into the
space above the rotating blades. Among the stage
seals, the leak through the tip seals is the largest
source of efficiency loss due to:
Therefore, today this design is used by some
turbomachinery manufacturers only as an
additional component to the other more
advanced tip seal designs.
 Largest leak area.
 Highest reaction in this area.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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Radial Rigid Seal
this seal is much more stable compared to the
axial rigid seal. Another positive feature of the
segmented plate seal is that this plate can be
made from different materials (special antifriction bronze, Ni-resist, or stainless steel –
depending mainly upon steam temperature).
However, the thin sheet is usually made from
stainless steel. Also, radial rigid seals can be
easily replaced during turbine overhauls.
This seal is formed by a thin (approximately
2.54mm or 0.100” thick) segmented plate
(Figure 23) or by a very thin (approximately
1.0mm or 0.040” thick) sheet (Figure 24)
installed into a radial groove.
In spite of the previously described positive
features, these seals have the following
substantial drawbacks:
 They (especially the sheet metal design) are
vulnerable to mechanical damage that may
occur during turbine operation (e.g. foreign
object damage due to steam impurity) and
from handling during maintenance/repairs,
overhauls and transportation.
Figure 23 - Radial Rigid Seal (Segmented Plate)
 Being rigidly fixed in the diaphragm, these
seals are subjected to severe radial rubs when
rotor radial displacements are larger than the
radial clearance. This situation takes place:
o During start-ups and shut-downs (when
rotor passes through first critical speed).
o During sharp dynamic regimes – “cycling”,
(changes in power output steam conditions,
etc.) leading to sudden significant
misalignments in the steam path.
Figure 24 - Radial Rigid Seal (Thin Sheet)
This radial groove is located either in the
diaphragm outer ring (if one fin is required), or
in the special tip ring, attached to the outer ring
by bolting or welding (if several fins are
required). Each radial groove is oriented
perpendicular to the shaft axis. Segments of the
plate are staked in the groove, while the sheet
metal is fixed in the groove by a caulking wire.
The geometry of the plate or sheet metal at the
inner diameter provides a thin fin which is
engaged with a cylindrical surface of the rotating
blades shrouding. The radial clearance “b”
between the thin fin and blades shroud controls
the steam leakage.
During rubbing, the seal surfaces (the thin fin
and blades shrouds) wear down (thus increasing
the radial clearance). In some cases, the rigid
seal may cut a groove in the shroud which
requires replacement of the entire row of rotating
blades (Figure 25).
Since the radial clearance remains relatively
unchanged * during different operational regimes,
*
In reality, during operation the radial clearances
also change their value from the original (cold)
condition, because of inevitable misalignments
between rotor and stator, as it is described below
in the “Interstage Seals and Low-Frequency
Vibration” section. However, these changes are
much smaller compared to the axial clearance
changes in the axial rigid seals.
Figure 25 – Shroud Damage Due to Heavy Tip
Seal Rub
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Proceedings of the Second Middle East Turbomachinery Symposium
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Radial Flexible (Spring-backed) Seal
to foreign object damage or mishandling. It is
also maintenance friendly and can be easily
replaced during brief maintenance or overhaul.
However, this type of tip seal is more expensive
and requires more space in axial and radial
directions compared to radial rigid seals.
Because of the positive features described above,
this seal can provide the effective leak control
while increasing longevity of the seal ring and
preventing rotor damage. These tip seals are
implemented in the current Saudi Aramco
diaphragms.
Multi-Fin Tip Seals
Figure 26 - Radial Flexible (Spring-backed) Seal
This seal is formed by a segmented ring,
installed into a diaphragm radial groove with
complex geometry (Figure 26). The location of
this groove and its orientation is the same as the
radial rigid seal. The outer portion of this ring
has a special geometry which allows its reliable
positioning in the groove. The ring ID has one or
two shaped thin fins engaged with the cylindrical
surface of the rotating blades shroud.
Similarly to the previous design, the radial
clearance “b” between the thin fin(s) and the
shroud controls the steam leakage. Since this
clearance remains relatively unchanged (see
previous note *) during different regimes, this
seal also remains stable during turbine operation.
Figure 27 - Radial Multi-fin Tip Seal (Solid and
Staked Design)
Each seal ring consists of several independent
segments (usually between 8 and 12-depending
upon outer diameter of the rotating wheel). Each
segment is spring backed by flat or coiled
springs, located on its OD. These springs fix
each segment in the proper position in the groove
providing the designed radial clearance between
the fins during turbine assembly. During
operation, steam pressure over the seal OD,
reinforces the spring action, reliably keeping all
the segments in the proper position. However,
during some operational regimes when the rotor
would otherwise contact the fin(s), the rotor
overcomes the joint forces of spring and steam
pressure and pushes the segments outwards, thus
minimizing all the negative effects from rubbing
as described above. When turbine vibration
and/or rotor misalignments die down, the springs
and steam pressure will push segments back into
their original position, thus restoring the radial
clearance. An additional advantage of this design
is that the segmented seal ring is much stronger
than the rigid seal and therefore is not vulnerable
Figure 28 - Radial Multi-fin Tip Seal (Staked
Design)
This seal is formed by two sets of multiple radial
rigid fins, one set of fins in the stator (diaphragm)
and another set in the shroud OD of rotating
blades. Fins in the diaphragm are of rigid radial
design described above. Fins in the shroud can
be machined either directly from the shroud
(Figure 27) or using the same design as in the
diaphragm (Figure 28). The number of fins (and
pitch) in each set is different. Radial clearances
between fins and cylindrical surfaces of the
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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 They also should prevent leakage, (energy
loss) from the main steam flow into the shaft
area, through equalizing holes in the disc.
diaphragm and rotor “b 1 ” and b 2 ” are much
larger compared to radial seals, while radial
clearances between fins in both sets “b 3 ” are
close to clearances in rigid seals. Effective
sealing ability of this design is provided by a
combination of different clearances and
resistance of chambers with different geometry.
While less common than tip and shaft seals,
nevertheless, root seals up to this day remain a
less investigated component in turbomachinery
design due to the very complicated conditions of
their operation. As it was stated earlier, every
wheel rotates in a separate compartment formed
by two neighboring diaphragms, and creates two
cavities which are interconnected with the
equalizing holes. Steam flow in this area is
complicated, as steam moves, swirling in all
(axial, radial and circumferential) directions,
depending upon numerous variables, such as:
geometry (shape and dimensions) of both
cavities, equalizing holes (location, diameter,
inlet and exit contour), steam conditions
(pressure, stage reaction), regime of operation,
etc. Therefore, steam flow in this area is unstable
and can be dramatically changed depending upon
balance of those variables.
In spite of the high efficiency of such a design, it
has the following deficiencies:
 It is expensive.
 It is extremely vulnerable to damage by
foreign objects in steam or by mishandling.
 Replacement of damaged fins in diaphragm
is costly and complicated.
 The longevity (service life) is limited because
restoration/replacement of damaged fins in
the shroud is almost impossible without
replacement of the entire row of blades.
Another configuration of multi-fin design is
described below in the “Interstage Seals and
Low-Frequency Vibration” section.
The first attempts to describe the steam flow
behavior in this area are dated to the late 1930’s
(Lomakin, 1940) and continued since then with
periods of higher and lower intensity, as it can be
seen in the works of Samoylovich and Morozov
1957, Shvetz et al. 1960, Kapinos et al. 1983,
Wilson et al. 1997. It still remains a point of
considerable interest (Cao et al. 2003, Moroz and
Tarasov, 2003 and 2004). Theoretical research
was combined with testing of various root seal
designs. These theoretical and experimental
investigations resulted in better understanding of
the processes inside the described areas but did
not define general criteria for root seal design.
Therefore, each turbomachinery facility develops
its own approach in resolving this problem.
Below is given a brief description of the
following major known root seal designs: axial
rigid seal, radial rigid seals, aerodynamic
“curtain” seal and cantilevered semi-flexible seal.
All in all, any “healthy” radial tip seal improves
the stage efficiency by approximately 2.5 percent
compared to axial rigid seal. It is possible to
achieve even better results using different
combinations of radial tip seals in HP and LP
stages. For example, implementation of the
contoured end wall and effective tip seals in the
control stage of a large MHI turbine resulted in
increased stage efficiency by about 7% which
was obtained by MHI bench tests (Troyanovski
1993).
References:
Paul et al. 1989, Shcheglyav 1993, Troyanovsky
1993, Cofer 1996, Leyzerovich 1997
ROOT SEALS
Root seals perform two functions:
 They should prevent the shaft leakage from
re-entering in the main flow at the diaphragm
exit and mixing with it. Such a collision
forms additional wakes (eddies) and thus
affects:
o Efficiency, because of diminishing high
tangential momentum of the main flow.
o Reliability, since wakes and eddies in the
main flow increase stimulus acting on the
rotating blades.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
Axial Rigid Seal
Radial Rigid Seal
Similarly to tip seals, this seal is formed by
either a thin (approximately 2.54 mm or 0.10”
thick) segmented plate (Figure 31.1) or by a very
thin (approximately 1.0 mm or 0.040” thick)
sheet (Figure 31.2) installed into the diaphragm
radial groove. The radial groove is located in the
shoulder of the diaphragm inner ring, near the
rotating blade fasteners and is oriented
perpendicular to the shaft axis. Segments of the
plate are staked in the groove, while the sheet
metal is fixed in the groove by a caulking wire.
The geometry of the plate or sheet metal at the
inner diameter also provides a thin fin, which is
engaged with a cylindrical surface on the
shoulder of the disk. The radial clearance “b”
between the thin fin and disk shoulder controls
the steam leakage.
Figure 29 - Axial Rigid Flat Design Root Seal
Figure 30 - Axial Rigid Staggered Design Root
Seal
This seal is formed by one or two so called spillstrips (thin fins) working against a mating flat
face or towards each other. Spill-strips can be
located either on the rotating blade platforms or
on the diaphragm face (Figure 29) or on both
components (Figure 30). Being comprised of
blades and diaphragm, both mating components
of this seal are made from hard metal.
The seal shown on Figure 30 appears to be more
efficient since it does not allow direct collision
of the shaft leak steam with the main flow and
also provides more resistance to steam leaking
from the main flow down. However it is more
costly and provides higher shaft leak entrance
into the rotating blades due to the suction effect
of the main flow.
Figure 31 – 1) Radial Rigid Root Seal Made
From Thin Segmented Plate; 2) Radial Rigid
Root Seal Made From Very Thin Sheet Metal
Since the design of this seal is almost identical to
the radial rigid tip seals, their operation and
drawbacks are the same and are described in the
previous section.
Since the design of this seal is almost identical to
the axial rigid tip seals, its operation and
drawbacks are the same as described in the
previous section.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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Aerodynamic Curtain
positive effect of the “aerodynamic curtain”
seals.
By design, during operation, a small amount of
steam is extracted from the main flow and is
directed axially from the bottom portion of the
diaphragm vanes (near base diameter) into the
rotating wheel passages “p” between the disk OD
and blades platforms (Figure 32). This additional
steam flow forms an aerodynamic barrier
(“curtain”) which prevents collision of the steam
leakage from the shaft seals with the main steam
flow. The shaft seal leakage mixes together with
the “curtain steam” and moves through passages
at the disk OD. In order to eliminate any
additional routes for steam leaks and to minimize
amount of the extracted “curtain steam”, the
balance (equalizing) holes in the rotating disks
are eliminated.
Cantilevered Semi-Flexible Seal
Figure 33 - Cantilevered Semi-Flexible Root
Seal
This design consists of a cantilevered semiflexible segmented seal ring and specially shaped
both (front and exit) cavities which are
interconnected by carefully shaped equalizing
holes. The main design features of the entire root
seal area are as follows (Figure 33):
 The cantilever style segmented seal ring
consists of three major elements: a relatively
thick short “root”, an elongated “neck” and a
“nose” with a thin fin at the end.
 The “root” of this ring is installed into an
axial groove of the diaphragm and is reliably
fixed there by staking.
Figure 32 - "Aerodynamic Curtain" Root Seals
for Axial Entry Rows of Blades
According to testing (Troyanovsky 1993),
properly designed “steam curtains” can improve
stage efficiency by up to 0.5 percent. While
effective at base load regimes, this design has the
following drawbacks:
 The thin fin of the ring “nose” is engaged
with the cylindrical surface of the rotating
disk shoulder, forming a small radial
clearance “b” which controls the steam
leakages.
 Limited application, since this design can be
used only with rotating blades having axial
(not radial) fasteners.
 The proper geometry of the “neck”
(dimensions “L”, “t”, etc.) and entire segment
provides the necessary flexibility to this seal.
 Limited effectiveness at partial and dynamic
regimes.
 The inlet cavity is divided by a cantilevered
seal ring in two areas. The lower portion
protrudes radially from the shaft up to the
seal ring and is shaped to provide the
favorable conditions for steam leaking
through shaft seals to be evacuated through
equalizing holes to the next turbine stage.
The upper portion of this cavity protrudes
radially from the seal ring up to the rotating
blades platform and is formed to minimize
steam leak in radial direction.
 Elimination of the balance holes in the disks
results in significant increase of the thrust
load, which requires implementation of a
special balance (“dummy”) piston on the
turbine front seal and requires a larger thrust
bearing. The substantially increased leakages
of high potential steam through this balance
piston together with increased friction losses
in the larger thrust bearing diminishes the
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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 The equalizing holes have proper diameter,
quantity and profiled inlet and exit entrance
to provide the steam flow passage with
minimum resistance.
The cantilevered semi-flexible seal design has
the following major advantages:
 It effectively prevents shaft steam leaks from
entering and colliding with the main steam
flow by creating:
o Minimal resistance for shaft steam leakage
to pass through the balance holes.
Figure 34 - Different Turbine Stages
o Maximum resistance against moving
radially upward – into the main steam flow.
This improves reliability.
 At the same time, it minimizes steam leakage
from the main steam flow (through the axial
space between the rotating blade roots and
diaphragm) and provides stable leak control
during operation (radial clearance in the root
seal does not change with rotor thermal
expansion).
 This is a very robust design, since the
cantilever seal ring can be individually sized
and made from different materials. Thus
choosing a material combination for each
stage gives optimal flexibility depending
upon particular conditions. Therefore, when
the turbine experiences vibrations and there
is a chance of rubs, each segment can absorb
these shaft movements and minimize the
wear/damage.
Figure 35 - Influence of the Root Seal upon the
Internal Efficiency of an Impulse Turbine Stage
 This design does not require additional axial
space compared to radial rigid seals and can
be incorporated in each stage.
References:
Lomakin 1940, Samoylovich et al. 1957, Shvetz
et al. 1960, Nedzvetzkiy et al. 1974, Kapinos et
al. 1983, Hirota et al. 1985, Wilson et al. 1997,
Pigott et al. 1986, Shcheglyaev 1993,
Troyanovsky 1993, Cao et al. 2003, Moroz et al.
2004
 In addition to its longevity, this design is
maintenance-friendly because it is very easy
to replace the cantilevered seal ring.
Figure 34 and Figure 35 present combined
results of testing some root seals by MHI (Hirota
et al. 1985) and by a Russian R&D facility. A
proper root seal design can result in 1.5 – 2.0
percent efficiency increase per stage.
SHAFT SEALS
Meanwhile, extensive research in seal
technology has also resulted in significant
advances of the shaft seal design. The main R&D
efforts were to create a labyrinth seal ring able to
operate with minimal clearances between its fins
and the rotor, and, at the same time, do not rub
against the rotor and wear which takes place
during long-term operation (Figure 36).
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
Figure 36 - Wear Pattern of Standard Shaft
Seals
Figure 38 - Case Thermal Deformation Due to
Cover-to-Base Temperature Difference
The long-term factors which severely affect
reliable operation of the labyrinth seals are as
follows:
 Case thermal deformation due to temperature
differences across its wall thickness (Figure
37), and top-to-bottom halves of the case
(Figure 38).
 Start-up and shutdown regimes when turbine
rotor runs through 1st critical speed which is
linked to with high vibration (Figure 39).
Figure 39 - Rotor Displacements during Typical
Turbine Startup and Shutdown
References:
Traupel 1997, Shcheglyaev 1993,
Troyanovsky1993, Cofer IV 1996
RETRACTABLE/BRUSH SEALS
Figure 37 - Case Thermal Deformation Due to
Temperature Difference Across the Wall
The next significant step in seal technology was
the development of a retractable seal design.
The effective solution in overcoming the case
thermal deformation is to align all the
diaphragms and packing boxes in the case with
the “thermal key-cross” (as it was described
previously), and provide a sufficient radial
clearance between the case bore and
diaphragm/labyrinth box OD along the
circumference. Also, suspension of the
diaphragms and packing boxes at the case joint
allows for different radial clearances between the
fins and the shaft along the circumference (larger
ones in the lower half and smaller in the upper
half). This design solution makes seals less
vulnerable to case thermal and mechanical
deformations and decreases their wear, thus,
increasing turbine efficiency and reliability.
Figure 40 - Standard Spring-Back Shaft Seal
Ring
In the standard labyrinth seal (Figure 40), each
segment has on its OD springs that force it to be
in the design position during cold start-up and
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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Brush insert bristles are oriented not
perpendicular to the shaft surface (as the
standard fin seals are) but at a certain angle in
the direction of shaft rotation. Such orientation,
together with bristles flexibility, allows
minimizing radial clearance down to 0.0000.005” (0.00-0.13mm). Using retractable brush
seals will result in additional efficiency gain of
1-2 percent without compromising reliability
(Little et al. 2001, Sulda 1999, Foley 2000, Neef
et al. 2006).
shut-down turbine conditions. During the
operation, in addition to springs, steam pressure
acting over OD segments keeps them in design
position.
Contrary to the standard design, in retractable
seals springs are located in the sides of a segment
(Figure 41). They push segments away from the
shaft, creating large clearances between seal fins
and the shaft during turbine start-ups or shutdowns. As these clearances are much larger than
rotor displacements during the 1st critical speed,
the retractable labyrinth seal does not suffer
damage or wear during start-up and shut-down
regimes. During turbine operation at loads
exceeding approximately 5 percent of the rated
capacity, steam pressure overcomes the spring
resistance and pushes the segments back into
working position closing the radial clearances to
design values. Using retractable seals improves
efficiency by approximately 2-3 percent and
increases reliability (Little et al. 2001).
Retractable brush seals are not only more
efficient than standard spring-backed seals; they
are much more stable during long-term operation.
The leakage through a new standard seal is 5-8
times higher than through a retractable brush seal.
During turbine operation, standard seals wear out
and the steam leakage though them increases by
2-3 times more (compared to new condition),
while leakage through a brush seal remains
mostly unchanged. Figure 43 is a plot of tested
leakage through different seal ring designs
versus pressure ratio across the seals, showing
substantial advantages of retractable and brush
seals, especially for long term operation.
Figure 41 - Retractable Shaft Seal Ring
The latest seal technology achievement is a
retractable brush seal design. In this seal, one
standard fin seal in a retractable seal ring is
replaced by a brush seal insert which consists of
a pack of brush bristles secured by two plates
(Figure 42).
Figure 43 - Tested Effectiveness of Different
Shaft Seals
References:
Schofield 1981, TurboCare Retractable Packing
1999, Sulda 1999, Foley 2000, Little et al. 2001,
Neef et al. 2006
Figure 42 - Retractable Brush Seal Ring
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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 Shroud forces which are the result of nonuniform pressure distribution along
circumference of the area beyond the rotating
blades shrouds, i.e. in tip seals.
INTERSTAGE SEALS AND LOWFREQUENCY VIBRATION
It is evident from the above description of the
interstage seals, that the best are the seals with
the minimum technically allowable radial
clearances, because they can minimize the steam
leaks by-passing the main steam flow or
colliding with it. In reality this is not always the
case. Extremely small clearances in the
interstage seals can cause serious subsynchronous (low-frequency) rotor vibration.
This situation usually occurs in steam turbines
operating at high pressures and high power
outputs in which rotors tend to be relatively light
weight. These rotors are typically found in high
pressure turbine-generator sets and in high
pressure / high speed steam turbines used as
mechanical drives in petrochemical and other
industries.
 Labyrinth forces which are caused by nonuniform pressure along circumference of the
multiple chambers of shaft seals.
This theoretical analysis, supported by numerous
tests and experience gained while resolving the
sub-synchronous vibration in actual steam
turbines, revealed that the aerodynamic
excitation forces in the high pressure tip seals are
the dominant cause of the low-frequency
vibration.
The aerodynamic excitation forces are generated
by the tip seals due to inevitable misalignments
between rotor and stator. During turbine
operation, it is impossible to provide perfect
concentricity between rotor and stator because of
the following reasons:
Low-frequency vibrations are self-excited rotor
oscillations in the journal bearings oil film. They
occur at operating speed with a frequency equal
to the first natural rotor frequency in the entire
train rotor system, which incidentally, is close to
one-half of the steam turbine rotor frequency.
This vibration is called "self-excited" because a
turbine rotor which is operating smoothly will
reach a load, called the "threshold" load, where it
suddenly loses its stability and produces intense
vibrations - without any visible outside causes.
 Rotor precession in journal bearings (i.e. its
constant movement within the oil film limits).
 Thermal misalignment due to different rotor
and stator temperatures in axial and radial
directions.
 Mechanical deformations due to steam
pressures, vacuum in condenser, piping
influence, etc.
These misalignments cause different clearances
in the interstage seals along the circumference of
the rotor. The uneven clearances cause uneven
steam leaks. Figure 44 explains the mechanism
of steam low-frequency vibration.
There are two main sources of the low-frequency
vibration:
 "Oil swirl" (rotor oscillations) caused by
hydrodynamic forces in oil film of the journal
bearings at low static load.
 "Steam swirl" (rotor oscillations) caused by
unbalanced aerodynamic forces in the steam
path which can act separately or in
combination with the additional
hydrodynamic excitation in journal bearings.
The theory of "steam swirl" anti-vibrational
stability of the interconnected rotors system was
developed by A.G. Kostyuk in 1972 and V.I.
Olimpiyev in 1976, and was described in detail
by A.D. Trukhny in 1990.
“Steam swirl”, according to the theory, can be
generated by three types of aerodynamic forces
that occur within a steam path:
 Circulating forces which are produced by the
rotating row of blades due to irregularities in
the rotating force.
Figure 44 – Interstage Seals and Low Frequency
Vibration
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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One of the best anti-vibration tip seal designs is
shown on Figure 45. The main feature of this
design is large radial clearances δ 4 and δ 5 in all
the fins. The effectiveness of this design is
achieved by the proper combination of axial (δ 1
and δ 3 ) and radial clearances (δ 4 and δ 5 ). Large
radial clearances are insensitive to radial rotor
displacements: they do not cause substantial
change in steam leakages, maintaining the
balance between main steam flow and leakages
around circumference during turbine operation.
In this figure the rotor is shown displaced in
vertical direction from its original (cold
assembly) position “O” down by a value "a". The
clearances in lower half are smaller than in the
upper half. The minimal clearance is at the 6
o’clock position, while the maximum clearance
in upper half is at 12 o’clock position.
So, steam leaks increase through the upper half
and decrease through the lower half. As a result
of uneven leakages, the main steam flow through
the stage will also be uneven: it will be decreased
in the upper half and increased in the lower half,
generating uneven circulating force (F upper <
F lower ). The resulting unbalanced aerodynamic
radial force "C" is applied to the center of the
rotor and rotates together with the rotor but 90
degree ahead of the dynamic rotor displacement.
Concurrently, uneven clearances in the interstage
seals cause another radial force "D" in the area of
reduced clearances, due to the increased steam
pressure there. This force acts as a "lifting"
element trying to return the rotor from the
misplaced position back to the center of the
stator. These two unbalanced radial forces
depend upon numerous variables, such as: steam
path (including interstage seals) design, steam
pressure, stage reaction, value of rotor radial
displacements, etc.
Figure 46 – Standard Shaft Seal with AntiVibration Addition
The turbomachinery industry has developed
effective anti-vibration seal designs and
recommendations for their implementation.
For tip seals it is recommended to use a singlefin seal with the minimal possible radial
clearances or to use two (inlet and exit) singlefin seals with different clearances. The seal on
the inlet side of the flow should have the smaller
clearance and the seal on the exit side of the flow
should have the larger radial clearance. Also, the
chamfer between the seals in a two seal design
should have a certain volume to compensate for
the "shroud" excitation forces. This is the exact
design implemented in the current Aramco
diaphragms.
Figure 47 – Retractable Brush Seal with AntiVibration
The effective designs of anti-vibration shaft seals
are shown on Figure 46 and Figure 47. A special
"nose" (1), with a set of axial channels (2),
(located along circumference of its bore) is
arranged in front of seal fins. These axial
channels damp irregularities in the leakage steam,
minimizing labyrinth forces. Those seals were
not used in Aramco turbines since calculated
labyrinth forces in the subject steam path are
insignificant.
References:
Kostyuk 1972, Olimpiyev et al. 1975, Olimpiyev
1976, Runov 1982, Trukhny 1990, Leyzerovich
1997
Figure 45 – Anti-Vibration Tip Seal
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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PROGRESS IN MECHANICAL
STRENGTH CALCULAIONS
theoretical research (Muster and and Sadowsky
1956, Naumov 1960, Ingultsov 1958) and
experimental testing (Taylor 1951, Sentsov 1958,
Kulagina 1960, Mellerovich and Bliznyukova
1961) – in order to improve the existing
calculation methods.
A turbine diaphragm, due to its complex
structure and harsh operating conditions, is a
very difficult subject for strength calculations.
As previously stated, structurally, a diaphragm is
a complicated 360 degree plate, which is
composed of several major components made
from different materials. The outer and inner
rings are connected together with a set of
beams/vanes of complex geometry. The
diaphragm is split in two halves at the horizontal
joint. The outer ring is supported by its OD in
the case while the inner ring is “free.” The
diaphragm is loaded in two main ways. First, the
entire diaphragm is subjected to a uniformly
distributed steam pressure load across its inlet
face. Second, the vanes are subjected to a
bending moment due to the impact of high
velocity steam flow. In addition to these loading
conditions, other factors of concern are thermal
gradients in the radial direction and between
diaphragm halves in the range of 25-65°F (1535°C), and vibration forces caused by the steam
flow.
The tests revealed that, contrary to theoretical
assumptions, all the connections between the
vanes and both (outer and inner) rings, and
between the outer ring and the case, are elastic
(not rigid). Elasticity in those connections turned
out to be critical for strength calculations.
All in all, this R&D allowed to:
 Modify the Smith Method by introducing
new semi-empirical factors which more
accurately reflect real loading in the
diaphragm components.
 Together with the gained experience of
turbines operation, establish semi-empirical
“safe” allowable stresses and deflections in
the diaphragms calculated by the Smith
Method.
Despite the fact that measured stresses in the
vanes located at the horizontal joint were still
significantly higher than calculated values, the
“Modified Smith Method” suited well for
designing diaphragms of small and medium size
turbines.
The combination of two diaphragm halves
(instead of a solid continuous plate), and the
plurality of the vanes with complex geometry
present the main problems for stress and
deflection calculations. Because of this structural
complexity, strength calculations were possible
in the past only by simplifications of the
mathematical model.
However, this method showed its limitations
when applied to larger turbines operating with
higher steam conditions and/or higher loads.
Several catastrophic failures occurred: due to
vanes overloading they were broken completely
resulting in disintegration of the entire
diaphragm with severe damage to the rotor and
case (Dodd et al. 1975).
However, mechanical strength calculations for
diaphragms have also been progressing along
with the evolution in their design.
Up to the 1950’s, OEM’s calculated diaphragms’
strength using classical methods developed by
A.M. Wahl (1930, 1932), and D.M. Smith
(1938). A.M. Wahl analyzed diaphragms as a
solid annular half ring plate (without vanes)
which was rigidly supported by its outer contour
and was subjected to a uniformly distributed load.
D.M. Smith considered diaphragms as a plurality
of rigid beams, each beam being supported
rigidly in the case and rigidly connected with a
massive inner ring. This structure was also
considered to be subjected to a uniformly
distributed load.
Such failures stimulated turbine manufacturers to
develop new, more accurate calculation tools for
large steam turbine diaphragms. One of these
methods was introduced in the mid 1980’s
(Kostyuk 1982) based upon the mathematical
model proposed by Pakhomov in 1934, and the
results of numerous theoretical and experimental
works done in the former Soviet Union.
However, being quite simple and handy, this
method still did not reflect the actual stresses,
deformations, and deflections; especially in the
vanes.
The real break-through in diaphragm strength
calculations occurred with the advent of the
finite element analysis (FEA) method; a
numerical procedure in which the solution
Naturally, stresses and deflections produced by
these calculations significantly deviated from the
actual values. Therefore the turbomachinery
industry continued intensive R&D, both
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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domain (i.e. diaphragm) is discretized into nodes
and elements, producing a large set of algebraic
equations representing the entire system. After
the application of boundary conditions and loads,
these equations are solved simultaneously to
obtain information at the location of the nodes
such as displacement or stress. While the finite
element method can be traced back to the early
1900’s, its true value was not realized until
combined with computers in the 1970’s. With
the help of computers, a component can be
modeled with thousands of nodes and elements,
allowing for highly accurate solutions (Moaveni
2008). With this valuable tool, a diaphragm can
be quickly modeled and analyzed, allowing the
designer to retrieve data (accurate stresses and
deformations) at any point in the component.
at these two locations in the diaphragm arc. The
diaphragms are represented by the blue, tan,
purple, orange and teal lines are all remaining
known diaphragms that have experienced
failures during 40 plus years of operation due to
greatly exceeding the yield limits of the vane
material at the joint. The red, yellow, and pink
lines represent the subject Aramco diaphragms
(all at the current pressure differential). Clearly,
the vanes of the original design exceed yield
strength limits (above 2 percent of nodes).
However, with the introduction of EB welding
and 360 degree construction, the stresses are
significantly reduced (as shown by the yellow
line) to less than 1 percent of nodes exceeding
yield strength. The current diaphragm
represented by the pink line, has even smaller
stresses with essentially zero nodes exceeding
the yield strength limit.
FEA of Saudi Aramco Diaphragms
During the production of all Saudi Aramco
diaphragms (except for the “current design”), the
mechanical strength was calculated using the
“Modified Smith Method.” Mechanical strength
of all the “current design” diaphragms was
calculated by FEA.
Since the majority of Saudi Aramco steam
turbine rotor damage occurred in stage #4 area,
the following stage #4 diaphragms were
analyzed at the current operating conditions:
 The original diaphragm with welding kerfs
using 2 x 180 degree sub-arc welding
technology.
 An “imaginary” diaphragm: the original
diaphragm design geometry but fabricated
using 1 x 360 degree electron beam (EB)
welding with 100 percent penetration.
Figure 48 - FEA Calculations: Deflections and
Equivalent Stresses in Original Design
 The current diaphragm design (2010 –
present)
FEA results are presented on Figure 48, Figure
49, Figure 50 and Table 1. Figure 48 and Figure
49 show calculated deflections and equivalent
stresses in the entire original (A) and current (C)
diaphragm halves. Figure 50 demonstrates the
distribution of vane stresses (as percentage above
yield strength) in one diaphragm half for these
diaphragms and for all remaining known
diaphragms that failed during 40 plus years of
operation. The “0” and “180” degree locations
are representative of the vanes at the horizontal
joint of the diaphragm. The acceptable allowance
has been set at below 2 percent of vane nodes
above yield strength. It is clearly shown that no
matter the diaphragm, stresses are always highest
Figure 49 - FEA Calculations: Deflections and
Equivalent Stresses in Current Design
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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Figure 50 - Historical Data: Vanes Stress Evaluation
Table 1 - FEA Results for the Stage #4
Diaphragm Scenarios
 The original vanes were not strong enough.
Even fabricated by EB welding (as
demonstrated by the “imaginary” diaphragm)
with 100 percent penetration, they still have
high stresses and produce nearly the same
deflections as the original sub-arc welded
diaphragm. Therefore, the diaphragms
failures occurred not due to operation and
maintenance (O&M) mistakes, but rather to
the mismatch of the original diaphragms
design specifications and the new operating
conditions that were imposed by the plant
requirements.
Table 1 presents maximal stresses and
deflections for all three chosen diaphragm
designs. It is evident from Table 1 that:
 The mechanical strength of the current
diaphragm is substantially higher compared
to the original diaphragm due to the
implemented advances in design, described
above, and manufacturing, described in the
following section. Calculated stresses and
deflections for current diaphragms are well
within the design limits for current operating
conditions.
 Even with ideal kerfs and welding processes
( i.e. no deviations from the nominal design
dimensions and procedures), the original
diaphragm is unsuitable for the current
operating conditions due to the high
deflections and the very high calculated
stresses in the vanes at the horizontal joint,
which exceed the material yield stresses even
at room temperature. In actual original
diaphragms having typical deviations from
the design made during pre-machining,
assembly and welding, the stresses and
deflections may be significantly higher than
the values given in Table 1.
References:
Wahl 1930 and 1932, Pakhomov 1934, Smith
1938, Taylor 1951, Muster et al. 1956, Ingultsov
1958, Sentsov 1958, Kulagina 1960, Mellerovich
et al. 1961, Dodd et al. 1975, Kostyuk 1982,
Moaveni 2008
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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industries. New tools and manufacturing
processes (advances in metallurgy, numerically
controlled machinery, electron beam welding,
shot peening, laser cutting, water jet cutting and
electro-discharge machining (EDM) wire cutting,
etc.) made it possible to produce steam turbine
diaphragms and other steam turbine components
with a higher degree of accuracy and a better
surface finish in a more cost-effective manner.
EVOLUTION IN DIAPHRAGM
MANUFACTURING
Generally, a steam turbine diaphragm is a very
difficult component to manufacture due the
following conflicting requirements:
o It consists of different parts made of
different materials.
o It must have exceptional mechanical
strength in harsh environments.
Below is a brief description of the evolution in
diaphragm manufacturing. This evolution is also
shown in Figure 51, which summarizes the
progress in design and manufacturing of turbine
diaphragms achieved during the last 40 years.
o The diaphragm vanes must be precisely
oriented, and have uniform pitches,
heights, throats, and leans, to form and
direct the steam flow (jet), into the rotating
blades with maximum efficiency and
minimal stimulus.
Original Diaphragms
The original diaphragms of the subject turbines
were built using 1950’s design and technology
(Sentsov 1956, German and Kulakova 1957,
Zemzin and Frenkel 1960, Houldcroft 1977).
The cage and the entire diaphragm were made in
two 180 degree halves. The profile holes in the
spacing strips were punched while the strips
were in the flat condition. Only after punching
the profile holes, were the spacing strips rolled to
form two 180 degree semicircles of the proper
diameter. The profile holes and their pitching
were inaccurate. The plunging process limited
Combination of heavy welding with high
accuracy presents approximately the same
dilemma as “making a Swiss watch with a sledge
hammer”. Therefore diaphragm manufacturing
includes numerous intermediate check/inspection
points, including charting of the steam passing
area, and a significant amount of hand-dressing
prior to shipment.
Progress in electronics, computerization and
modeling resulted not only in the break-through
design improvements described above. It also
accelerated significant progress throughout many
Figure 51 - Diaphragm Design and Fabrication Evolution
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
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the spacing strip thickness to 0.140” (3.6mm)
maximum and required larger holes and radial
orientation of the plunger tool axis. Therefore the
exit edge of the profile was parallel to the radial
line which resulted in the negative tangential
lean of the vanes with all the deficiencies
described above.
welding process and by the loose metal chips
produced during final machining.
So the described manufacturing method could
not provide desired accuracy and required
significant hand-dressing of each set of vanes.
This resulted in reduced efficiency (from poor
channel aerodynamics), reliability, (because of
the large stimulus spectrum acting on the
rotating blades), and the reduced diaphragm
strength.
The welding kerfs in the outer ring and center
had non-optimal geometry which could not
provide the maximum possible mechanical
strength. Four massive sub-arc welds generated
an enormous amount of heat, which caused
significant thermal deformation. Also during
welding, the thin spacing strips were often
burned through, allowing melted metal to
penetrate into the steam path and damage the
vanes. Figure 52 and Figure 53 show the major
steps in such diaphragm fabrication.
Improved Diaphragms
Manufacturing progress in the 1970’s and 1980’s
led to significant improvements in cage
construction. Industry started laser or water-jet
cutting profile holes while the spacing strips
were in the rolled condition (Matsui et al. 1987).
This allowed for a substantial increase of the
spacing strip thickness (1.4 times and more)
compared to the punching method, making the
spacing strips sturdier and preventing burnthrough damage. Better accuracy of the profilehole geometry and pitching resulted in improved
efficiency and decreased stimulus on the rotating
blades.
Although cage construction improved, the
original diaphragm fabrication method of
producing two separate 180 degree halves using
the original vane welding kerfs and sub-arc
welding remained. Therefore, this manufacturing
method had the same major original flaws as
described above.
Figure 52 - Obsolete Building of a 180 Degree
Cage
So, the improved diaphragm also required
significant hand-dressing of the vanes and had
limitations in efficiency, reliability and strength.
Modified Diaphragms
The limitations of the improved diaphragms
were still troublesome. The next step in
diaphragm evolution was the “modified
diaphragm”. The modified diaphragm utilized a
new machining method: Electrical Discharge
Machining, (EDM), which was developed in the
late 1980’s to the early 1990’s (Fuller 1989).
Figure 53 - Obsolete Fabrication of 2 x180
Degree Diaphragms by Sub-Arc Welding
During the machining of the diaphragm
horizontal joint, the vanes in both diaphragm
halves were unavoidably split and partially cut.
The cut vanes at the horizontal joint of one half
of the diaphragm did not accurately match with
their mating parts in the opposite half. Also due
to the minimal vane set back, all the vanes were
vulnerable to mechanical damage during the
While the sub-arc welding method and the vane
airfoils remained the same, the modified
diaphragm utilized substantially improved cage
and diaphragm construction. The diaphragm was
built as a single, 360 degree ring with
significantly improved welding kerfs, (instead of
two 180 degree halves with the original kerfs).
Improved welding kerfs provided better bonding
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
between the outer ring, “squirrel cage”, and the
center ring, increasing the diaphragm mechanical
strength by approximately 15 percent and
decreasing thermal deformation. EDM wire
cutting was used to split the diaphragm in halves
using a wire diameter of 0.010 – 0.015 inch
(0.25-0.38 mm). The subject cutting method did
not remove much metal and therefore, allowed
accurate mating of the split vanes at the
horizontal joint. This modified diaphragm had
increased accuracy, efficiency and reliability
(due to decreased stimulus) while the amount of
hand-dressing of the diaphragm vanes was
reduced. But their mechanical strength soon
became inadequate to the changed operation
conditions.
amount of hand-dressing required. Figure 54 and
Figure 55 show the major steps in diaphragm EB
welding, while Figure 56 and Figure 57 show the
differences in diaphragm quality produced by
sub-arc and EB welding.
Figure 54 - Current Building of a 1 x 360
Degree Cage
Current Diaphragms
The current diaphragms utilize further
improvements which occurred in the 1990’s: the
development of new analytical tools and
manufacturing technologies combined with
continuous design improvements.
Diaphragm stresses and deflections in all major
components were obtained using FEA which was
calibrated against in-house test results.
It became possible to use Electron-Beam (EB)
welding technology instead of traditional sub-arc
welding (Akutsu 1980, Schiller et al. 1982,
Matsui 1987). EB welding, as the name implies,
uses a high energy electron beam in deep
vacuum, directed at the joint of the tightly
assembled parts. The high temperature of the
beam melts the metal at the joint where the
assembled parts meet, welding them together in a
very short period of time, which minimizes their
heating. Welding in deep vacuum produces
higher quality welds which have no porosity or
inclusions. The possible axial depth (penetration)
of EB welds is significantly higher than the
depth of sub-arc welds. EB welds in many
diaphragms penetrate 100 percent of the axial
vane width. This deep weld penetration,
combined with the high quality of the EB weld
and new stronger airfoil, substantially increases
the diaphragm strength when compared to a
similar diaphragm fabricated by sub-arc welding.
Also the EB welding method produces 8-10
times less heat compared to sub-arc welding
methods. Therefore, EB welding causes
insignificant thermal deformation of the
diaphragm, which in turn results in higher
efficiency and reliability, (due to reduced
stimulus on the rotating blades) and minimizes
Figure 55 - Advance Fabrication of a 1 x 360
Degree Diaphragm by EB Welding
So, utilization of the advances in stress analysis,
(FEA), design achievements, (vane profiles, vane
lean, tip and retractable shaft seals, etc.), and
progressive manufacturing technologies, (CNC
machining, EB welding and EDM wire cutting)
allowed creation of the new, current generation
of diaphragms having much higher efficiency,
reliability, and mechanical strength without
reworking the steam turbine cases or shafts.
Saudi Aramco has already replaced existing
diaphragms with new ones in 5 of their turbines
and plan to complete this replacement on the
remaining 3 units by the end of 2013.
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
turbine efficiency. However, root seals require a
special shoulder in the inlet face of each rotor
wheel to interact with the fin of the seal ring that
would be installed into the exit face of the
diaphragm. Retractable brush seals require
additional axial space on the shaft in order to
accommodate the brush which imposes change
the standard rotor tongue-and-groove
configuration.
The existing steam path is characterized by an
abrupt 50 to 100 percent steam admission due to
non-optimal steam path geometry of both the
diaphragms and the rotor which does not
accommodate the specifics of natural steam
expansion. High velocity steam flow leaving the
first (control) stage with a partial (approximate
50 percent) arc admission does not have enough
room to spread 100 percent circumferentially.
Therefore it can’t pass through the second stage
and beyond (designed for full arc admission)
without suffering significant energy losses that
seriously affect the overall steam turbine
efficiency.
Figure 56 - Deficiencies of Obsolete 2 x 180
Degree Diaphragm Fabricated by Sub-Arc
Welding
New diaphragms paired with a new/modified
rotor allow a steam path design with a smooth,
gradual transition from 50 to 100 percent steam
admission designed in accordance with natural
steam expansion for a given turbine geometry.
This will minimize/eliminate the energy losses
associated with abrupt steam admission and thus
substantially increase the overall steam turbine
efficiency.
Figure 57 - Advantages of Current 1 x 360
Degree Diaphragm Fabricated by EB Welding
with EDM-Wire Cutting
Optimal Diaphragms
Current diaphragms, being superior to previous
designs, nevertheless cannot provide the highest
possible benefits (efficiency and reliability) to
the steam turbine since they do not utilize the
following design improvements which require
changes to the existing rotors:
Implementation of these advanced features in
developing the optimal steam path, (optimal
diaphragms and new/modified rotors
combinations within the steam turbine), will
further improve efficiency by 8 to 10 percent due
to minimizing energy losses in the main steam
flow and decreased leakages while also
improving reliability due to decreased wakes and
smoother main steam flow.
 New rotating blades with advanced airfoils
and integral shrouds;
 Root seals;
 Retractable brush seals;
However, this design is much more expensive
compared to all previous diaphragm designs.
Therefore Saudi Aramco has decided to postpone
its implementation pending the further
development of the entire UGP and SGP plants.
 Optimal smooth 50 percent to 100 percent
steam admission.
New vanes with advanced airfoils in the current
diaphragms cannot be installed in the optimal
position, (set-up angle), because they have to
operate with the original vintage rotating blades.
Therefore, they could not realize their full
potential efficiency. Only a combination of new
vane with new rotating blade can provide the
highest possible efficiency.
References:
Sentsov 1956, Bistritzkiy 1956, German et al
1957, Zemzin et al. 1960, Houldcroft 1977,
Akutsu 1980, Schiller et al. 1982, Matsui et al.
1987, Fuller 1989, Sun et al. 1996, Dragunov et
al. 2005
As it was described before, root seals and
retractable brush seals also improve the steam
Copyright © 2013 by Siemens Energy, Inc.
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Proceedings of the Second Middle East Turbomachinery Symposium
17 – 20 March 2013, Doha, Qatar
CONCLUSIONS
 The entire story of manufacturing several
generations of diaphragms for Saudi Aramco
UGP and SGP steam turbines presents an
evolution in diaphragm design, calculations,
and manufacturing in the steam
turbomachinery industry for almost forty
years.
Cao, C., Chew, J.W., Millington, P.R., Hogg,
S.I., 2003, “Interaction of Rim Seal and
Annulus Flows in an Axial Flow Turbine”, Proc.
of ASME Turbo Expo 2003, Power for Land,
Sea, and Air, June 16-19, 2003, Atlanta, Georgia,
USA, GT2003-38368
 Evolution in steam turbine diaphragm design,
strength calculations, and manufacturing
resulted in substantial increase of their
efficiency, reliability, mechanical strength,
and maintainability.
Cofer IV, J.I., April 1996, “Advances in Steam
Path Technology”, ASME Journal of
Engineering For Gas Turbines and Power Vol.,
118, pp. 337-352
 Each diaphragm generation, for UGP and
SGP, starting from the original parts, was
designed and manufactured at the edge of
existing technology. While being reliable for
the provided design operating conditions, the
margin of their mechanical strength
decreased at aggravated loads imposed by
increased plant output. Therefore, some of
them deteriorated and failed causing serious
rotor damage.
Dawes, W.N., 1988, “Development of a 3D
Navier Stokes Solver For Application to All
Types of Turbomachinery”, @ ASME Paper 88GT-70.
Dawes, W.N. , 1990, “Towards Improved
Through Flow Capability: The Use of 3D
Viscous Flow Solvers a MultiStage
Environment”, @ ASME Paper 90-GT-18.
 The current generation of diaphragms, being
in the process of installation in the UGP and
SGP turbines with existing rotors, has
substantially increased mechanical strength,
efficiency, reliability, and maintainability.
These diaphragms are completely adequate
for current and planned future increases in
plant output.
Deckers, M., Simon, V. and Scheuerer, G., 1997,
“The Application of CFD to Advanced Steam
Turbine Design”, International Journal of
Computer Applications Technology
Dejch, M.E. and Dejler, S., 1994, “Experimental
Investigations of a Stator Cascade Generating
Reduced Secondary Losses”, Teploenergetika,
No. 10 (in Russian)
 There is potential for implementation of
optimal diaphragm design, which, combined
with the modernization of existing rotors and
other components, will further increase
turbine efficiency, reliability, and
maintainability.
Dejch, M.E., Zariankin, A.E. and Filippov, G.A.,
1960, “ The Method of the Efficiency
Improvement in Turbine Stages with Short
Blades”, Teploenergetika No. 2, pp. 18-24 (in
Russian)
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Atkins, M.J., 1987, “Secondary Losses and EndWall Profiling in a Turbine Cascade”, @ Imeche
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Akutsu, Y., May 1980, “Application of Electron
Beam Welding to Steam Turbine Diaphragms”,
Proc. of the International Beam Technology
Conference, Essen, Germany, pp. 63-69
Dodd, V.R., Dubner, R.M. and Caruso, W.J.,
September 1975, “Steam Turbine Diaphragm
Failure”, Proc. of the ASME Petroleum
Mechanical Engineering Conference, Tulsa,
Oklahoma, USA.
Bistritzkiy, N.D., 1956, “Steam Turbine
Diaphragms”, Mashgiz (In Russian)
Copyright © 2013 by Siemens Energy, Inc.
29
Proceedings of the Second Middle East Turbomachinery Symposium
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Dragunov, V.K., Goncharov, A.L. and Nemytov ,
D.S., 2005, “Formation of Diffusion Interlayers
in Electron Beam Welded Joints in Diaphragms
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Jansen, M. and Ulm, W., 1995, “Modern Blade
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Filippov, G.A. and Zhong-Chi Wand., 1964,
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Kapinos, V.M., Matveev, Yu.Ya., Pustovalov,
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Foley, M., July 2000, “Retractable Brush Seals
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Kawagishi, H. and Kawasaki, S., 1991, “The
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German, S.I. and Kulakova, G.N., 1957,
“Investigation of the Welded Joints Produced by
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29-33 (in Russian)
Kulagina, G.F. ,1960, “Experimental
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Goel, S., Singh, H. and Cofer IV, J.I., 1993,
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AKNOWLEDGEMENTS
The authors wish to express their gratitude to the
following people for their active participation in
the entire project, and valuable assistance in
preparing and organizing this paper: Carolyn B.
Smith, Kenneth E. Bruce, Timothy Ewer, Russel
V. Caggiano, Tyler H. Aaron, Shannon Pedersen,
and Tariq Al-Shaikh.
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Copyright © 2013 by Siemens Energy, Inc.
33
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