This invention relates to a closed circuit fluid cooled
gas turbine blade of new land based gas turbine in simple or
combined cycle configuration, which permits a user to incorporate
air or steam cooling of hot gas turbine parts with minimal change
in components, and which also incorporates design changes
enabling certain turbine components to be used without change
in both 50 and 60 Hz turbines. The invention here specifically
relates to cooling steam circuits for the gas turbine blades
in the first and second stages of a four stage combined cycle
gas turbine.
BACKGROUND
Gas turbine blades have historically used compressor bleed
air as the cooling medium to obtain acceptable service
temperatures. Cooling passages associated with this design
technology are typiclly serpentine arrangements along the mean
camber line of the blades. The camber line is the locus of points
between the low pressure and high pressure sides of the air-
foil. Adjacent radial passages are connected alternately at
the top and bottom by 180 degree return Ubends to form either
a single continuous passage or independent serpentine passages,
with the cooling air exiting into the gas path by one or a
combination of the following schemes (a) leading edge holes,
(b) hole exits along the trailing edge, (c) hole exits on the
high pressure side and low pressure sides of the blade airfoil,
and (d) tip cap holes.
Each radial passage typically cools both the high pressure
and low pressure sides of the blade airfoil. The specific
geometry of each radial cooling passage is designed to balance
the conflicting demands for low pressure drop and high heat
transfer rate. Schemes used in the state of the art............

to enhance heat transfer rate include raised rib turbulence promoters (also
known as trip strips or turbulators), passage crossover impingement, the use of
impingement inserts, and the use of banks or rows of pins. These schemes
increase the local turbulence in the flow and thus raise the rate of heat
transfer. The effectiveness of open circuit air cooling is further improved by
the coverage of the blade airfoil by an insulating film of air bled through
openings in the airfoil surface. The disadvantage of using compressor bleed
flow, however, is that it is inherently parasitic. In other words, turbine
component cooling is achieved at the expense of gas turbine thermodynamic
efficiency. Cooling schemes involving high pressure and high density fluids,
such as steam, on the other hand, have not yet been employed for blade
cooling or reduced to practice in commercially available gas turbines.
DISCLOSURE OF THE INVENTION
The object of this invention is to provide a turbine blade design which
can be used to operate under gas turbine conditions with very high external
combustion gas temperatures (about 2400° F) and high internal pressure
coolant supply conditions (600-1000 psi) typical of extraction steam available
from the steam turbine cycle of a combined cycle steam and gas turbine power
plant. Commonly owned co-pending application S.N. 5,685,639 (atty ^kt.
839-346) entitled "Removable Inner Turbine Shell With Bucket Tip Clearance
Control" discloses a removable inner shell which permits easy access and
conversion of stage 1 and 2 stator and rotor components from air to steam
Patent _ ,..,_ 07.
cooling. Commonly owned co-pending application S.N. ■'■^i^'1 (attv.
dkt. 839-358) entitled "Closed Or Open Circuit Cooling Of Turbine Rotor
Components" discloses the manner in which the cooling steam is fed to the
blades
stage 1 and 2 -buckets. Both applications are incorporated herein by reference.

This invention relates to the stage 1 and 2 turbine blades per se, and
seeks to maximize the thermodynamic efficiency of the gas turbine cycle by
using steam as the turbine blade coolant instead of air bled from the gas
turbine compressor for the first and second stages of the gas turbine, i.e., the
stages where cooling is most critical. In reaching the desired goal, the design
of closed circuit steam cooled blades and associated coolant passages is
determined in accordance with the following additional criteria;
1) minimum coolant pressure loss;
2) predictable and adequate heat transfer;
3) metal temperature consistent with part life objectives;
4) minimization of secondary flow effects; and
5) ease of manufacture.
By way of additional background, the high gas inlet temperatures
required to maximize gas turbine thermodynamic efficiency are sufficient to
melt metals used in gas turbine blade construction. The blades used in the
first few stages are cooled to prevent melting, stress rupture, excessive creep
and oxidation. The cooling must be judiciously applied to prevent premature
cracking due to low cycle fatigue. The continuing increases in gas turbine
inlet temperature, and the use of combined cycles to maximize the thermal
efficiency of power plants bring into consideration the use of steam as a
coolant for gas turbine hot gas path components.
The use of steam as a coolant for gas turbine blade cooling can provide
several advantages. One advantage is that of potentially superior heat transfer.
For example, when comparing typical high pressure extraction steam to
compressor bleed air, steam has an up to 70% advantage in heat transfer
coefficient in turbulent duct flow by virtue of its higher specific heat (other
considerations being equal). The more important advantage is higher gas
turbine thermal efficiency. Since the compressor bleed air is no longer needed

for cooling the first and second stages, it can be put to good use as increased
flow in the gas path for conversion into shaft work for higher turbine output
for the same fuel heat input. There are problems associated with steam as a
coolant, however, which stem from the requirement of maintaining a closed
circuit and the already mentioned high supply pressures typical of reheat
extraction in a steam power plant. In closed circuit cooling, the coolant is
supplied and removed from the shank of the blade, and a single serpentine
circuit is provided within the blade, including multiple radial outflow and
radial inflow passages.
Closed circuit cooling (as opposed to open circuit cooling typically used
when air is the cooling medium) is preferred because: (a) otherwise, large
amounts of make-up water would be required in the steam turbine cycle
(assuming a combined cycle configuration), and (b) it would be more
deleterious for thermodynamic efficiency to bleed and mix steam into the gas
path (as compared to air) because of steam's greater capability to quench and
reduce the work capability of the hot combustion gas because of steam's higher
heat capacity.
High coolant pressures are required because reheat steam is usually
extracted at high pressure to optimize steam turbine cycle thermodynamic
efficiency. Thin airfoil walls, usually required for cooling purposes, may not
be sufficient for the pressure difference between the internal coolant, steam,
and the gas path, resulting in excessive mechanical stresses. Steam pressures
may be in excess of 3-5 times typical compressor bleed air (e.g. 600-1000 psi
steam versus 200 psi air). A new design is thus required which can operate
under high heat fluxes and high supply pressures simultaneously.
Other problems arise from the high pressure and high density steam
used as a coolant. For example, the density of steam at 1000 psia is 3 times

the density of air at 200 psia (at the same temperature, for example, 800" F).
At the same time, the heac capacity of steam is roughly twice that of air under
the same conditions. This means chat lesser amounts of steam mass flow are
required for the equivalent convection cooling. The Buoyancy Number, Bo,
obtained from the ratio of the buoyancy to inertia force of the forced
convection flow is defined by the Grashof number divided by the Reynolds
number squared (Gr/Re2). With air cooled blades, undesirable buoyancy
effects are typically small, Bo < < 1. The buoyancy effects are greater with
steam, however, and as the buoyancy factor Bo approaches unity, the
undesirable effects become even more significant. The internal coolant
passages for a steam cooled system must therefore be designed to accounc for
Coriolis and buoyancy effects, also known as secondary flow effects, explained
in greater detail below.
More specifically, at the higher densities and low flow rates (lower
flow velocities for a given passage cross sectional area) of steam, the cooling
fluid in the internal blade cooling passages is more prone to develop secondary
flows from Coriolis and centrifugal buoyancy forces which (a) affect the
predictability of heat transfer and (b) impair the heat transfer by uneven heat
pickup or potential flow reversal. As the blade rotates abouc the shaft axis,
one side of the airfoil is ahead of the other in the direction of rotation. The
side of the airfoil which is ahead is the leading side and the one which is
behind is the trailing side. It is shown in the literature (for example, see
Prakash and Zerkle, "Prediction of Turbulent Flow and Heac Transfer in a
Radially Rotating Square Duct," Paper HTD-Vol. 188), that, when air is the
coolant, flow tends to move from the high pressure region near the leading
side to the low pressure region near the trailing side in the plane of the coolant
passage cross section. The effects are more severe when steam is the coolant.

It has also been determined that Coriolis and buoyancy forces or effects
are most significant in the radial outflow passages of the serpentine cooling
circuit, particularly in the region from the pitchline (halfway between the hub
blade
and the tip of the-bueket) to the cip of the bucket or blade. Accordingly, the
blade
focus in this invention is on the-bueket radial outflow passage design. Any
such design requires prior knowledge of the flow conditions which would set
up these adverse flow recirculations at which point, passage size, and shape
can be used to minimize any adverse effects.
The parameters which must be taken into account in any such design
process include:
a) mass flow rate of the combustion gases entering the gas turbine;
b) heat transfer coefficients of coolants;
c) surface area to be cooled;
blade
d) temperature of combustion gases at-buc-ket leading edge;
e) temperature of the *uifstt; and
0 heat flux.
• In addition, certain material limitations dictate certain aspects of the
design. For example, in one embodiment, the rotor itself dictates that the
temperature of the coolant exiting the turbine be no more than about 1050° F
due to the properties of Inconel, for example, of which the rotor is formed.
This, in turn, dictates that the steam coolant entering the turbine should be
about 690°-760° F (given a pressure of about 600-1000 psi). By the time the
steam coolant reaches the first and second stages of the turbine, the
temperature will be somewhat higher (about 1000° F) and the pressure
somewhat lower (about 700 psi).
In accordance with the anticipated operating parameters of this new gas
turbine, combustion gases are likely to enter the first stage at about 2400° F

and the maximum metai temperature needs to be reduced to below about 1800°
F. Corresponding second stage temperatures are likely to be 2000° F and
1650°.
With these conditions set. the mass flow of coolant and coolant passage
areas can be determined. At the same time, given a mass flow and inlet
temperature (T^) for the coolant, the passages can be designed to
accommodate (i.e., minimize) Coriolis and buoyancy effects.
The novel features of the turbine blade designs in accordance with this
invention are thus found in the blade cooling passages and the exclusive use of
high pressure steam as the blade cooling fluid in the gas turbine first and
second stages. The third stage remains air cooled and the fourth stage remains
uncooled in conventional fashion.
In a first exemplary embodiment, radial passages in the turbine blade
are configured in a single serpentine, closed circuit, with steam entering along
the trailing edge of the blade and exiting along the leading edge of the blade.
The number of radial inflow and outflow passages may be any number
depending upon the demands of the above design criteria. The radial passages
are connected alternately by 180 degree return U-bends and each passage
includes 45 degree angle raised rib turbulence enhancers.
[n a transverse cross section through the pitchline of the airfoil, the
radial outflow passages are made deliberately smaller than the radial inflow
passages, with the exception of the radial inflow (or exit) passage along the
leading edge of the airfoil. The reasons for this exception are explained
further herein.

The smaller radial outflow passages counteract the tendency for any
radial secondary flow recirculation resulting from centrifugal buoyancy forces
acting on the cooling fluid. This adverse tendency is counteracted by making
the bulk flow velocity as large as possible in radial outflow within the confines
of producibility and pressure drop. The radial outflow passages are designed
with aspect ratios (length to width cross-section dimensions for the passages),
such that buoyancy parameters lead to maximized heat transfer rate on the
leading side of the passage as substantiated by test results. The target regime
of operation in radial outflow is a Buoyancy Number of less than 0.15 or
greater than 0.8 for passages with an aspect ratio of 3.3 to 1. As already
noted above, it is known that the adverse effect of Coriolis and buoyancy
forces are more benign to radial inflow passages when air is used as the
coolant. (See, for example, Wagner, J.H., Johnson, B., and Kopper, F.,
"Heat Transfer in Rotating Serpentine Passages with Smooth Walls," ASME
Paper 90-GT-331, 1990.) We have confirmed that this is also the case for
steam. As such, the radial Inflow passages are kept relatively large within the
confines of desired heat transfer coefficients and pressure drop constraints.
The above embodiment also features the use of turbulence enhancing
raised ridges or trip strips to enhance the heat transfer rate. These features
have the additional benefit of reducing the adverse effects of buoyancy and
Coriolis forces as the local turbulence breaks up secondary flow tendencies.
This effect also has been documented (for air) in the literature (see, for
example, Wagner, J.H., Steuber, C, Johnson, B., and Yeh, F., "Heat
Transfer in Rotating Serpentine Passages with Trips Skewed to the Flow".
Rows of pins may also be used in trailing edge passages for both mechanical
strength and heat transfer.
Cooling the tip portion of a closed circuit cooled blade presents
additional problems. Typical high technology open circuit air cooled designs

bleed coolant near the tip to reduce the heat flux around the tip periphery of
the airfoil. The reduced heat fluxes reduce the temperature gradient through
the wall and the associated thermal stresses. In closed circuit cooling, the
mechanism for solving the problem is solely by internal convective cooling.
Tip cooling is addressed by incorporating raised ribs on the underside
of the blade tip cap. These ribs increase the local turbulence and thus enhance
the rate of heat transfer.
Another feature is the incorporation of bleed holes at the juncture where
the rib meets the wall and the tip cap. The aforementioned feature provides
relief from high thermal stresses by u neons training the corner region from the
relatively cold rib. The situation is further improved by chamfering or
radiusing the external corner at the juncture of the airfoil wall and the tip cap.
This reduces the effective wall thickness and reduces the temperature gradient
across the wall of the airfoil around the periphery of the tip cap.
In a variation of the above design, the flow is. reversed, i.e., the flow
moves radially outward through the leading edge passage and then follows a
similar serpentine arrangement, in reverse, exiting through the trailing edge
passage.
It has also been found that incorporation of the disclosed embodiments
in actual blade design may require coupling with a thermal barrier coating on
the blade outer surface to keep blade temperatures within acceptable limits.
In one aspect, therefore, the present invention may be defined as
comprising a gas turbine-kueket having a shank -pett-ion. a fadial tip-pottten
and an airfoil having leading and trailing edges and pressure and suction sides,
and an internal fluid cooling circuit, the improvement comprising the internal

fluid cooling circuit having a serpentine configuration including
plural radial outflow passages and plural radial inflow
passages,the radial outflow passages shaped avoid an undesirable
Buoyancy Number associated with the aspect ratio selected.
As an example, the radial outflow passages are shaped to
lave aspect ratios of about 3.3 to 1 and Buoyancy Numbers of
<^0.15 or^fl-SO.
In another aspect,the invention may be defined as comprising
a gas turbine blade having a shank, a radial tip and an airfoil
extending between the shank and the radial tip. the airfoil having
leading and trailing edges and pressure and suction sides, and
an internal fluid cooling circuit, the improvement comprising
the internal fluid cooling circuit having a serpentine
configuration including plural radial outflow passages and plural
radial inflow passages, the radial outflow passages having,
on average, smaller cross-sectional areas than the radial inflow
passages.
In still another aspect, the invention relates to a method
of determining a configuration for steam cooling passages for
blade stage in a gas turbine comprising the steps of:
a) determining combustion gas inlet temperature and mass
flow rate of combustion gases passing through the gas turbine
stage;
b) taking into account Coriolis and buoyancy flow effects
in the steam coolnt caused by rotation of the blade stage; and
c) configuring the radial inflow and outflow coolant
passages to have a size and shape to provide aspect ratios of
about 3.3 to 1 and Buoyancy Numbers of <^0.15 or ^0.8 in said
radial outflow passages.
The advantages which accrue from this invention can be
summarized as follows:

1. Closed circuit steam cooling using high pressure
steam achieves bulk cooling effectiveness greater than that of
open circuit air cooling.
2. Closed circuit steam cooling of turbine blades
increases gas turbine thermodynamic efficiency by eliminating
parasitic compressor bleed flow for turbine blade cooling.
3. The adverse effects of the rotational Coriolis and
buoyancy forces and possible flow reversal in outward flow
have been reduced through proper passage design for the flow
rate of coolant, particularly in the radial outflow passages.
4. The adverse effects of the rotational Coriolis and
buoyancy forces and possible flow reversal have been further
reduced by the use of turbulator ribs or trip strips.
5. A more even distribution of heat transfer rate around
the periphery of the coolant cavity has been maximized by the
passage design.
6. Regions of flow stagnation in the tip turnaround have
been eliminated by the use of turning vanes and/or raised rib
turbulators.
7. Tip cooling has been enhanced by use of raised rib
turbulators on the underside of the cap.

8. Thermal stresses at the outer periphery of the tip cap
are relieved by bleed holes which are placed at the juncture of
the rib, the airfoil wall and the tip cap.
9. The passages have been designed to maximize heat
transfer and sustain high internal pressures.
Advantages and benefits beyond those discussed above will become
apparent from the detailed description which follows.
ACCOMPANYING
TfTlJYJ flTpSf^ROPTION OF THE£)RAWINGS
FIGURE 1 is a schematic diagram of a simple cycle, single shaft,
heavy duty gas turbine;
FIGURE 2 is a schematic diagram of a combined cycle gas
turbine/steam turbine system in its simplest form;
FIGURE 3 is a partial cross section of a portion of the gas turbine in
accordance with the invention;
FIGURE 4 is a section through a typical turbine blade with internal
cooling passages;
FIGURE 4A is an enlarged, planar representation of a flow passage
from Figure 4, and illustrating secondary flow effects;
FIGURE 5 is a perspective view of a first stage turbine blade in
accordance with this invention;

FIGURE 6 is a perspective view similar to Figure 5 but broken away to
show internal cooling passages;
FIGURE 7 is a planar side view of the blade shown in Figure 5, with
internal passages shown in phantom;
FIGURES 8A-C are sections of a first stage gas turbine blade in
accordance with the invention, the sections taken at the hub, pitchline and tip
of the blade, respectively;
FIGURE 9 is a perspective view, partly in section, of a second stage
turbine blade in accordance with the invention;
FIGURES 10A-C are sections of a second stage blade, taken at the
hub, pitchline, and tip, respectively;
FIGURE 11 is a partial, enlarged section of a blade tip, illustrating
internal tip cooling in accordance with che invention;
FIGURE L2 is a view similar to Figure 11 but illustrating an alternative
blade tip cooling arrangement;
FIGURE 13 is a view similar to Figure 11 but illustrating another blade
tip cooling arrangement in accordance with the invention;
FIGURE 14A is a section through a blade illustrating bleed holes in the
passages dividers in accordance with the invention;
FIGURE 14B is a partial section taken along the line 14B-L4B in Figure
14A;

FIGURE 15 is a partial section of a first stage turbine blade in
accordance with another exemplary embodiment of the invention;
FIGURE 16 is a partial section of a first stage turbine blade in
accordance with still another exemplary embodiment of the invention;
FIGURE 17 is a partial section of a first stage turbine blade in
accordance with still another exemplary embodiment of the invention; and
FIGURE 18 shows a variation of Figure 15.
BEST MODE FOR CARRYING OUT THE INVENTION
Figure 1 is a schematic diagram for a simple-cycle, single-shaft heavy
duty gas turbine 10. The gas turbine may be considered as comprising a
multi-stage axial flow compressor 12 having a rotor shaft 14. Air entering the
inlet of the compressor at 16 is compressed by the axial flow compressor 12,
and then is discharged to a combustor 18 where fuel such as natural gas is
burned to provide high energy combustion gases which drive a turbine 20. In
the turbine 20, the energy of the hot gases is converted into work, some of
which is used to drive compressor 12 through shaft 14, with the remainder
being available for useful work to drive a load such as a generator 22 by
means of rotor shaft 24 (an extension of the shaft 14) for producing electricity.
A typical simple-cycle gas turbine will convert 30 to 35% of the fuel input into
shaft output. All but one to two percent of the remainder is in the form of
exhaust heat which exits turbine 20 at 26.
Figure 2 represents the combined cycle in its simplest form in which
the energy in the exhausc gases exiting turbine 20 at 26 is converted into
additional useful work. The exhaust gases enter a heat recovery steam

generator (HRSG) 28 in winch water is converted to steam in the manner of a
boiler. The steam thus produced drives a steam turbine 30 in which additional
work is extracted to drive through shaft 32 an additional load such as a second
generator 34 which, in turn, produces additional electric power. In some
configurations, turbines 20 and 30 drive a common generator. Combined
cycles producing only electrical power are in the 50% to 60% thermal
efficiency range using the more advanced gas turbines.
blades
In the present invention, steam used to cool the gas turbine b-uc-kets in
the first and second stages may be extracted from a combined cycle system in
the manner described in commonly owned application S.N. 08/161,070 filed
December 3, 1993. This invention does not relate to the combined cycle per
se, but rather, to the configuration of internal steam cooling passages in the
blades
first and second stage gas turbine •feuek-ess, consistent with the discussions
above.
Figure 3 illustrates in greater detail the area of the gas turbine which is
the focus of this invention. Air from the compressor 12' is discharged to the
several combustors located circumferentially about the gas turbine rotor 14' in
the usual fashion, one such combustor shown at 36. Following combustion,
the resultant gases are used to drive the gas turbine 20' which includes in the
instant example, four successive stages, represented by four wheels 38, 40, 42
and 44 mounted on the gas turbine rotor for rotation therewith, and each
including-bycl«&e63r blades represented respectively, by numerals 46, 48, 50
and 52 which are arranged alternately between Fixed stators represented by
vanes 54, 56. 58 and 60. This invention relates specifically to steam cooling
K1 fi rj f-\ <-i
of the first and second stage buetc-ets-r represented by blades 46, 48, and the
minimization of secondary Coriolis and centrifugal buoyancy forces or effects
in the internal blade cooling passages.

Referring to Figures 4 and 4A, a typical passage 2 is showr .n a blade
4 having a leading (or suction) side 6 and a trailing (or pressure) side 8. The
Coriolis induced secondary flow (assume rotation in the direction of arrow A)
transports cooler, higher momentum fluid from the core to the trailing side 8,
; whereby the radial velocity, the temperature gradient and hence the '.-onvective
effects are enhanced. Centrifugal buoyancy increases the radial velocity of the
coolant near the trailing side 8, further enhancing the convective effect. For
the leading side 6, the situation is just the reverse. Due to the Coriolis
induced secondary flow, the fluid exchanges heat with the trailing side 8 and
side walls before reaching the leading side 6. The fluid adjacent to the
leading side 6 is warmer and the temperature gradient in the fluid is lower,
weakening the convection effect. For the same reason, the Coriolis induced
flow leads to a lower radial velocity adjacent to the leading side 6, weakening
the convection effect further. Buoyancy effects become stronger at high
; density ratios such that flow reversal can occur adjacent to the leading side 6
of the passage 2. One of the objectives of this invention is to account for the
presence of these secondary flows in order to mitigate the adverse effects by
blades
appropriate design of the internal cooling passages in the buefcets, and
particularly the radial outflow passages where the secondary flow effects are
i more severe.
Referring now to Figure 5, the external appearance of the gas turbine
blade
first stage -bueket-46 in accordance with this invention is shown. The external
appearance of the blade or bucket 46 is typical compared to other gas turbine
'• blades, in that it consists of an airfoil 62 attached to a platform 64 which seals
blade
the shank 66 of the-bucket from the hot gases in the flow path via a radial seal
pin 68. The shank 66 is covered by two integral plates or skirts 70 (forward
and aft) to seal the shank section from the wheelspace cavities via axial seal
pins (not shown). The shank is attached to the rotor disks by a dovetail
i attachment 72. Angel wing seals 74, 76 provide sealing of the wheelspace

cavities. A novel feature of the invention is the dovetail appurtenance 78
under the bottom shank of the dovetail which supplies and removes cooling
blades
steam from the te«eket~via axially arranged passages 80, 82 shown in phantom,
which communicate with axially oriented rotor passages (not shown).
Figure 6 illustrates in simplified form, the internal cooling passages in
the first stage bucket 46. Steam entering the bucket via passage 80 flows
through a single, closed serpentine circuit having a total of eight radially
extending passages 84, 86, 88, 90, 92, 94, 96 and 98 connected alternatively
by 180° return U-bends. Flow continues through the shank via the radial
inflow passage 98 which communicates with the axially arranged exit conduit
82. Outflow passage 84 communicates with inlet passage 80 via passage 100,
while inflow passage 98 communicates with exit passage 82 via radial passage
102. The total number of radial passages may vary in accordance with the
specific design criteria.
blades
Figure 7 is a schematic planar representation of the Tbt^ek-et-shown in
Figure 4, and illustrates the incorporation of integral, raised ribs 104 generally
arranged at 45° angles in the radial inflow and outflow passages, after the first
radial outflow passage, which serve as turbulence enhancers. These ribs also
appear at different angles in the 180° U-bends connecting the various inflow
and outflow passages. Referring to Figures 8A-8C, it can be seen that
turbulator ribs 104 are provided along both the leading (or low pressure) side
and the trailing (or pressure) side of the blade erbucket 46.
Pins 106 (Figs. 6, 7) provided in the radial outflow passage 84 adjacent
the trailing edge improve both mechanical strength and heat transfer
characteristics. These pins may have different cross-sectional shapes as
evident from a comparison of Figures 6 and 7.

Figure 8A represents a transverse section through the root of the blade
4-6 and [he flow arrows indicate radiaJ inflow and outflow in the various
passages 84, 86, 88, 90. 92. 94. % and 98. Note again that the cooling steam
blade
flows into che-fewe-ket initially via passage 84 adjacent the trailing edge 108 and
exits via passage 98 adjacent the leading edge 109. The radial outflow
passages 84, 88, 92 and 96 are made smaller than radial inflow passages 86,
90. 94 with the exception of the radial inflow passage 98 adjacent the leading
edge 109 for reasons explained below. As already noted, the adverse effect of
Coriolis and buoyancy forces are more benign in radial inflow passages, and
these passages are therefore kept relatively large.
The leading edge passage 98 requires a high heat transfer coefficient.
This is forced by reducing the flow area to raise the bulk flow velocity, which
in turn raises the heat transfer coefficient which is proportional to mass flow
divided by the perimeter raised to the 0.8 power. The smaller cross section of
passage 98 results in a smaller perimeter, thus raising the heat transfer
coefficient.
The generally smaller radial outflow passages 84, 88, 92 and 96
counteract the tendency for any radial secondary flow recirculation resulting
from Coriolis and centrifugal buoyancy forces acting on the fluid in radial
outflow. This adverse tendency is counteracted by making the bulk flow
velocity as large as possible in radial outflow within the confines of
producibility and pressure drop. The radial outflow passages 84, 88, 92 and
96 are thus designed such that buoyancy parameters lead to enhanced heat
transfer rate on the leading side of the outflow passages.
blade
Figure SB illustrates the same-bucket-46, but with the cross-section
taken at the pitchline of the blade, halfway between the hub or root and the
tip. Figure 8C shows the same blade at the radially outer tip. From these

views, the relative changes in passage geometry from rooc to tip may be
appreciated.
With a judicious selection of aspect ratios (the ratio of length dimension
"L" to the width dimension "W" as shown in Figure 8B) and cross-sectional
area ratios in the radial outflow passages, as explained below, it is possible to
achieve, for a given aspect ratio, a buoyancy factor (for steam) of < 1, and
even as low as 0.15 in the radial outflow passages 84. 88, 92 and 96 where
secondary flow effects are critical, In this way, the unwanted secondary flow
effects (buoyancy and Coriolis) can be minimized particularly in the radial
outflow passages, while at the same time maximizing local heat transfer, [n
this regard, it has been determined that it is desirable to achieve a heat transfer
enhancement factor (- actual heat transfer ^ high ^
heat transfer xn a smooth tube
possible. For example, when the radial outflow passages are shaped to have
an aspect ratio of about 3.3 to I, it has been determined that, with regard to
heat transfer enhancement and Buoyancy Number (Bo), an enhancement factor
of 2 is achievable with a corresponding B, of 0.15. Between B0's of 0.15 and
0.80, it has been discovered that the enhancement factor drops below 2. As a
result, radial outflow passages should be designed to have B0's of less than
0.15 or greater than 0.80 when the aspect ratio is about 3.3 to 1.
For purposes of the above analysis, the passages were also provided
with turbulators 104.
It is expected that a similar undesirable range of Buoyancy numbers
will be identified for other aspect ratios, but this has not yet been confirmed.
Ft will be appreciated that these aspect ratios will change somewhat
along the length of the blade, from hub to tip due to the changing curvature

and twist of the blade. At the same time, the cross-sectional area ratio
between the larger radial inflow passages (with the exception of the smaller
radial inflow passage along the leading edge) and the smaller radial outflow
passages at the picchline, on average, should be about 1 '/: to I.
Since secondary flow effects are typically more significant in first stage
■bueket&r it follows that aspect ratio effects are also more significant in the first
blades blades
stage bttefceEs- Thus, in the second stage buckets-, the aspect ratios may be on
the order of 1 to 1 or 2 to I, while the cross-sectional area ratios may remain
blades
substantially as for the first stage buckets Once having determined the
configuration of the radial outflow passages, the radial inflow passages can be
configured consistent with requirements relating to heat transfer coefficients
and pressure drop constraints.
Ft should be noted here that the turbulence enhancing ribs or turbulators
104 also tend to reduce the adverse effects of buoyancy and Coriolis forces as
the local turbulence breaks up secondary flow tendencies.
blade
Figures 9 and 10A-10C illustrate a second stage -b-udcet in views which
blade
generally correspond to the first stage btieket shown in Figures 6 and 8A-8C.
blade
The stage two bucket 110 has six cooling passages, as opposed to the eight
blade
passages in the first stage btieket, reflecting the reduced cooling requirements
in the second stage. Thus, radial outflow passages 112, 116 and 120 alternate
with radial inflow passages 114, 118 and 122 in a single, closed serpentine
circuit. The first radial outflow passage 112 is connected to axial supply
conduit 124 via passage 126 while the last radial inflow passage 122 is
connected to axial return conduit 128 via passage 130. Pins 132 appear in the
last radial inflow passage 122, and it will be appreciated from Figs. 10A-10C
that raised ribs 134 are provided as in the stage one buckets. The Buoyancy
Number, aspect ratio and cross-sectional area ratios are as stated above.

An alternative design variation is also illustrated in Figure 9.
Specifically, the steam coolant flow path is reversed, i.e., steam enters the
blade
-bucket 110 and flows radially outwardly in leading edge passage 112 and exits
the bucket via trailing edge passage 122. This arrangement may be
advantageous in some circumstances.
blade
In both first and second turbine stages, the backet tips are cooled by
providing raised ribs on the underside of the tip cap as shown in Figures 11-
blade
13. In Figure 11, for example, the tip cap 136 of a-btieket 138 is formed with
integral ribs 140 on the underside of the cap in a U-bend between radial
outflow passage 142 and radial inflow passage 144. Turning vanes 146 may
be located in outflow passage 142 to direct flow into the turnaround cavity
corner 148 which is a typical location of stagnant flow and insufficient cooling.
In Figure 12, integral ribs 240 of squared off configuration are provided on the
underside of the tip cap 236, in further combination with turning vanes 246
and 246' in both outflow and inflow passages 242, 244, respectively. In
Figure 13, raised rib turbulators or trip strips 149 are provided in the 180° U-
bend region and on the underside of the tip cap 336 in combination with
rounded ribs 340 on the underside of the tip cap. These features also increase
local turbulence but, at least with regard to the turning vanes 146 and
turbulators 149, may not provide any heat transfer enhancement.
In Figures I4A and 14B, it can be seen that bleed holes 150 may be
provided where the passageway divider rib 152 meets the blade walls 154, 156
and the tip cap 158. This feature tends to provide relief from high thermal
stresses by unconstraining the corner region from the rib. Additional benefits
may be gained by chamfering or radiusing the external corners of the blade at
160. This reduces the effective wall thickness and reduces the temperature
gradient across the wall of the airfoil around the periphery of the tip cap 158.

Turning to Figures 15-18. alternative design configurations for first
blades
stage turbine btreleets are shown which are intended to enhance heat transfer in
the generally triangularly shaped (in cross section) trailing edge cooling
passage. The flow adjacent the trailing edge is laminar due to the constriction
of the core flow between the boundary layers, ft should be noted that the
second stage bucket does not experience the same trailing edge phenomenon.
so long as the trailing edge wedge angle is below about 12°.
With specific reference now to Figure 15, parallel flow passages 162,
164 are provided near the trailing edge 166 of the blade 168, fed from the
same entry passage 170. One passage 164 is intended to enhance heat transfer
at the trailing edge through an arrangement of opposed baffles 172, 174. The
other branch or passage 162 is intended to enable a high through flow by
providing a bypass to minimize overall pressure drop. Both passages meet
near the blade tip to continue into the serpentine circuit, and specifically into a
radial inflow passage 176. In this embodiment, the trailing edge passage 164
with its arrangement of baffles 172, 174, forces turbulence through the trailing
edge region via vortices caused by U-return bends (similar to the return bends
at the blade tip) between adjacent baffles projecting alternately from opposite
sides of the passage 164. Passage 164 will have 10-20% of the total flow from
entry passage 170 because of the high flow resistance from the head losses in
all of the U-bends. In the exemplary embodiment, there are about 10 such U-
bends (eleven baffles 172, 174 are shown).
Tests indicate that enhancement factors of 1.5 to 2 are possible at the
U-bends at the blade tip. With ten baffles in the passage 164, an exit
hydraulic diameter prior to the U-bends of about 0.35 inches will result in a
smooth wall heat transfer coefficient of about 500 BTU/ft.2. The turbulence
enhancement will bring the effective heat transfer coefficient to about 1000
BTU/ft.2. In addition, the number of serpentine inflow and outflow passages

can be reduced in this embodiment to six, in order to keep
overall flow in excess of 30 pps. It is important to keep total
flow rate at about 30 pps or greater, in order to keep exit
temperatures below 1050 F, and to maximize leading edge heat
transfer.
The flow split along the trailing edge 166 of the blade
168, and the overall pressure drop, will be controlled by several
variables including (a) the relative size of the bypass radial
outflow passages; (b) the degree of overlap of the baffles 172,
174; (c) the number of baffles; (d) the angle of inclination
of the baffles, and particularly the radially innermost baffle;
and (d) inlet and/or exit constrictions in the trailing edge
flows.
A variation of the above trailing edge passage configuration
is illustrated in Figure 16 where two parallel bypass 178 and
180 extend parallel to the trailing edge passage 182. Here again,
the radial outflow passages 178, 180 and 182 split from a common
entry or supply passage ( not shown) similar to passage 170
in the Figure 15 embodiment. This arrangement increases the
percent of coolant bypassing the trailing edge passage 182.
Turning to Figure 17, a radial outflow passage arrangement
involves parallel passages 184, 186 along the trailing edge
188 of the blade 190. Flow from radial outflow passage 186 splits
at the blade tip, with some of the flow moving into the narrow
diameter inflow trailing edge passage 184, and some of the flow
moving into an interior radial inflow passage 192 in the closed
serpentine circuit. The edge passage 184 exits into a passage
194 leaving the blade.

Figure 18 illustrates a variation of Figure 15 where vanes 196 are
utilized in the trailing edge passage 164' in place of baffles 172, 174 to
promote turbulence. Here again, the flow distribution is controlled by
variables discussed above in connection with Figure 15.
[t should also be noted that chevron turbulators 198 as illustrated in
Figures 15-18, may be preferred in particular circumstances over the 453
turbulators 104 in the earlier described embodiments, in light of higher heat
transfer enhancement with this type of turbulence promotor for the same
pressure drop. Some 45° angle turbulators may be retained, however, if
particular passages are too small to accommodate a chevron-shaped turbulator.
It will be appreciated that various configurations of 45° and chevron-shaped
turbulators may be included. It has also been determined that the first one
third of the passage length, as measured from the flow entry point, may be left
unturbulated in order to minimize pressure drop, tn addition, inlet entry
turbulence provides the necessary enhancement so that turbulators are not
required in this part of the passage length.
While the invention has been described in connection with what is
presently considered to be the most practical and preferred embodiment, it is to
be understood that the invention is not to be limited to the disclosed
embodiment, but on the contrary, is intended to cover various modifications
and equivalent arrangements included within the spirit and scope of the
appended claims.

WE CLAIM:
1. A closed circuit fluid cooled gas turbine blade having
a shank, a tip and an airfoil having leading and trailing edges
and pressure and suction sides, and an internal fluid cooling
circuit, the improvement comprising said internal fluid cooling
circuit having serpentine configuration having plural radial
outflow passages and plural radial inflow passages, said radial
outflow passages shaped to avoid an undesirable Buoyancy Number
associated with the aspect ratio selected.
2. The gas turbine blade as claimed in claim 1 wherein said
radial inflow passages have larger cross- sectional areas than
said radial outflow passages.
3. The gas turbine blade as claimed in claim 2 wherein a
radial inflow passage adjacent the leading edge of the bucket
has a smaller cross sectional area than said radial outflow
passages.
4. The gas turbine blade as claimed in claim 2 wherein a
ratio of the cross sectional area of the radial inflow passages
to the cross sectional area of the radial outflow passages is
1.5 to 1.

5. The gas turbine blade as claimed in claim 1 wherein
the aspect ratio is about 3.3 to 1 and the Buoyancy number is
<0.15 or >0.80

In a gas turbine blades having a shank portion, a radial tip portion and
an airfoil having leading and trailing edges and pressure and suction surfaces,
and an internal fluid cooling circuit, an improvement wherein the internal fluid
cooling circuit has a serpentine configuration including plural radial outflow
passages and plural radial inflow passages. The radial outflow passages, in
one example, are shaped to have aspect ratios of about 3.3 to 1 and Buoyancy
Numbers of < 0.15 or > 0.80. A method of determining a configuration for

steam cooling passages for a blades stage in a gas turbine is also provided
which includes, in one example, the steps of:
a) determining combustion gas inlet temperature and mass flow rate of
combustion gases passing through the gas turbine stage;
b) caking into account Coriolis and buoyancy secondary flow effects in

the steam coolant caused by rotation of the blades stage; and
c) configuring the radial outflow coolant passages to have a size and
shape sufficient to produce aspect ratios of about 3.3 to 1 and Buoyancy
Numbers in the radial outflow passages of < 0.15 or > 0.80.