Note: Descriptions are shown in the official language in which they were submitted.
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TURBINE BLADE WITH RECESSED TIP
FIELD OF THE INVENTION
The invention lays in the field of turbine blades for highly loaded axial
rotor turbines as e.g.
used in aero engines or for power generation.
DESCRIPTION OF THE ART
In axial turbine the tip clearance flow occurring in rotor blade rows is
responsible for about
one third of the aerodynamic losses in the blade row and in many cases is the
limiting factor
for the blade lifetime.
io The tip leakage vortex forms when the leaking fluid crosses the gap between
the rotor blade
tip and the casing from pressure to suction side and rolls up into a vortex on
the blade suc-
tion side. The flow through the tip gap is both of high velocity and high
temperature, with
the heat transfer to the blade from the hot fluid being very high in the blade
tip area. In
order to avoid blade tip burnout and a failure of the machine, blade tip
cooling is commonly
is used.
The tip clearance flow has been investigated recently with a number of
contributions in the
open literature. An important contribution always referred to i s the work by
Rains [1 ](see
subsequent list of references), who studied tip clearance flow for axial flow
pump motivated
by concerns of cavitation. Moore & Tilton [2] investigate the tip leakage flow
both experi-
20 mentally and analytically. The flow structure inside the gap and heat
transfer to the blade
have been discussed. A flow model assuming the gap losses coming from complete
mixing
behind the vena contracta (in general, point in a fluid stream where the
diameter of the
stream is the least) leading to uniform flow conditions at the gap outlet is
presented. Bin-
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don [3] measured and investigated the tip clearance loss formation. He could
divide the
total endwall loss into loss generated inside the tip gap, mixing loss of the
tip leakage vor-
tex, and secondary and endwall losses. He concluded that not only tip leakage
mass flow is
important for loss generation (48% of overall loss seen in mixing loss), but
also the flow
structure inside the gap would play a significant role (39% of overall loss
generated inside
the gap). Furthermore, he showed a conceptual model for tip clearance loss
formation. Bin-
don & Morphis [4] investigated loss for different blade tip geometries. They
found that the
overall loss remained unchanged although gap losses strongly varied between
the base line
sharp edged flat tip and differently contoured blade tips. Whereas the sharp
edge case
showed high losses inside the gap with a strong separation bubble at the
pressure side lip,
the contoured cases showed less losses since no separation bubble was formed
on the gap
inlet but in turn an increased tip gap mass flow was found.
In a study of different squealer tips by Heyes, Hodson & Dailey [5] it was
also concluded
that the separation bubble with the associated vena contracta is effectively
sealing the gap
is and by reducing tip mass flow the tip clearance losses may be decreased. In
this study also a
flow model for squealer tips is included as well as a model for the flow at
the gap exit which
is represented as a combination of an isentropic jet between the casing and
gap mid height
and a wake formed behind the separation bubble. A computational study by Ameri
[6] on
heat transfer on the blade tip showed that a flat tip with sharp edges
performs best in terms
of efficiency and total pressure loss compared to a mean camberline squealer
tip and a flat
tip with radiused blade tips. Further computational studies by Tallman [7],
[8] on the ex-
perimentally investigated linear cascade by Bindon & Morphis [4] discuss the
effect of tip
clearance height and relative casing wall movement on the flow physics in the
tip gap. A
further experimental investigation on the tip clearance flow physics due to
moving casing
wall is presented in a two part study by Yaras & Sjolander [9,10]. It was
found that the
moving belt simulating relative casing motion significantly decreased tip gap
mass flow.
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Also the tip passage vortex is drawn to the suction side, providing a
throttling effect. Fur-
thermore a reduced pressure difference driving the flow into the gap was
observed.
In a recent study, detailed heat transfer to recessed blade tip was first
investigated by Dunn
et al. [12]. A recessed blade tip was equipped with heat flux gauges and
experimentally
investigated in a full stage rotating turbine. Nusselt Number was shown for
different
vane/blade spacings. It was found that the leading edge Nusselt Number on the
cavity
bottom were in excess of the blade stagnation value.
LIST OF REFERENCES
[11 Rains, D.A., 1954, "Tip Clearance Flows in Axial Flow Compressors and
Pumps", Cali-
fornia Institute of Technology, Hydrodynamics and Mechanical Engineering
Laborato-
ries, Report No. 5, June 1954.
[2] Moore, J. & Tilton, J.S., 1988, "Tip Leakage Flow in a Linear Turbine
Cascade,",ASME
Journal ofTurbomachinery, Vol 110, pp. 18-26.
[3] Bindon, J.P., 1989, "The Measurement and Foramtion of Tip Leakage Loss,"
ASME
Journal of Turbomachinery, Vol 11 1, pp. 257-263.
[4] Morph is, G. & Bindon, J.P., 1992, "The Development of Axial Turbine
Leakage Loss for
Two Profiled Tip Geometries Using Linear Cascade Data," ASME Journal of Turbo-
mach i nery, Vol 114, pp. 198-203.
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[5] Heyes, F.J.G., Hodson, H.P., & Dailey, G.M., 1992, "The Effect of Blade
Tip Geometry
on the Tip Leakage Flow in Axial Turbine Cascades," ASME Journal
ofTurbomachinery,
Vol 114, pp.643-651.
[6] Amen Ali, A., 2001, "Heat Transfer and Flow on the Blade Tip of a Gas
Turbine
Equipped with a Mean-Camberline Strip," ASME Paper 2001 -GT-01 56.
[7] Tallman, J. & Lakshminarayana B., 2000, "Numerical Simulation of Tip
Clearance
Flows in Axial Flow Turbines, With Emphasis on Flow Physics, Part I - Effect
of Tip
Clearance Height," ASME Journal of Turbomachinery, Vol 123, pp. 314-323.
[8] Tallman, J. & Lakshminarayana B., 2000, "Numerical Simulation of Tip
Clearance
io Flows in Axial Flow Turbines, With Emphasis on Flow Physics, Part II -
Effect of Outer
Casing Relative Motion," ASME Journal ofTurbomachinery, Vol 123, pp. 324-333.
[9] Yaras, M.I. & Sjolander, S.A., 1992, "Effects of Simulated Rotation on Tip
Leakage in a
Planar Cascade of Turbine Blades: Part I - Tip Gap Flow," ASME Journal of
Turbo-
mach i nery, Vol 114, pp. 652-659.
[10] Yaras, M.I. & Sjolander, S.A., 1992, "Effects of Simulated Rotation on
Tip Leakage in a
Planar Cascade of Turbine Blades: Part II - Downstream Flow Field and Blade
Load-
ing," ASME Journal of Turbomachinery, Vol 114, pp. 660-667.
[111 Basson, A.H. & Lakshminarayana, B., 1995, "Numerical Simulation of Tip
Clearance
Effects in Turbomachinery," ASME Journal ofTurbomachinery, Vol 109, pp. 545-
549.
[12] Dunn, M.G. & Haldemann, C.W., 2000, "Time Averaged Heat Flux for a
Recessed
Blade Tip, Lip and Platform of a Transonic Turbine Blade", ASME Paper GT2000-
0197.
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[13] Behr, T., Kalfas, A. I., Abhari, R. S., " Unsteady Aerodynamics in Casing
Injection for Tip
Leakage Treatment in an Oneand-1 /2-Stage Unshrouded Turbine", ASME Paper No.
GT2006-90959.
[14] Sell M., Schlienger J., Pfau A., Treiber M., Abhari R.S., 2001, "The 2-
stage Axial Turbine
5 Test Facility LISA", ASME Paper No. 2001 -GT-049.
SUMMARY OF THE INVENTION
The invention is directed to a recessed blade tip for a highly loaded axial
rotor turbine blade
with application in high pressure axial turbines in aero engine or power
generation.
io To overcome the problems known from the prior art a blade tip design is
suggested with a
recess in the blade tip instead of a simple flat blade tip. The recess
(cavity) inside the blade
tip acts as an aerodynamic seal which improves the performance of the turbine
and/or re-
duces the heat load of the turbine tip, in that it takes influence on the
flows distribution.
Since material at the upper surface of the blade is moved to a lower radius
from the axis of
rotation, blade root mechanical stresses can be lowered. Also in case of a tip
rub, i.e. when
the rotor blade touches the casing during rotation, only the thin cavity rim
is damaged.
Wear damage to the casing is also limited and since the purge holes for the
blade tip cool-
ing are located inside the cavity, the rubbing does not damage the outlet of
the holes. Effi-
cient cooling is hence assured even if rubbing occurs. Finally, the recess
cavity may act as a
labyrinth seal, which could be beneficial in reducing tip clearance mass flow.
It has been observed that by an appropriate profiling of the recess shape, the
total tip heat
transfer Nusselt Number can significantly reduced, e.g. being 15% lower than
the flat tip
and 7% lower than the baseline recess shape. Experimental results also showed
an overall
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improvement of 0.3% in the overall turbine total efficiency with the improved
recess deign
compared to the flat tip case, validating a 0.38% prediction from the CFD
analysis.
With use of a three dimensional Computational Fluid Dynamics (CFD), the flow
field near
the tip of the blade for different shapes of the recess cavities is
investigated. Through con-
trol of cavity vertical structures, an improved design is achieved and the
differences to the
blade tips as common in prior art are highlighted.
Tip clearance between the blade tip of a rotor and the casing is necessary for
a free rotation
of the rotor blade row. The gap however allows fluid to cross the blade tip
from the pressure
side of the blade to the suction side due to the pressure difference on the
pressure and the
suction side. This flow is associated with two main problems. Firstly, roughly
one third of all
the aerodynamical losses in a rotor row are related to the tip leakage vortex,
which forms
when the tip leakage over the blade tip enters the passage flow again on the
blade suction
side. It creates both mixing loss when it mixes out with the main flow and
perturbs the pres-
sure field on the blade tip wall that is responsible for the blade lift.
Furthermore, the fluid
is crossing the gap is not turned by the blade and therefore no work is
extracted from it. It is
therefore interpreted as lost work extraction. Secondly, the fluid crossing
the tip clearance
has a relatively high temperature due to hot streak migration, resulting in a
high thermal
load for the blade tip. In fact, blade tips burn away if not adequately cooled
and are hence
one of the limiting factor for the blade lifetime.
2o Additionally, it is desirable to minimize the tip clearance gap height in
order to improve the
performance through reduction of the tip leakage mass flow. This reduced gap
height, how-
ever, increases the risk of the rotor blade rubbing at the casing sometime
during the opera-
tional envelope. This can occur for example if the rotor expands further than
the casing due
to transients, a rotor dynamic excursion, an ovalization of the casing, or
through casing
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thermal distortions. In the case that a blade with a flat tip rubs severely at
the casing, catas-
trophic coolant loss could occur if the tip wears off. Even in a case of a
relatively minor rub
for a flat tip, any cooling holes located on the tip may be damaged resulting
in an inade-
quate cooling eventually leading to blade tip burnout.
Unlike for the case of a flat blade tips, the more complex flow physics for a
recessed blade
tip is more difficult to understand it's complexity. Also systematic design
procedures for cav-
ity size and shape are not available. By research it was possible to overcome
this problem
and to better understand aerodynamics and heat transfer physics of recess
cavities and to
provide new design boundaries for a standard, highly loaded rotor blade
representative of a
high pressure turbine. A special three dimensional CFD tool has been
extensively used for
this purpose.
NOMENCLATURE
Cartesian coordinates X (axial), Y, Z
is Cylindrical coordinates R (radial), 0, X
Blade axial chord Cax =XLEblade - XLE blade
Tip clearance height h= RB1adeTip -RHub
RHub - RCasing
Recess length L
Recess depth D
Recess rim thickness t(L,D)
Static pressure ps
Relative Total pressure pt,
Mass Flow m= f c- dA
A
Normalized quantity to flat tip Q* = Q
QFLAT TIP
Total pressure Loss coefficient Cpt = ptr in -ptr out
Ptr out - ps out
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K-I
Efficiency 77m =~w M~ CpTn 1- p` u` K
m Pt,in
The invention as such and computational tools to improve the results are
described in gen-
eral in accordance with the drawings. Because not available on the marked the
computa-
tional tools for pre-processing and the solver have been developed by the
inventors. These
tools may interact in parts with commercial products for post-processing.
In order to perform the intended computational and experimental study, a
previously de-
signed axial turbine test case has been utilized. The geometry of the one-and-
1 /2-stage,
unshrouded turbine models a highly loaded (DH/U2 = 2.36), low aspect ratio gas
turbine
environment. The air-loop of the test rig is of a quasi-closed type and
includes a radial com-
pressor, a two-stage water to air heat exchanger and a calibrated venturi
nozzle for mass
flow measurements. Before the flow enters the turbine section, it passes
through a 3 Meter
long straight duct, which contains flow straighteners to ensure an evenly
distributed inlet
is flow field. Downstream of the turbine the air-loop is open to atmospheric
conditions. A DC
generator absorbs the turbine power and controls the rotational speed of the
turbine. An
accurate torque meter measures the torque that is transmitted by the rotor
shaft to the gen-
erator. The TET (turbine entry temperature) is controlled to an accuracy of
0.3% and the
RPM (Rounds Per Minute) is kept constant within 0.5 min-' by the DC
generator. More
information on the turbine design (Behr et al. [13]) as well as on the
operation of the ex-
perimental facility (Sell et al. [14]) can be found in the open literature.
The following table shows the main parameter of "LISA" 1.5-stages axial
turbine research
facility at design operating point (Table 1):
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Turbine
Rotor speed [RPM] 2700
Pressure ratio (1.5-Stage, total-to-static) 1.60
Turbine entry temperature (TET) [ C] 55
Total inlet pressure [bar abs norm] 1.4
Mass flow [kg/s] 12.13
Shaft Power [kW] 292
Hub/Tip Diameter [mm] 660/800
1 st Stage
Pressure ratio (1 st Stage, total-to-total) 1.35
Degree of reaction [] 0.39
Loading coefficient y=Dh/u2 [-] 2.36
Flow coefficient f=cx/u [] 0.65
A computational design optimization for a nominal recess cavity commonly used
in axial
turbine rotor blades has been presented. From extensive parametric study, an
improved re-
cess cavity design is presented. Extensive aerodynamic and heat transfer
comparisons be-
tween the new design and the flat tip blade and the nominal recess tip are
presented. The
computational data was compared to experimental data of the Swiss Federal
Institute of
Technology (ETHZ) where the 3D flow structures and the performance of rotor
blades with
flat tip and the new recess design were measured. Qualitative comparisons to
the experi-
mental data from OSU have also been used to validate the predicted heat
transfer data. The
following concluding statements can be drawn from this study.
= A better understanding of the three dimensional flow inside recess cavities
was gained.
Three cavity vortices were found to govern the leakage flow through the
cavity.
= Change of the cavity geometry influences the generation and the interaction
of the
main recess vortices. A particular recirculation at the suction side front
responsible for
is high heat transfer could be eliminated and lead to a new design with
improved heat
transfer behaviour.
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= The beneficial effect of creating an aerodynamic seal has been shown for
both recess
designs. Tip leakage mass flow could be lowered by as much as 25 % in the new
recess
design according to the present invention compared to the flat tip. CFD showed
in-
creased power output for the new design too. Experimental measurements showed
a
5 0.3% increase in the turbine efficiency between a flat tip and the new
recess tip at de-
sign point.
= The heat load on the blade tip is found to be a balance between the heat
load on the
different blade tip components, i.e. the tip rim, the cavity rim walls, the
cavity bottom
and the rear flat blade tip. The new recessed design is about 7% lower on the
overall
10 heat load compared to the base line recessed design and 15% lower compared
to the
flat tip.
= To the best of the author's knowledge, this is the first time that detailed
profiling of
blade tip recess cavity is shown to improve performance and reduce heat load.
= Three-dimensional geometric profiling of blade tip recess cavity walls
significantly im-
is proves overall efficiency and effectively reduces in the same time both
heat load at the
blade tip and mechanical stress.
= Three-dimensional geometric profiling of recess cavity walls can be achieved
by using
non-uniform rim wall thickness as well as three-dimensional shaping of the
inside of the
cavity recess volume.
= Three-dimensional geometric profiling occurs through varying the recess rim
thickness
and cavity depth either separately or combine.
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= Three-dimensional geometric profiling of recess cavity optimize leakage flow
and its
interaction with vortices within the cavity and suppresses vertical flow
formation leading
to loss in the cavity through flow and high heat load on the cavity walls.
= Three-dimensional geometric profiling of blade tip recess cavities showing
the above
features attenuates and suppresses the secondary vortex formation and by
restraining
the available space required for vortex formation both circumferentially and
radially.
= Three-dimensional profiling of recess cavity walls leads to higher work
output of the
blade because of additional blade surface of the cavity walls in case of a
favourable cav-
ity pressure gradient.
= Three-dimensional geometric profiling according to the above features define
an opti-
mum cavity volume to combine.
= Three-dimensional geometric profiling exhibiting the above features reduces
the number
of circumferentially aligned vortices.
= Three-dimensional geometric profiling exhibiting the above features leads to
less dissi-
pative vortex patterns inside the cavity.
= Three-dimensional geometric profiling of blade tip recess cavities showing
the above
features keeps the advantages of non-profiled, constant rim thickness
cavities, such as
reduced rubbing surface area, cooling hole protection, lower mechanical
stresses.
= Three-dimensional geometric profiling of blade tip recess cavities showing
the above
features is applicable for unshrouded and certain partially shrouded turbine
blades.
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= Three-dimensional geometrically profiled tip recess cavity walls showing the
above fea-
tures provide passive flow control of leakage flow by acting as a throttling
mechanism.
= Three-dimensional geometrically profiled tip recess cavity walls based on
the three-
dimensional flow structure showing the above features can be used to re-design
the
blade tips of both existing blades for the upgrade and replacement blades as
well as be-
ing incorporated into new design configurations.
The invention is directed to a turbine blade with a side surface having an
aerodynamic pro-
file. The turbine blade has an end surface arranged in a mounted position in a
turbine de-
limited by a gap from a casing of the turbine. The end surface comprises a
recess which is
shaped such that it acts as an aerodynamic seal and/or reduces the blade tip
heat load. The
recess is delimited by a side wall which has a variable wall thickness in
circumferential direc-
tion. The side wall may have an in general constant wall thickness in height
direction (in the
direction of the length of the turbine blade). In certain embodiments the wall
thickness of
the side wall has an overall maximum in the area where the turbine blade has
its maximum
thickness. In a further embodiment the wall thickness of the side wall has a
maximum on
the suction side of the turbine blade between +0% and +50% relative length
with respect
to the blade leading edge (trailing edge +/-100%). In certain embodiments this
maximum
is an overall maximum. However, in certain other embodiments the wall
thickness of the
side wall has on the pressure side of the turbine blade a minima. Depending on
the field of
application this may be located in the range between -0% and -70% relative
length with
respect to the blade leading edge stagnation point (0%). The wall thickness on
the pressure
side may be at least partially constant. Good results may be achieved in that
the overall
maximum of the wall thickness of the side wall is about 5.5 to 6.5 times
bigger then the
overall minima. In an embodiment the recess has an in general constant depth.
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BRIEF DESCRIPTION OF THE DRAWINGS
The herein described invention will be more fully understood from the detailed
description
given herein below and the accompanying drawings which should not be
considered limit-
ing to the invention described in the appended claims. The drawings show in a
simplified
and schematic manner
Fig.1 an example of a three dimensional computational grid turbine rotor blade
with
a recess cavity;
Fig. 2 an example of a recess cavity design according to the prior art;
Fig.3 a three dimensional CFD flow distribution over flat tip blade as known
from the
prior art;
Fig. 4 a three dimensional CFD flow distribution around a first embodiment of
a tip
with a recess according to the prior art;
Fig. 5 a two dimensional CFD flow distribution in the recess according to
Figure 4;
Fig. 6 a three dimensional flow distribution around a blade tip with a recess
according
is to the invention;
Fig. 7 the flow distribution according to Figure 6 in a cut view;
Fig. 8 a pressure side CFD normalized tip mass flow;
Fig. 9 CFD Predicted Tip Rim Static;
Fig. 10 Suction Side CFD Predicted Normalized Tip Mass Flow;
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Fig. 1 1 CFD Blade Tip Nusselt Number Distribution;
Fig. 12 CFD Predicted Normalized Heat Load for Flat Tip, Nominal Recess and
New
Recess;
Fig. 13a Experimental Relative Total Pressure Coefficient Distribution at 14%
axial chord
downstream Rotor Blade Trailing Edge For FlatTip Blade;
Fig. 13b CFD Predicted Relative Total Pressure Coefficient Distribution at 14%
axial
chord downstream Rotor Blade Trailing Edge, Flat Tip Blade;
Fig. 13c Steady CFD Prediction vs pitch averaged Experimental Relative Yaw
Angle at
14% axial chord downstream Rotor Blade Trailing Edge, Flat Tip Blade;
io Fig. 14a Experimental Relative Total Pressure Coefficient Distribution at
14% axial chord
downstream Rotor Blade Trailing Edge, New Recess Tip Blade;
Fig. 14b CFD Predicted Relative Total Pressure Coefficient Distribution at 14%
down-
stream Rotor Blade Trailing Edge For New Recess Tip Blade;
Fig. 14c CFD Predicted vs Experimental pitch averaged Relative Yaw Angle at
14% axial
chord downstream Rotor Blade Trailing Edge For New Recess Tip Blade;
Fig. 15 experimental pitch averaged relative yaw angle at 14% axial chord
downstream
rotor blade trailing edge for new recess tip blade
Fig. 16 shows a first graph of a wall thickness distribution of a recess in a
turbine tip;
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Fig. 17 shows a second graph of a wall thickness distribution of a recess in a
turbine
ti p.
DESCRIPTION OF THE EMBODIMENTS
5 Subsequent embodiments of the invention are described in more detail.
Similar features are
in the different drawings marked with the same numbers.
Figure 1 shows a computer model of two turbine blades for numerical
calculation of the
behaviour of a recess arranged at the tip of a turbine blade.
The analytical models applied are in general based on a numerical grid (mesh)
of a turbine
10 blade 1. The numerical grids used were generated with an in-house developed
grid genera-
tor called MELLIP. A multi block structured grid generator uses a two
dimensional NURBS
library as input data to mesh the computational domain boundaries. Using a set
of geomet-
rical transformations the interior block boundaries are defined according to
the intended
grid topology. High grid quality, i.e. smooth gridlines, limited aspect ratio,
skewness and cell
is to cell ratios are achieved using both non linear interpolation algorithms
with flexible clus-
tering specification and two dimensional Poisson type elliptic partial
differential equations
during the meshing of each block. Several topologies are implemented and
partition the
computational domain for a blade tip with recess area in 18 blocks for the
blade recess
case. Especially the use of up- and downstream wake blocks with adjustable
sizes and grid
density helps to keep low grid skewness in the trailing edge area and prevents
the numerical
diffusion of the shed wake. The grids used for this study show a high
resolution in the blade
tip area in order to capture the flow gradients in this region.
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This helps in keeping the number of grid points at about 900'000 points, since
clustering
near walls does not need to be as aggressive as in two layer turbulence model
computa-
tions. Hence the high number of grid points in the blade tip area leads to
homogeneous
mesh density with smooth cell to cell ratios distribution. The densely packed
tip region grid
block spans over about the top 10% the blade span.
A numerical flow solver preferably used is called MBStage3D, a three-
dimensional, struc-
tured Navier-Stokes solver for multistage turbomachinery applications. The
time marching
algorithm preferably used in MBStage3D is a Jameson-type algorithm, i.e. an
explicit
method with a residual-averaging technique applied for improving stability.
The time discre-
tization is preferably accomplished by a five stage Runge-Kutta technique,
which is of
fourth-order accuracy. All computations discussed here were conducted with the
algebraic
Baldwin-Lomax turbulence model together with the Sommerfeld logarithmic wall
function to
compute the turbulent viscosity at the wall.
Extensive post processing necessary to gain understanding of flow physics is
achieved
is through 3D, 2D, 1 D, and scalar investigation of the flow fields of
interest. The 3D visualiz-
ing is done with TECPLOT, a collection of integrators and the 3D data
generation subrou-
tines for TECPLOT are developed in-house.
Figure 2 schematically shows a common turbine blade 51 as known from the prior
art. The
Turbine blade 51 has an outer side surface 57 with an aerodynamic profile. It
further com-
prises an end surface 56 which is in a mounted position in a turbine arranged
at a distance
to a casing 58 of the turbine (see Figure 5). The end surface 56 comprises a
recess 52 which
is shaped such that it acts as an aerodynamic seal. The recess 52 at the blade
tip 53 has a
length L and a uniform depth D. The wall 54 surrounding the recess 2 has a
constant thick-
ness t in vertical and circumferential direction. An investigation of the
three dimensional
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flow-field around the blade tip 53 and the recess 52 having a length L of 80%
of the axial
chord (depth of wing) and a depth D of twice the tip gap height allowed to
identify the flow
features in an exemplary manner.
The geometry of the tip recess 52 and its impact on the distribution of the
flow field around
the blade tip 53 was investigated based on a standard (nominal) recess design
52 (see Fig-
ures 2, 4, 5) to find further improved recess designs. The main geometrical
parameters var-
ied were the length L of the recess cavity, the depth D of the cavity and the
shape of the
recess rim which most generally can be any function of the two former
parameters. However
the shape was mostly determined as a function of the length.
io The cavity walls of the recess 2 (see Figure 6) offer additional surface
which generate addi-
tional power that needs to be added to the main power generated by the outer
blade wall.
It was also observed that the tip leakage flow for a recessed cases blade tip
is lower than for
a flat tip. The overall mass flow through the computational domain remained
unchanged,
suggesting that the change in the leakage mass flow occurred prior to the
suction side
is throat region, hence resulting in a constant corrected flow through the
rotor. It is important
to also note, that variations in total pressure loss coefficients followed
those of the tip gap
mass flow. The relation between aerodynamic loss and tip leakage mass having
been veri-
fied, necessitated further reduction of the tip leakage mass flow as an
important design
criteria. The changes with depth were strongly non-linear, with an optimum
depth leading to
20 minimum tip leakage mass flows being identified. This is also an important
difference to the
tip leakage mass flow evolution in a flat tip case where the tip leakage mass
flow varies
linearly with the gap height. The changes in recess length were almost linear.
The understanding of the detailed flow physics is therefore particularly
important in the
design process and will be investigated next on behalf of three major test
cases. The first
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case is a flat tip blade. The second test case is a nominal recess cavity as
known from the
prior art with a length of 80 % axial chord and twice as deep than the tip
gap, the rim
thickness being kept constant. This test case represents a current design for
recess cavities.
The final test case is the newly designed recess geometry based on the
extensive physical
modelling and geometric perturbation.
Figure 3 shows in an exemplary manner a flow field around a common turbine
blade 100
with a flat tip 101 with two main flow structures. The first one is the tip
passage vortex 102
which forms when incidence driven fluid enters the tip gap from the suction
side at leading
edge and leaves it again after approx. 20 % axial chord. Furthermore tip
leakage occurring
on the pressure side between leading edge and approx. 15 % axial chord crosses
the tip gap
and mixes with the incidence flow in the tip passage vortex. The feeding of
the tip passage
vortex is well organized, with the flow past a dividing streamline triggering
the formation of
the tip passage vortex. The succeeding pressure side leakage flow feeding the
tip passage
vortex can also be identified.
The second main flow feature observed is the tip leakage vortex 103 that forms
from the tip
leakage flow crossing the gap from the pressure side starting at approx. 15 %
axial chord.
The dividing streamline between the pressure side leakage feeding the tip
passage vortex
and the tip leakage vortex outer fluid layer can be identified. The outer
fluid layers in the tip
leakage vortex 103 result from the main part of the pressure driven, low gap
shear loss gen-
erating leakage jet. The tip leakage vortex core is formed by blade tip
boundary layer fluid.
A cutting plane orthogonal to the blade mean camberline reveals the well known
gap flow
structure. When the tip leakage flow enters the gap from the pressure side, a
separation
bubble is formed, leading to a vena contracta. The leakage jet leaving the
vena contacta
would then form the wake fluid in the lower part of the gap. This wake creates
mixing loss
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and is found later in the tip leakage vortex core. The leakage jet above the
wake part is of-
ten modelled as an isentropic jet, it forms the outer fluid layers around the
tip leakage vor-
tex core depending on the axial position when it left the gap on the suction
side.
Figure 4 schematically shows a flow structure around the turbine blade 51
according to
Figure 2 with the recess 52 at its tip 53. The recess 52 has in a
circumferential and in verti-
cal direction (in the direction of the length of the turbine blade) over most
of its length -
with exception to the trailing edge 62 - an in general constant wall
thickness. It should be
noted that the flow structure is dependent on the aerodynamic design of the
turbines, vary-
ing somewhat for different turbine blade rotor designs and may therefore not
exactly match
the following descriptions. Despite of this, it is believed that many of the
flow features in
general remain the same for modern axial high work turbines, with the
sensitivities and the
trade-studies relating to geometrical variation remaining applicable. It was
found that in
total six main flow features influencing the cavity flow. Starting at the
leading edge 60 of
the pressure side, fluid above the blade leading edge stagnation point 61
fluid of the pas-
i s sage enters the cavity 52. It crosses the cavity with low loss and at the
flow angle of fluid at
leading edge, and impinges on the corner between the cavity bottom and cavity
suction side
wall shortly following the peak suction. After hitting the corner wall this
pressure side lead-
ing edge jet rolls up into a vortical structure C and moves downstream inside
the cavity 52,
partly leaving the cavity 52 and entering the suction passage flow again. The
boundary
layer on the recess rim entering the cavity along the whole pressure side 20
due to the tip
clearance pressure gradient and on the first 20% axial chord on the suction
side immedi-
ately rolls up into a vortex A in the corner between the cavity bottom and the
cavity walls,
stretching along the whole pressure side 20 to the end of the cavity and
reaching up to
20% of axial chord on the suction side 21 cavity wall. After 10% of the axial
chord, this
suction side part of the vortex C is lifted off the cavity bottom by an
incidence driven suction
side leakage jet entering the cavity. A third important flow feature is vortex
B which forms
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when casing boundary layer fluid rolls up against the pressure side tip
leakage jet TL. This
vortex stays on the casing wall and deflects the pressure tip leakage inside
the cavity as
shown in Figure 5. Between the above two vertical flow structures, a dividing
streamline
DS1 establishes.
5 Downstream of the 20 % axial chord, the flow behaviour of the pressure side
leakage is
similar to the flat tip case with the difference that the leakage is deflected
by the cavity
vortices and interacts with them. After leaving the gap on the suction side,
this fluid forms
again the outer layer of the tip passage vortex and the tip leakage vortex.
The core of the tip
passage vortex is formed by the same incidence tip leakage that lifts off the
cavity corner
io vortex when entering and leaving the cavity between 10% and 20% axial
chord. The core of
the tip leakage vortex is wake fluid behind the separation bubble on the
suction side rim
that forms when the pressure side leakage jet leaves the cavity.
To additionally clarify the flow features inside the cavity a cutting plane
orthogonal to the
camberline located downstream of the formation of the vortex formed by the
pressure side
is leading edge jet is shown in Figure 5. The three main cavity vortices above
are referenced.
Between the casing vortex B and the rim boundary layer roll up vortex A the
pressure side
leakage crosses the cavity lifting up vortex C caused by the pressure side
leading edge cavity
jet. Two separation bubbles form, one at the pressure side rim edge to the
outer blade wall
when the tip leakage jet enters the gap, and the other on the suction side rim
edge with the
20 cavity wall when the tip leakage leaves the cavity again. The tip leakage
jet is recognizable
as a low entropy zone between the higher entropy zones where the vortices are
located. The
casing vortex is rather squeezed, this explains why downstream of the blade
gradually more
and more fluid from the neighbouring vortices mix in.
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The nominal design showed many vortical structures inside the recess cavity.
Particularly the
front part of the cavity is affected by these structures. As seen above, the
boundary layer
fluid leaking from the rim into the cavity rolls up in a vortex along the
corner between cavity
bottom and cavity rim wall. The aim of the new design was to eliminate the
recirculation
zone in the front part of the blade to minimize aerodynamic losses and reduce
the head
transfer coefficient.
Figure 6 shows a flow distribution around a tip 3 of a of turbine blade 1
according to the
present invention. The recess 2 is surrounded by a wall 4 which has in
circumferential direc-
tion a variable wall thickness t. Whereby the wall 4 has a maximum thickness t
on the suc-
tion side 21 in the area where the turbine blade 1 has the maximum over all
thickness. A
side surfaces 5 of the recess 2, respectively the wall 4, is here arranged in
general parallel to
the length axis of the turbine blade 1(in general perpendicular to an end
surface 6 of the
turbine blade 1). However, depending on the field of application, alternative
designs may be
appropriate. One advantage consists in that the shown design is relatively
easy to make by
is standard grinding processes.
One effect of the improved design is that the streamline that separates the
recirculation
zone from the pressure side leading edge jet is moved. In Figure 6 the effect
of the im-
proved design on the cavity flow pattern is shown. The recirculation on the
leading edge 10
is suppressed. The pressure side leading edge jet now spreads inside the whole
cavity 2. The
casing boundary layer fluid rolls up into vortex B and interacts with the
fluid from the pres-
sure side leading edge jet. The fluid trapped in the recirculation in the
nominal design now
enters the cavity and is pushed out again by the pressure side leading edge
jet. After leaving
the cavity it feeds the tip passage vortex.
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Figure 7 illustrates the vortical flow pattern inside the cavity 2 and around
the tip 3 of the
turbine blade 1 along a cutting plane parallel to the length direction of the
turbine blade 1.
As it can be seen the end surface 6 of the turbine blade 1 is delimited by a
gap 12 from the
casing 8 of the turbine. The vortices A and B are actively interacting.
Therefore vortex B is
not confined to the casing anymore but can occupy the whole cavity volume.
Vortex C is
formed by the fluid that led to the suction side leading edge recirculation in
the nominal
case. It is formed when this fluid separates on the cavity rim while being
pushed out of the
cavity by the pressure side leading edge jet.
Subsequent the results regarding aerothermal performance of three test cases
are intro-
duced. For the aerodynamic performance, the tip leakage mass flow is
investigated since it is
intensively related to the total pressure loss. Nusselt Number distribution
and integrated
heat flux vector on the blade tip walls are compared to assess the impact of
the new design
on heat transfer.
Figure 8 shows the variation of accumulated tip gap mass flow from leading to
trailing
is edge for the pressure and the suction side for the three investigated test
cases are shown
separately. In the diagram x means length, c_ax the axial cord and mdt* the
leakage mass
flow over the blade tip as a function of the axial distance. On the pressure
side, the most
important feature is the growth of the accumulated tip gap mass flow from
leading edge to
trailing edge for the flat tip compared to a linear increase as long as the
recess cavity opens
behind the rim. Downstream of the cavity trailing edge the accumulated tip gap
mass flow
also varies non-linearly and is similar to the flat tip case for in both
recess designs. The lin-
ear increase may be explained through the average static pressure variation at
the tip gap
entry on pressure side and its exit on the suction side.
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As shown in Figure 9, the static pressure decreases both for the pressure and
the suction
side non-uniformly in the flat tip case. The diagram shows the static pressure
variation as a
function of the axial length. For the recess cases however static pressure
remains at constant
level whenever tip leakage enters the cavity, which is the case for the entire
pressure side
gap but also for the front part of the suction side, where incidence fluid
enters into the cav-
ity. The recess cavity acts like a reservoir where pressure remains constant.
The sealing effect
and the resulting reduction of the accumulated tip gap mass flow is clearly
observed. The
nominal recess case showed a reduction of 23% in the leakage mass flow when
compared
to the flat tip, the new design had 25% less mass flow crossing the gap
compared to the
io flat tip.
Figure 10 shows the accumulated tip clearance mass flow on the gap exit for
the blade
suction side is shown for a flat tip blade (prior art), the a recess with a
constant wall thick-
ness (nominal recess, prior art) and an improved recess with variable wall
thickness (indi-
cated as new design, according to the present invention). The incidence driven
tip leakage
is mass flow is much more intense in the recess cases than in the flat tip
case. Incidence leak-
age reaches for all three cases from leading until 22 % axial chord. However
the amount of
mass flow entering the recess cavities is almost the double of the one for the
flat tip. This
sustains again the cavity in acting as a reservoir to be filled up with fluid.
The difference in
the amount of mass flow additionally having entered the cavity compared to the
flat tip gap
20 almost exactly matches with the reduction in total cumulated tip leakage
mass flow over the
entire suction side. No large differences in tip gap mass flow are noted
between the nominal
and the new recess designs.
Figure 11 shows the heat transfer at the tip of three different turbine
blades. While Figure
11 a shows in a comparative manner the Nusselt Number (Nu Number) distribution
of a
25 common turbine blade with a flat end surface as shown in Figure 3, Figure
11 b and 11 c
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show the distribution at turbine blades with recesses according to the prior
art (indicated as
"nominal") and the present invention (indicated as "new"). High Nusselt
Numbers occur on
the leading edges in all three cases. This is where hot fluid meets the blade
tip first and heat
transfer is highest. On the pressure side edges Nusselt Number has the same
magnitude in
all three cases. The thin rim of the nominal recess shows a similar
distribution on the suction
side front part compared to the new design. When the tip leakage vortex at
forms (about 25
% axial chord), the value of the Nusselt Number drops for both the flat tip
case and the
nominal recess case. This observation can however not be made for the suction
side rim of
the new design. Both the flat tip and the nominal recess show high Nu Numbers
at the lead-
ing edge. In the new design however, Nu Number values do not reach as high on
the cavity
bottom. From Figure 11 the differences between the improved design with a
recess accord-
ing to the present invention (indicated as "new") and a flat tip (indicated as
"flat") is visual-
ized. The blocking of the suction side recirculation zone by a thicker rim has
not lead to
higher Nusselt Numbers on the rim.
is Figure 12 visualizes the heat load of different turbine blades. The
integrated heat flux vec-
tor on the blade tip of a common flat blade tip (indicated as "flat") and a
blade tip with a
recess according to prior art (indicated as "nominal") and the present
invention (indicated as
"new") gives the heat load for the three test cases.
From Figure 12 where heat load is split according to the affected wall type
the following
observation can be made. The overall predicted tip heat load is highest for
the flat tip and
lowest for the improved design with a variable wall thickness ("new" design).
The reduction
of heat load between the flat tip and the new design (about 14%) is about
twice as high
than between the nominal and the new recess design (7%). However it must be
noted, that
the new design also has higher regions of heat load than the nominal design.
This is the
case for the tip rim and the cavity walls to which the cavity rim walls and
the remaining rear
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flat blade tip portion belongs. Increasing the rim thickness, as well as a
deeper and shorter
cavity are the reasons for this increased heat load. However it has also been
shown that
changing the flow field on the leading edge inside the cavity through the
elimination of the
suction side recirculation zone has proved effective in reducing heat load on
the cavity bot-
5 tom.
In the Figures 13 and 14 the computed results from CFD predictions (see
Figures 3-12) are
compared to experimental results which have been performed in a confidential
manner at
the Swiss Federal Institute of Technology (ETHZ). The flat tip and the new
recess design
blades have been experimentally evaluated at ETHZ axial turbine facility LISA.
10 Figure 13a shows the experimental relative total pressure coefficient
distribution at 14%
axial chord downstream rotor blade trailing edge for flat tip blade. First the
predicted and
experimental relative total pressure loss coefficients for the flat tip and
the new recess de-
sign are compared in 2D axial cutting planes located 14% axial chord
downstream of the
rotor trailing edge. The experimental data for the flat tip blade shown in
Figure 13a are
15 snapshots at a given point in time of an unsteady flow. Therefore unsteady
flow features are
resolved which are not present in the single row steady state CFD results. The
fact of un-
steady data also explains the modulation of the low relative total pressure
zones identifying
the secondary flow vortices and the tip leakage vortex.
Figure 13b shows the CFD predictions from a single row steady state
computation. The
20 computationally predicted relative total pressure loss coefficient resolves
in sufficiently good
accuracy the secondary flow structures measured by the experiment. The hub and
tip pas-
sage vortex are captured both in their spatial extension and in the loss
magnitude. Also the
trailing edge wake of the rotor blade is captured in the CFD results. The loss
region associ-
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ated to the tip clearance vortex is however over predicted compared to the
experiment. As
was mentioned above
Figure 13c shows CFD predicted and measured pitch averaged radial
distributions of rela-
tive flow yaw angle at 14% of the axial chord downstream of the rotor blade
trailing edge
for the flat tip blade. It can be seen that the variation in the relative yaw
angle due to the
tip leakage vortex between 80% span and 100% span is largely over-predicted by
the com-
putational result. The magnitude of relative yaw angle variation due to the
hub and tip sec-
ondary flow structure is well predicted.
Figures 14a and 14b show experimental and CFD predicted data for the turbine
blade with
an improved recess according to the present invention (indicated as "new"
design) with a
variable wall thickness in circumferential direction are compared. The
experimental data
shown in Figure 14a are again a snapshot of unsteady data. The snapshots for
the flat
blade tip and the new recess design were both taken at the same point in time.
The CFD
predicted relative total pressure coefficient from steady state computations
are shown in
is Figure 14b. It can again be seen that the predicted CFD results resolve the
same features
that are also captured by the measured flow field. Predicted relative total
pressure loss coef-
ficients showing the hub and tip passage vortices agree well with the
experimental data.
The over prediction of the tip leakage vortex noted in the flat tip case is
also found in the
new recess design case. Compared to the computationally predicted relative
total pressure
loss coefficient for the flat tip case, it can be noted that the spatial
extension of the tip
leakage vortex loss core for the new recess design has been reduced.
Figure 14c shows the measured and predicted pitch averaged relative flow yaw
angle dis-
tributions for the new design. The CFD predicted relative flow angle
distribution matches
the experimental data up to about 80% of the span. The part from 80 % span to
100%
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span is influenced by the tip leakage vortex. Whereas the secondary flow
features on hub
and tip are correctly predicted, the difference in flow angle due to the tip
leakage vortex is
very much over predicted.
In Figure 15 the experimentally measured relative yaw angle distributions for
the flat tip
and the new recess design are presented for the tip region from 60% span to
the casing at
100% span. It can be seen that the new recess design shows less over turning
than the flat
tip blade. This result clearly illustrates that the recess cavity influences
the tip leakage vortex
which is responsible for the stated overturning.
Experimentally measured performance data shows that for the turbine used the
"new" recess
design has a 0.3 % total efficiency when compared to the flat tip at exactly
the same over-
all turbine operating conditions. The predicted difference between both
efficiencies was
0.38 %, which is a good quantitative agreement with the experimental data.
The predicted heat transfer data is qualitatively compared to data presented
by the Ohio
State University Gas Turbine Laboratory [12]. A turbine blade with a recess
cavity similar to
the nominal design presented here was equipped with heat transfer gauges to
measure heat
transfer on the cavity bottom near leading edge, trailing edge and in the
middle. Also the
rim was equipped with several gauges. Nusselt Numbers were reported for
different
vane/blade spacings. The trend in the variation of Nu Number according to the
investigated
location is similar. The highest Nu Number is found in the leading edge
region. The second
heat flux gauge was positioned 12.5 % blade axial chord from peak suction
downstream,
reporting almost half of the Nu Number at leading edge. This can also be
observed in the
nominal recess case from Figure 11. Finally the third heat gauge at 62% blade
axial chord
corresponding to 80% cavity axial chord reports almost the same Nu Number for
all vane
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spacings that was reported for the low vane spacing on the previous one for
the short spac-
ing.
Figures 16 and 17 are showing the distribution of the wall thickness in
circumferential di-
rection of several recesses of turbine blades according to the present
invention. Such that it
becomes possible to compare different shapes of recesses of turbine blades,
the graphs are
shown in a normalized manner in that on the y-axis the thickness t is shown in
relation to a
reference thickness tO. With reference to Figure 6, the local thickness t of
the wall 4 is
measured at the level of the end surface 6 of the turbine blade 1 along and
perpendicular
to an outer edge 13 of the turbine blade 1, which is formed by the side
surface 7 of the
io turbine blade 1 and its end surface 6. The local wall thickness t is shown
in the graph on the
x-axis starting from the blade leading edge stagnation point 0(0%
circumferential length)
in both circumferential directions, whereby the relative position of the
trailing edge 11 of
the turbine blade 1 is indicated by the values 1(100%), respectively -1 (-
100%). The suc-
tion side is positive (+) and the pressure side of the profile is negative (-
). The trailing edge
11, which is in the graph located approximately in the ranges between +0.6
(+60%) and +1
(+100%) and -0.6 (-60%) and -1 (-100%), is not considered in the diagram
because of the
reducing thickness of the turbine blade in this area. It can be seen that the
wall thickness t
increases on the suction side and has a (local) maximum 15 on the suction side
of the tur-
bine blade, depending on the field of application, in the range between +0.1
(+10%) and
+0.4 (+40%) relative length. In other embodiments the maximum 15 is arranged
at +11%,
19%, 30%, respectively 37% (aberration +/- 5%). The maximum 15 is in certain
embodi-
ments arranged in the area of the maximum thickness 14 of the turbine blade.
As it can be
retrieved from the shown graphs, certain embodiments have a shoulder 16 in the
range
+0.05 (+5%) and +0.15 (+15%), with a locally reduced augmentation of the
thickness (e.g.
see Figure 6). In an embodiment, the augmentation of the wall thickness
(indicated by ar-
row 17) already starts on the pressure side of the turbine blade, e.g. in the
range between -
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0.1 and 0. As it can be seen in the graphs, in certain embodiments the
increase and the
decrease of the wall thickness are in general similar to each other and in the
range of 300%
to 400% per 0.1 normalized circumferential length. As it can further be
retrieved from the
graphs, the maximum wall thickness is about 5.5 to 6.5 times thicker than the
minimum
wall thickness at the pressure side.
On the pressure side the wall thickness is in certain embodiments constant
which is indi-
cated by the in general horizontal progression of the graph in this area.
However, certain
embodiments (see profile 3) may also have a local maximum 18 at the pressure
side.
Although the present invention has been described in relation to particular
embodiments
thereof, many other variations and modifications and other uses will become
apparent to
those skilled in the art. It is preferred, therefore, that the present
invention be limited not by
the specific disclosure herein, but only by the appended claims.
