In current and advanced gas turbine engines, increased speeds, pressures and temperatures are used to reduce specific fuel consumption and increase thrust/weight ratios. Hence, the turbine airfoils are subjected to increased heat loads escalating the cooling requirements to satisfy life goals. The efficient use of cooling air requires that the details of local geometry and flow conditions be adequately modeled to predict local heat loads and the corresponding heat transfer coefficients. The objective of this program is to develop a heat transfer and pressure drop data base, computational fluid dynamic techniques and correlations for multi-pass rotating coolant passages with and without flow turbulators. The experimental effort is focused on the simulation of configurations and conditions expected in the blades of advanced aircraft high pressure turbines. With the use of this data base, the effects of Coriolis and buoyancy forces on the coolant side flow can be included in the design of turbine blades

Hajek, T. J.

Johnson, B. V.

Wagner, J.

English

NASA Technical Reports Server

N89- 1289@
COOLANT PASSAGE HEAT TRANSFER WITH ROTATION*
T. J. Hajek
United Technologies Corporation
Pratt and Whitney
J. Wagner and B. V. Johnson
United Technologies Research Center
In current and advanced gas turbine engines, increased speeds, pressures and
temperatures are used to reduce specific fuel consumption and increase thrust/weight
ratios. Hence, the turbine airfoils are subjected to increased heat loads escalating
the cooling requirements to satisfy life goals. The efficient use of cooling air
requires that the details of local geometry and flow conditions be adequately mod-
eled to predict local heat loads and the corresponding heat transfer coefficients.
Improved turbine airfoil local temperature and hence, life predictions can be
realized by accurately accounting for the effects of rotation on internal cooling.
Although the effects of rotation which give rise to Coriolis and buoyancy forces can
be large, they are currently not adequately included in the heat transfer designs
of blades. Experimental data is particularly needed for the higher Rayleigh and
Reynolds number conaitions that are characteristic of turbine airfoils cooling
passages. This data is crucial for development of design correlations and computer
codes as well as their verification. Accurate prediction of local heat transfer co-
efficients enables the designer to optimize cooling configurations to minimize both
metal temperature levels and thermal gradients. Consequently, blade life and engine
efficiency can be significantly improved.
OBJECTIVE
The objective of this 36-month experimental and analytical program is to develop
a heat transfer and pressure drop data base, computational fluid dynamic techniques
and correlations for multi-pass rotating coolant passages with and without flow
turbulators. The experimental effort is focused on the simulation of configurations
and conditions expected in the blades of advanced aircraft high pressure turbines.
With the use of this data base, the effects of Coriolis and buoyancy forces on the
coolant side flow can be included in the design of turbine blades.
EXPERIMENTAL MODEL
The coolant passage heat transfer model features a four-pass serpentine arrange-
ment designed to reflect the passages within a gas turbine blade. For the present
experiments, the model was fitted with skewed turbulators, as indicated in figure
I. Figure 2 shows a schematic diagram of the model with the instrumentation and wall
sections indicated. Heat transfer coefficients are obtained for each wall section
element. These wall elements, numbered l to 64, consist of a copper block backed
with a thin film electrical resistance type heater and instrumented with two thermo-
couples. The copper wall sections are I0.7 mm x 49.3 mm (0.42 in. x 1.94 in.). Each
section is thermally isolated from the adjoining section by a 1.5 mm (0.060 in.)
thick divider strip of low thermal conductivity G-ll composite material. The
straight radial passages have a square cross section, 12.7 mm x 12.7 mm (0.5 in. x
0.5 in.).
*NASA Contract NASA-23691
193
https://ntrs.nasa.gov/search.jsp?R=19890003528 2020-03-20T04:36:16+00:00Z
DATA REDUCTION
Data acquisition/analysis consists of three general categories: equipment cali-
bration, model heat loss measurement, and heat transfer coefficient calculations.
The equipment calibration follows standard experimental procedures. Model heat loss
measurements precede each test. These measurements are executed with no coolant flow
and uniform wall temperature steady-state conditions, identical to the subsequent
test less the coolant flow. Heat transfer coefficients are then calculated for each
wall section element by applying the following procedure.
For each copper element the net energy convected to the fluid is calculated by
subtracting the electrical line losses and conducted heat losses from the total
energy supplied. Bulk fluid temperatures are then calculated based on an energy
balance for each flowpath section as follows:
q
T b = net_ 4 walls + Tb
out mCp in
where the model inlet bulk temperature is measured. Once bulk fluid temperatures are
determined, heat transfer coefficients are calculated from the equation:
q
net, wall
n =
A (T w - Tb)
where Tb is the average of the inlet and exit bulk temperatures. Thus for each
test case, 64 heat transfer coefficients are calculated.
Table I shows the test conditions for which data were acquired with the skewed
rough wall model. A total of 30 tests has been conducted to isolate the effects on
heat transfer of rotation rate, flow rate, coolant-to-wall temperature variations,
radius length and passage angle.
RESULTS
All of the experiments listed in Table I were completed to date. Measurements
of both the passage heat transfer and the channel pressure drop were conducted. Due
to the large number of data points obtained, comprehensive discussion of all the
results is beyond the scope of this paper. Instead, the following paragraphs will
mainly focus on comparisons between the smooth and the skewed trip strip channel
data at specific operating conditions and physical locations within the model. This
will facilitate better understanding of the underlying physical principles. It
should also be noted that due to the complexity of the subject matter, many of the
explanations presented herein are hypotheses and will require further substantiation.
In order to fully understand the channel heat transfer behavior under the in-
fluence of rotation, it is imperative that key _ifferences between the smooth and
the augmented channel be examined in stationary frame first. Figure 3 depicts a
comparison between the smooth and augmented models for the baseline condition of
Re : 25,000. In the first leg of the model, the smooth channel heat transfer
exhibits classical thermal development behavior (decreasing Nusselt number with
194
distance), whereas the rough wall shows nearly constant augmentation (factor of 3)
throughout the first leg. This is consistent with other Pratt & Whitney data.
Also quite different is the turn heat transfer. For the smooth wall case,
classical heat transfer increase (factor of 2) is present. In the augmentea channel
turn, the Nusselt number decreases through the turns (note that it reaches values
below the smooth duct). This is followed by an immediate increase, just downstrea_
of the turn. Subsequently, the second leg heat transfer decreases along the passage.
A similar pattern repeats for the second turn and the third leg. It is important to
note that the second and the third passages exhibit progressively decreasing average
Nusselt number. In fact, the third passage augmentation approaches levels expected
from normal trip strips. Even though good understanding of this phenomenon has not
been gained to date, it is clear that the first passage (with its well behaved
inlet) acts very much like a straight duct, whereas the subsequent passages are
strongly affected by the turn generatea secondary flows. It is felt that this fact
will play an important role in understanding the rotating results. Consistent with
other results, the rough channel side wall (rib wall) heat transfer is augmented
somewhat by the presence of the trip strips, but in general behaves similarly to
the smooth duct rib walls.
Figure 4 shows comparison of heat transfer results for smooth wall and skewed
trip rough wall at baseline rotating flow conditions of Ro = 0.24 and Re = 25,000.
As was the case with the smooth passage, the rough wall leading surface heat trans-
fer is significantly reduced (40-50% reduction) by the introduction of rotation.
Because the trailing side shows heat transfer augmentation of only 30%, the average
channel Nusselt number is reduced. While the rough model heat transfer in the first
passage is strongly affected by rotation, the subsequent legs and turns show very
little dependence. This fact is further supported by examining figure 5, where neat
transfer ratios are presented for several rotation rates. In this figure, the first
passage is again quite active, whereas the other channels exhibit very little varia-
tion with rotation. This finding further supports the hypothesis of turn generated
secondary flows dominating the subsequent passages.
The influence of buoyancy on rotating channel heat transfer is depicted in
figure 6. With the exception of the first leg, the leaaing (stablilized) wall shows
very weak dependence on temperature (density) variations. The trailing surface, on
the other hand, does show considerable dependence on _T. Better representation can
be found in figure 7. Leading and trailing surfaces are plotted for the last heating
element in the first leg. Note that the data at Ro = 0.0 do not coinciae. This is a
direct result of small manufacturing inconsistencies in the heating elements. The
density (temperature) ratio is shown with flag symbols. It can be clearly seen that
the trailing surface heat transfer is more sensitive to density variations than the
leading surface. When the trailing side data are examined closely, it can be seen
that the absolute heat transfer change due to temperature variations is approximate-
ly the same for both smooth and rough ducts. The rough wall leading surface, how-
ever, shows significantly smaller variation for a given _T change. This information
indicates that the buoyancy forces do not play as important a role in augmenting
heat transfer on "stabilized" surfaces for rough walls as they do for smooth walls.
This fact is further supported in figure 8, where the same data is plotted against
the Buoyancy Parameter. Both leading and trailing surfaces for smooth aucts as well
as trailing surface for rough ducts correlate well with the Buoyancy Parameter.
However, the leading surface data for the rough wall model do not collapse. Poten-
tially, this may indicate that in addition to the Duoyancy forces, some other
process, as yet unexplained, is becoming important.
195
It should also be noted in figure 8 that at high levels of the Buoyancy Para-
meter, the heat transfer ratios are approaching an asymptotic limit. In the case of
the smooth duct trailing surface, the limiting rough wall heat transfer level is
only 20% higher.
In reality, the difference is only 10% if the rough wall convection area is
corrected for the trip strip surface area. The important observation to be made here
is that at low values of the Buoyancy Parameter, the rough wall has significant heat
transfer advantage over the smooth wall. At high values of the buoyancy parameter,
however, the trip strip advantage is significantly diminished.
WORK PLANNED
Detailed analysis and correlation of the skewed turbulator data will continue.
Currently the model is being modified to include normal turbulators on the
leading and the trailing surfaces of the straight radial passages. A thirty point
test matrix, similar to the one in Table I will be executed.
196
TABLE 1
TEST MATRIX FOR ROTATING HEAT TRANSFER EXPERIMENTS FOR SKEWED TURBULATORS
Test
No.
UTRC
Run No.
201 6.7
202 9.9
203 8.8
204 10.8
Dimens iona I Parameters Basic
Dimension|ess Parameters
P _'_ _ AT H O_ Re Ro AI" H Ap _'_H
(pst) (rpm) (lb/sec) (F) (tn) (de9) Ttn d P V
147.7 0 0.013 81 25 0 25,337 0 0.15
147.8 0 0.006 80 25 0 12,490 0 0.15
149.5 0 0.025 80 25 0 50,715 0 0.15
145.0 0 0.024 80 25 0 75,348 0 0.15
205 15 80 25 0 25,000 0.006
206 145 80 25 0 25,000 0.06
207 275 80 25 0 25,000 0.12
208 412 80 25 0 25,000 0.18
209 550 80 25 0 25,000 0.24
210 825 80 25 0 25,000 0.35
211 145 80 25 0 12,500 0.12
212 550 80 25 0 50,000 0.12
213 825 80 25 0 75,000 0.12
214 275 160 25 0 25,000 0.12
215 550 160 25 0 50,000 0.12
216
217
218
219
220
221
222
223
224
145
412
412
412
55O
55O
550
825
825
825
550
275
275
55O
550
160 25 0 25,000 0.06
120 25 0 25,000 0.18
40 25 0 25,000 0.18
160 25 0 25,000 0.18
40 25 0 25,000 0.24
120 25 0 25,000 0.24
160 25 0 25,000 0.24
40 25 0 25,(_0 0.36
120 25 0 25,000 0.36
80 25 45 25,000 0.34
80 25 45 25,000 0.24
80 25 45 25,000 0.12
160 25 45 25,000 0.12
80 25 45 50,000 0.12
160 25 45 50,000 0.12
225
226
227
228
229
230
Secondary Dimension-
less Parameters
Grl Re2
0.00
0.01
0.05
0.12
0.22
0.45
0.09
0.06
0.05
0.14
0.14
0.03
0.17
0.07
0.20
O. 72
0.30
O. 36
0.28
0.64
0.42
9.22
0.05
0.14
0.06
0.14
Grx10-8
0.00
0.13
0.46
1.06
1.96
4.22
0.14
1.98
4.20
O. 87
3.49
0.21
1.55
0.63
O. 18
1.13
2.73
3.06
2.39
5.70
3.66
1.98
0.46
O. 87
1.98
3.49
Comments
No Rotation
Vary Ro
Hold AT, Re
Vary Re
Hold _T, Ro
Htah /_T
_An91e
IVarlatlon
- 450
Yaw /_T, Ro
at Reffi25,O00
Streamwise location of test sections identified b.y A to R.
All four test section surfaces for streamwise locations A through R are heated.
Leading test
section surfaces
Trailing test
section surfaces
I--i
Figure 1
I
I
In|et rxtt
Cross Sectional View of Coolant Passage Heat Transfer Model Assembly With
Skewed Trip Rough Walls
198
TEST SECTION ELEMENT IDENTIFICATION
SURFACES 1-32 ARE ON SIDE WALLS PERPENDICULAR TO VIEW SHOWN
SURFACES 33-4B ARE ON '_+ _ LEADING PLANE
SURFACES (49)-(64)ARE ON "'+ £t TRAILING PLANE
PRESSURE MEASUREMENT LOCATIONS [_ - [_
NOTE EACH TEST SECTION SURFACE IS INSTRUMENTED WITH TWO THERMOCOUPLES
r
/
i
[!]"
HEATED STRAIGHT
TEST SECTIONS
,,T}
GUARD
HEATER
5}-
MEASURED INLET
BULK TEMPERATURE
"-T,
5 39
2) {55)
5 40
I
i
I
) (56)
I 41
)) (57)
31 42
9 14
(E
10 13
((
11 12
(e
=
4
58) II (59
18
(64)
UNHEATED
SECTION
!
I
_] I_ MEASURED OUTLET
r,,, .,#.,,.,,, BULK TEMPERATURE
i_ SECOND --
TURN ]
Figure 2 Instrumentation Plan for Coolant Passage Heat Transfer Model
199
- 0.0 rpm m = 0.013 Ib/sec
P = 10 atm Re ==25,000
Open symbols - smooth wall data
Solid symbols - skewed rough wall data
symbol o • z_A o • 0 0
Surface Ist Leg Outside Inside Leading Trallinq
4.0
3.0
2.0
_" 1.0
"6
0.0
A B C D E F G H I J K L M N P R
_,_ Ou_,rd rtrst Inward Second OutwiP¢l Third
Stra(ght Turn Strat gt_t Turn Strot ght Turn
Streamwise Location
4.0
3.0
2.0
,,_
8
_- 1.0
"6
I I I I I l0.0
A B C D E F G H I J K L M N P R
Guard Outward First Inwilrd Second Outward lh|rd
StraSght Tur_ $trIH_t Tur_ Strltght Turn
Stream_tse Location
Figure 3 Comparison of Heat Transfer Results for Smooth Wall and Skewea Rough Wall
at Baseline Stationary Flow Conditions
200
_) = 550 rpm
P = 10 arm
- 0.013 Ib/sec
Re _ 25,000
Open symbols - smooth wall data
Solid symbols - skewed rough wall data
symbol o • I z_Ak o • 0 0
Surface Ist Leg Outside ilnside Leading Traillnq
S.O
4.0
3.0
_ 2.0
g
_ 1.0
0.0 m I m I
A B C D E F G H I a K L M N P R
5.0
4.0
"6 3.0
z
2.0
8
1.0
"6
0.0 I I I I m I
A B C D E F G H I J K L M N P R
Guard Outward First ImlP<l Second Outxel_d l_trd
Strltg_t Turn Strlt_t Turn Stretght Turn
Streamwlse Location
Figure 4 Comparison of Heat Transfer Results for Smooth Wall and Skewed Rough Wall
at Baseline Rotating Flow Conditions
201
AT - 80o F
Symbol
Rotation Number
Sl_ed (rpm)
a) Leading Surfaces 33-48
5.0
Re _ 25,000
O L$ 0 X
0 0.12 0.24 0.36
0 275 S50 825
4.0
Z
"_ 3.0
Z
b_
3
2.0t-
F-
"6
1.0
"1"
0.0 l J I
A B C D E F G
I I ! I
H I J K L W N P R
b) Trailing Surfaces 49-64
5.0
4.0
_: 3.0
Z
w,.
2.0
c
I--
1.0
"r
0.0
m
m
I I I I I I
A B C D [ F G H I J [ L H N P R
Stra_t Tu_ Stra_t Tv_ StriCt _m
Stream_(se Location
Figure 5 Effect of Rotation of Heat Transfer Results for Skewed Trip Rough Wall
Model
202
&T= 800 F
Symbol
Rotation Nunl)er
Spi)ed(rpm)
Re _ 25,000
0 ZS 0 X
0 0.12 0.24 0.36
4
0 275 550 825
c) Side Wall Surfaces 1-18
5.0 i4
z
3.0
Z
t.
2.0
e"
o
k--
1.0I
0.0 I I I I I I I
A g C D E F G g I J K L M g P R
d) Side Wall Surfaces 19-32
5.0
4.0
_-_ 3.0
Z
"_ 2.0
c
F,-
"_ 1.0
I
0.0 I
A D
I J I I
C D E F G _H I J K L M N P R
Ou_lrd Iri_t ifwaril Seccmd Out.rural 11titN
Strllght TII_ Strl4ght Tur_ Strl|ght Tl_n
Streanw|se Locatton
Figure 5 Effect of Rotation of Heat Transfer Results for Skewed Trip Rough Wall
(Cont'd) Model
203
C¢- 550 rpm
Symbol [3
4 p/p o.06
Temp. Diff. (OF) 40
c) Side Wall Surfaces 1-18
5.0
Re ==25,000
0
0.11 0.15
80 120
X
0.19
160
4.0
Z
3.0
Z
°
&,.
"_ 2.0
C
I--
"6
1.0
-r
0.0 I , I
A B C D E F G
d) Side Wall Surfaces ]9-32
5.0
I I I I
H I J K L M N P R
4.0
= 3.0
"_ 2.0
c
"_ 1.0
-r
0.0 I
A B
l I I I I J
C D E F 6 H I _ K L M N P R
O_e_l F|rst Imlrll Second Outmlrd 1flNird
Stflltl_t Tu_ $tra|pt Tvr_ $trltlht Tim
Streanwise Location
Figure 6 Effect of Density Ratio on Heat Transfer Results for Skewed Trip Rough
Wall Model
204
_= 550 rpm
Symbol ] rl
mp/p 0.06
Temp. Diff. (OF) 40
a) Leading Surfaces 33-48
5.0
Re _ 25,000
O
0.11 0.15
80 120
X
0.19
160
4.0
3.0
Z
g
_ 2.0
2
p-
_ 1.0
I
0.0 I I I j J I
A D C D [ F G H I J K L I'1 N P R
b) Trailing Surfaces 49-64
5.0
4.0
Z
3.0
z
ob.
= 2.0C
£
_ 1.0I
0.0 IA B
_m_
I I I I ' ]
C D E F G H 1 ,,1 K L M N P It
Oel_l_'d ft tilt Inelell $l¢:o,_1 Oel_l ell Iiih1 l,d
Strll_t Tir_ Strlll_t Term Stritght Turn
Strlamw4se Location
Figure 6 Effect of Density Ratio on Heat Transfer Results for Skewed Trip Rough
(Cont'd) Wall Model
205
43
0
0.5
OF POORQUAt,|TY
Leading Surface
/
/
I I I I
0.4 0.3 0.2 0.1
Trai Iing Surface 4
-o ¢/
+J1/ +
AT-4OOF
,, 25,000. ;q.SO°F O"
Open S?_ols - Smoth Wlll 4T.|2OOF
Solid S},g_w_l - Rovgh Wail 4T._iOOF
I ; i I
O.1 0.2 0.3 0.4 0.5
Rotation Number (Rd/V}
Figure 7 Effect of Rotation on Heat Transfer Ratio for Smooth Wall and Skewed
Rough Wall Models
5.0
g2.O
z
z
_1.0
w_
0.5
p-
0.2
+ ...... ,_+ - " "_-.. "-'..i"
--. , "'*,
_-4 _ _"<z, "'J "_ "......... /
I .
I I
J I
1.0 0.8
Leading Surface
0.I- l I I
1.2 0.6 0.4 0.2
Tral 11ng Surface
,, I'S.O00.
Opee S.Y_O|S - S_ot.h Wail
Solid SJ'mboll - Rough Wlll
I l I
0.2 0.4
Buoyancy Parameter, (_p/p)(nH/V)( rid/V}
$x_ol
_1|g$
_T'IOOF T
dT.BO°F
_T.I2OOF
t I
0.6 0.8 1.0 1.2
Figure 8 Effect of Buoyancy Parameter on Heat Transfer Ratio for Smooth Wall and
Skewed Rough Wall Models
206


Coolant passage heat transfer with rotation

Abstract

Similar works

Full text

Available Versions

NASA Technical Reports Server