Open Journal of Fluid Dynamics, 2013, 3, 28-35
http://dx.doi.org/10.4236/ojfd.2013.32A005 Published Online July 2013 (http://www.scirp.org/journal/ojfd)
Guide Vane with Current Plate to Improve Efficiency
of Cross Flow Turbine
Kiyoshi Kokubu1, Toshiaki Kanemoto2, Keisuke Yamasaki1
1Graduate School of Engineering, Kyushu Institute of Technology, Kitakyushu, Japan
2Faculty of Engineering, Kyushu Institute of Technology, Kitakyushu, Japan
Email: kiyoshi-kokubu@tanasui.co.jp
Received June 3, 2013; accepted June 10, 2013; accepted June 17, 2013
Copyright © 2013 Kiyoshi Kokubu et al. This is an open access article distributed under the Creative Commons Attribution License,
which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.
ABSTRACT
To get the sustainable society, the hydropower with not only the large but also the small/mini/micro capacities has been
paid attention to in the power generation. The cross flow turbine can work effectively at the comparatively low head
and/or low discharge, then the runner and the turbine profile has been optimizing. In this paper, the model turbine was
prepared in accordance with the traditional design, and the performance and the flow condition were investigated ex-
perimentally at the various operating conditions. The hydraulic efficiency is doubtlessly maximal while the guide vane
is at the normal/design position, and deteriorates in the lower discharges adjusted by the guide vane. Such deteriorations
are brought from the unacceptable flow conditions in the inlet nozzle. To improve the efficiency dramatically in the
lower discharge, the guide vane installed in the inlet nozzle was equipped with the current plate, and the fruitful effects
of the plate on the efficiency were confirmed experimentally.
Keywords: Hydraulic Turbine; Cross Flow Turbine; Guide Vane; Inlet Nozzle; Hydraulic Efficiency; Discharge
1. Introduction
It is desired to exploit the renewable energies, and the
hydro resources, such as small/mini/micro-scaled rivers,
agricultural/industrial channel/drain and so on, are ex-
pected to play a great role in the power generation [1]. It
is very important, for meeting such needs, to reduce the
initial cost of the power plant, since the plant does not
generate much electricity in comparison with the large-
scaled power plant [2]. The cross flow turbine is assem-
bled with few components, and can work effectively at
the comparatively low head and/or low discharge in the
onshore and offshore without nature disruptions. These
advantages have been demonstrated in many researches,
moreover, the turbine profile has been optimizing [2-4].
Nevertheless, the hydraulic efficiency of the cross flow
turbine is still lower than the efficiency of the turbines
such as the bulb/Francis/Pelton types used widely [5].
The supposable reasons for the lower efficiency are as
follows: 1) the flow crossing in the runner may contact
the main shaft at the partial load; 2) the flow direction
re-entering to the runner at the inner diameter does not
meet the blade angle; 3) the guide vane makes the flow
condition unacceptable for the runner blade in the low
discharge. To address these points, the installation of the
passage in the runner has been proposed to guide the
flow direction [6,7]. On the contrary, this paper proposes
to modify the guide vane profile in response to the dis-
charge.
2. Preparations to Improve Efficiency
2.1. Model Cross Flow Turbine in Tradition
The inlet nozzle and the runner profiles affect directly the
performance of the cross-flow turbine, and are prepared
for house works in accordance with the traditional design
[8] to understand the factors lowing the efficiency.
2.1.1. Inlet Nozzle
The inlet nozzle gives the angular momentum to the
runner, and it is desirable to give the same momentum in
the peripheral/tangential direction. Besides, the flow an-
gle at the nozzle outlet, measured from the circumferen-
tial direction, should be smaller in the same way as the
jet flow of Pelton turbine. Taking the material strength
and the fabrication into account, the angle was set at α1 =
17 degrees on the casing walls and the concave surface
of the guide vane, as shown in Figure 1. The opening
angle of the nozzle outlet is 109 degrees in accordance
C
opyright © 2013 SciRes. OJFD
K. KOKUBU ET AL. 29
Figure 1. Model cross flow turbine.
with reference [9]. The spiral type streamline may give
the same momentum in the circumferential direction, but
that makes the turbine size very large. Then, the most of
the walls are straight and the walls close to the outlet of
the nozzle are curved with α1 = 17 degrees.
The cross sectional areas along the concave and the
convex surfaces of the guide vane change gradually and
smoothly in the stream-wise, where the minimum width
is 40.4 mm at the convex side and 29.5 mm at the con-
cave side while the guide vane opening is at the nor-
mal/design position (GO = 100%) shown in Figure 1.
The guide vane opening is adjusted by the rotation
around the stem, and the passages can be closed (GO =
0) at the leading and the trailing edges.
2.1.2. Runner
The blade angle at the outer diameter of the runner was
designed as follows. The rotational velocity u1 at the
outer diameter of the runner is given by

12
11 1
π60 2uDn kvkgH (1)
where D1 is the outer diameter of the runner, n is the ro-
tational speed, k = 0.45 is the velocity coefficient and
brought from the coefficient of Pelton turbine [10], H is
the effective head and the absolute velocity ν1 is given by
(2gH)1/2. Then, the blade angle β1 given by Equation (2)
without the incidence angle, namely the relative flow
angle, is derived from the 111 11
sin( cos)tanvvu
1

(see Figure 2), Equation (1) and α1 = 17 degrees.

1
111
tansincos30degreesbaak



(2)
This angle is reasonable as well known [8].
Assuming the number of blades is infinite and the flow
is in the ideal condition without the gravitational effects,
At Section 1 At Section 3
At Section 2
Figure 2iangles.
me as the velocity at Section 1 (w1), but the flow direc-
lled in the runner may contribute
to
2.2. Model Setup and Experiments
e 1 is composed
. Velocity tr
sa
tion turns to 180 degrees. Then, the flow conditions with
α1 = 30˚, β1 = 17˚ and k = 0.45 at Section 1 give the flow
angle α3 = 100 degrees at Section 3. The angle α3 is
accepted taking account of Reference [3], though α3 de-
viating from the radial direction makes the discharge loss
increase more or less.
The main shaft insta
increase the rotational torque and deteriorate the flow
condition at Section 2, while the flow crossing in the
runner contacts with the main shaft. The flow pattern
crossing in the runner is determined with the blade angle
at the inner diameter of the runner and the outlet position
of the inlet nozzle. Then, the blade angle β2 was set at 87
degrees as recommended by Reference [3], and the outlet
position was set as shown in Figure 1 so as the flow
crossing in the runner does not contact with the imaged
shaft of 0.18 times as small as the inner diameter of the
runner.
Model cross flow turbine shown in Figur
of the inlet nozzle with the guide vane, the runner, and
the casing with the air breath. The runner, which was
designed at N11= nD1/H1/2 = 41.4 m, min1, has the outer
diameter of D1 = 250 mm, the inner diameter of D2 = 167
mm, and 30 blades formed with the single arc camber of
3 mm thickness and the height of B = 16 mm. The runner
has not the front shroud, and the clearance between the
blade tip and the front casing is 1 mm. Besides, the run-
ner has not the main shaft and is the overhang type. The
model has not the draft tube, that is, Section 3 is exposed
the relative velocity component at Section 3 (w3) is the
Copyright © 2013 SciRes. OJFD
K. KOKUBU ET AL.
30
to the atmospheric air.
In the experiments, the turbine head H and the dis-
ch
re
ev
2.3. Turbine Performances
ormances, where Q11 is
s the guide vane
op
2.4. Flow Conditions
Section 1
own in Figure 5,
arge Q were given by the booster pumping system
stood at the upstream, where H was defined by the head
at the center of the main shaft and estimated with the
static and the dynamic pressures at the turbine inlet, and
Q was measured by the orifice. The discharge Q is ad-
justed by the guide vane. In this paper, the guide vane
opening (GO) is defined as the percentage of the adjust-
ing/rotating angle (GO) from GO = 100% while the vane
is fully open (at the original/design position shown in
Figure 1) to GO = 0% while the passage is shut off by
the guide vane [3]. The runner speed is adjusted by the
inverter system with the regenerative braking circuit, and
is measured accompanying with the rotational torque.
The hydraulic efficiency η and the shaft power P we
aluated without the power losses by the sealing and the
bearing. The flow around the runner, at the middle blade
height of Sections 1, 2 and 3 is steadily measured with
the 3 hole Pitot tube composed of the wiry tube (the outer
diameter of 1 mm). Figure 3 shows the velocity distribu-
tion in the width direction at θ = 2˚ of Section 1, where θ
is the central angle measured from the upper edge of the
inlet nozzle (Figure 1), and z/z0 is the dimensionless dis-
tance from the front casing. It may be proved that the
boundary layers on the end walls scarcely affect the run-
ner performance and the flow conditions.
Figure 4 shows the turbine perf
the unit discharge [=Q/(BD1H1/2)], P11 is the unit output
[=P/(BD1H3/2)], η is the hydraulic efficiency [=P/(ρgQH)],
and ηmax is the maximum hydraulic efficiency at N11 =
40.7 m, min1 with GO = 100% which is in proximity
close to 41.4 m, min1 at the design stage. The maximum
output and hydraulic efficiency are at about N11 = 40.7 m,
min1 for GO = 100%. The output and the efficiency de-
teriorate and N11 giving the maximum η/ηmax and P11 is
slower, with the decrease of the guide vane opening
(GO), namely the unit discharge Q11. Besides, the unit
discharge Q11 increases slightly as N11 is slow, and above
characteristics has been well known [8].
The hydraulic efficiency deteriorates a
ening GO decreases, because the profile of the inlet
nozzle is optimized at GO = 100% in the design stage.
The experiment in house shows that the deterioration of
the efficiency is noticeable at GO smaller than 60%.
2.4.1. Flow Condition at
The flow conditions at Section 1 are sh
where the flow angles α1, β1, and the velocity ν1 are in
Figure 2. The relative flow angle β1 is in close to the
0.0
0.3
0.6
0.9
1.2
)1/2
0.0 0.2 0.4 0.6 0.8 1.0
v1 / (2gH
z/z0
Absolute velocity

= 2 deg. at Section 1
N11= 40.7 m, min-1
GO = 100%
Figure 3. Velocity distribution in the inlet nozzle w idth at
Section 1.
0.2
0.4
0.6
0.8
1.0
1.2
1
.
4
GO=100%
GO=80%
GO=60%
GO=40%
GO=20%
Efficiency
0.0
0.5
1.0
1.5
10 20 30 40 50 60 70
N11 m,min-1
Discharge
0
2
4
6
Output
Figure 4. Turbine performances.
ock less condition without the incidence angle for the sh
blade at the outer diameter of the runner, namely the zero
incidence angle at GO = 100% (Figure 5(b)), as ex-
pected at the design stage. At GO = 20%, the flow along
the concave/right surface of the guide vane runs out-
ward (0 degree < θ < 55 degrees in Figure 5(a)), and
runs inside accompanying with the flow along the con-
vex/left surface of the guide vane in the region larger
Copyright © 2013 SciRes. OJFD
K. KOKUBU ET AL. 31
(a)
θdeg.
(b)
θdeg.
(c)
Figure 5. Flow conditions at (a) Absolute flow
an θ = 55 degrees in Figure 5(a). That is, the runner
pposable that such flow conditions make the
ru
2.4.2. Flow Condition at Section 2
ection 2, where the
Section 1:
angle; (b) Relative flow angle; (c) Absolute velocity.
th
works effectively in the lower casing wall (large θ) at the
small GO. The flow may contribute to break more or less
the runner rotation near θ = 70 degrees, because the flow
is very slow with the comparatively large β1 (Figure 5(b)
and (c)).
It is su
nner work, namely the hydraulic efficiency, deteriorate
in the small guide vane opening (low discharge). Thus, it
is expected to improve the flow condition through the
passage surrounded by the concave/right surface of the
guide vane in the small opening.
Figure 6 shows the flow angles at S
flow angles are also given in Figure 2 and the boundary
between the flow crossing in the runner and the ambient
air was measured with the flow visualization. Section 2
has the flow discharging from the runner blade in α2 > 0
degree (discharging region) and the flow re-attacking to
the runner blade in α2 < 0 degree (attacking region),
where θ giving α2 = 0 degree is the border to separate
those two regions. The discharging region moves to the
large central angle θ at the small guide vane opening, but
the flow angle against θ is scarcely affected by the guide
vane opening, because the discharging flow direction
depends on the blade angle.
-60
-40
-20
0
20
40
60
6090120 150 180 210 240
GO=100%
N
11
=40.7 m,min
-1
GO=20%
N
11
=23.8 m,min
-1

deg.
Section 2
Absolute flow angle
Visualized boundary
(a)
-100
-50
0
50
100
150
200
6090120 150 180 210 240
GO=100%
N
11
=40.7 m,min
-1
GO=20%
N
11
=23.8 m,min
-1

deg.
Section 2
Relative flow angle
Inlet angle of
runner blade
Outlet angle of
runner blade
Visualized
boundary
(b)
Figure 6. Flow angles at Sec 2: (a) Absolute flow angle; tion
(b) Relative flow angle.
Copyright © 2013 SciRes. OJFD
K. KOKUBU ET AL.
32
The relative flow angle changes in the tangential di-
re
2.4.3. Flow Condition at Section 3
angle at Section 3,
2.4.4. Analysis of Output
lines obtained experimentally
ction and does not meet the blade angle, as shown
Figure 6(b). To rectify the flow direction and reduce the
shock loss, the flow passage was installed in the runner
[6,7]. The authors, however, cannot improve the hydrau-
lic efficiency by such passages, in house experiments
[11].
Figure 7 shows the absolute flow
which has the flow discharging not to the inside but to
the outside of the runner (Section 1-3) and the flow after
crossing in the runner inside (Section 1-3), where α3 is
given in Figure 2. Both flows are separated at θ = 187
degrees for GO = 100%, and θ = 203 degrees for GO =
20%. The discharging flow angle is larger than α3 = 90
degrees and may make the discharging loss increase at
GO = 20%, though the optimum angle is α3 = 90 de-
grees.
Figure 8 shows the stream
from the absolute flow, where these are at the N11 giving
the maximum hydraulic efficiency, respectively, S2 - S9
were estimated with the flow quantitatively measured at
Sections, S1 and S10 were estimated from the flow visu-
alization. Figure 9 shows the theoretical power fraction
estimated from the angular momentum measured
-100
-50
0
50
100
120 140 160 180 200 220 240 260
GO=100%, N
11
=40.7 m,min
-1
GO=20%, N
11
=23.8 m,min
-1

deg.
Section 3Absolute flow angle
Section 1-2-3
Visualized
boundary
Figure 7. Absolute flow angle at Section 3.
S
1
S
2
S
3
S
4
S
5
S
6
S
7
S
8
S
9
S
10
(a) (b)
Figure 8. S11 , min1);
tion is ε = the power
ne
treamlines: (a) GO = 100% (N = 40.7 m
(b) GO = 20% (N11 = 23.8 m, min1).
on each streamline, where the frac
on each streamline: PST)/the summation of PST: ΣPST). At
the small guide vane opening GO = 20% (Figures 8(b)
and 9(b)), the power fraction ε through Section 1-3
increases, that is, the fraction through Section 1-2-3,
especially 1-2, decrease obviously. On the contrary, the
flow through Section 1-2 contributes to increase the
power at the large guide vane opening (Figures 8(a) and
9(a)). These results and Figure 5 suggest that it is very
important to improve the flow condition in the smaller
central angle θ at Section 1 while the guide vane opening
is small.
3. Improving Efficiency in Low Discharge
3.1. Guide Vane Equipping with Current Plate
In response to the above suggestion, the guide va
equipped with the current plate of the single arc as shown
in Figure 10. Current Plate A is attached to the concave
surface of the guide vane at θ = 4 degrees of GO = 60%
and comes at the outer diameter of the runner with the
clearance of 1mm on θ = 31 degrees, where the plate
angle at the trailing edge coincides with the relative flow
angle β1 while operating at the maximum efficiency of
GO = 20%. In the same manner, Current Plate B is at-
tached to the guide vane at θ = 17 degrees and comes at
the runner outside on θ = 43 degrees. In the practical use,
the current plate may be made of the shape-memory alloy
plates.
Figure 9. Power fraction between streamlines: (a) GO =
100% (N11 = 40.7 m, min1); (b) GO = 20% (N11 = 23.8 m,
min1).
Current plate A
4°
Current plate B
17°31°
43°
GO=60
=20
Figure 10. Current plates.
Copyright © 2013 SciRes. OJFD
K. KOKUBU ET AL. 33
3.2. Improvemiency
the hy-
axi-
htly af-
fe
tional speed N. The distri-
bu
ditions
ct of the plate on the flow con-
erating at the highest effi-
ent of Hydraulic Effic
Figure 11 shows the effect of current plates on
draulic efficiency, where η and η are the m
Pmax Gmax
mum hydraulic efficiencies with the plate and without
the plate at each guide vane opening (GO). Current Plate
A is effective to improve the efficiency drastically in GO
smaller than 55%. Current Plate B also improves slightly
the efficiency in the region of 40% < GO < 60%, but the
plate is ineffective in total to improve efficiency. One
reason may be that the flow runs outward more or less
because the plate is at more downstream than Current
Plate A and the plate has the large turning angle. Besides,
it is not desired to improve the efficiency by the plate
attached to the surface at the far upstream of Current
Plate A, because of the extreme narrow passage.
Figure 11 also shows the unit discharge Q11opt giving
the efficiency ηPmax/ηGmax. The discharge is slig
cted by the current plates and is almost proportional to
the guide vane opening GO.
Figure 12 shows the hydraulic efficiency with Current
Plate A against the unit rota11
tion of the efficiency scarcely affected by the current
plate and the guide vane opening, but the hydraulic effi-
ciency is notably improved especially in smaller guide
vane opening.
3.3. Flow Con
Figure 13 shows the effe
dition at Section 1, while op
ciency of GO = 20%. The flow runs in the radial direc-
tion in the region of 15 degrees < θ < 60 degrees sur-
rounded by the back surface of the plate and the concave
surface of the guide vane, but the flow scarcely affects
0.5
0.6
0.7
0.8
0.9
1.0
1.1
1.2
1.3 1.8
Gmax
0. 2
0. 4
0. 6
0. 8
1. 0
1. 2
1. 4
1. 6
20 3040 5060 7080
GO %

Pmax
Q11opt m,m3/s


0.2
0.4
0.6
0.8
1.0
1.2
GO=60%
GO=40%
GO=20%
GO=6 0%
GO=4 0%
GO=2 0%
10 20 30 40 50 60
Without current plateWith current plate A
GO %
Figure 12. Hydraulic efficiency.
0
0.5
1
1.5
-50
-25
0
25
v without current plate
v with current plate A

without current plate

with current plate A
Se ction1
N
11
=23.8 m,min
-1
GO=20%
v
1
/(2gH)
1/2
1
Pmax Gmax
Q11opt m,m3/s
With current plate A
With current plate B
Without current plate
With current plate A
With current plate B
Without current plate
Figure 11. Efficiency improvement.
-200
-100
0
100
200
300
020 40 60 8010012
0

deg.
Section1 Relative flow angle
Inlet angle
of runner blade
GO=20%
With current plate
Without current plate
N
11
=23.8 m,min
-1
Figure 13. Effect of the current plate on flow conditions at
Section 1.
the runner work owing to the slow velocity. On the con-
trary, the flow affects mainly the runner work in the re-
gion not only from θ = 60 to 109 but also from 0 to 15
degrees. The flow in the latter region at Section 1 dis-
charges to Section 2 (see Figure 14), and improves con-
spicuously the runner work because the relative flow
angle is reasonable (Figure 13). Improvement of the
Copyright © 2013 SciRes. OJFD
K. KOKUBU ET AL.
34
flow condition at Section 1 makes the flow discharge at
the smaller central angle θ at Section 2, and makes the
flow crossing in the runner reinter to the runner with
nearly the same flow angle β2 as GO = 100% (see Figure
6(b)). The flow condition at Section 3 is also improved
as shown in Figure 15, and the results may bring the
decrease of the discharging loss.
3.4. Power Fraction between Streamlines
Figure 16 shows the effect of the current plate on the
theoretical power fraction ε defined at 2.4.4. The current
plate improves ε on streamline S1 to S4, by contribution
of the flow through Sections from 1 to 2. Figure 16 also
shows ηs = PST/ρgQH. The power is reasonably higher on
S7 to S10. The current plate deteriorates ε and ηs on S5 and
S6, because the flow passing convex surface of the guide
vane runs into the runner accompanying with de dead
water behind the current plate.
4. Conclusions
The relation between the performance and flow condition
-150
-100
-50
6090120 150 180 210 240
0
100
150
50
Without current plate
With current plate A

deg.
Section 2
Relative flow angle
Inler angle of
runner blade
Outler angle of
runner blade
GO=20%
Visualized boundary
N
11
=23.8 m,min
-1
Figure 14. Effect of the current plate on the relative flow
angle at Section 2.
-100
-75
120 140 160 180 200 220 240 260
-50
-25
0
Without current plate
With current plate A

deg.
Section 3Absolute flow angle
Visualized boundary
GO=20%
N
11
=23.8 m,min
-1
Figure 15. Effect of the current plate on the absolute angle
at Section 3.
0
2
4
6
8
10
12
14
16
18
20
22
12345678910
Sec.23
Sec.2
Sec.1
Without current plate
2Sec.13
3Sec.12
With current plate A
Sec.13
GO =2 0%
N=23.8 m,min-1
11
Stream line S
Figure 16. Effect on the current plate on the power fraction
between streamlines.
itions. Then, the hydraulic
proved in the low discharge adjusted by
the guide vane equipped with the current plate. These are
summarized as follows:
1) The flow condition in the inlet nozzle along the
concave surface of the guide vane is poor in the lower
discharge, and decreases the runner work.
2) The current plate installed at the concave surface of
the guide vane contributes to improve the efficiency in
the lower discharge.
ould like to deeply express their thanks to
Prof. Young-Do Choi from Mokpo National University,
Korea, for discussing the experimental results with the
numerical simulations [12], and Mr. Hirotaka Honda for
helping the experiments.
REFERENCES
[1] H. Kobayashi, “Small Capacity Hydropower,” Water
Culture, No. 39, 2011, p. 4. (in Japanese)
[2] L. Cho, K. Kurokawa, J. Matsui and H. Imamura, “Ap-
plication of Low Head Cross Flow Turbine to Micro Cap-
acity Hydropower (Simplified Structure and Improved
Performances),” JSME Series B, Vol. 67, No. 664, 2001,
pp. 222-227. (in Japanese)
[3] T. Toyokura, T. Kanemoto, T. Kitahora and A. Shiraisi,
“Researches of Cross Flow Turbine,” JSME Series B, Vol.
51, No. 461, 1985, pp. 143-151. (in Japanese)
was investigated experimentally with the model cross
flow turbine in accordance with the traditional design, at
e various operating condth
efficiency was im
5. Acknowledgements
The authors w
Copyright © 2013 SciRes. OJFD
K. KOKUBU ET AL.
Copyright © 2013 SciRes. OJFD
35
doi:10.1299/kikaib.51.143
[4] T. Kitahora, M. Inagaki, M. Uchida and S. Ooike, “In-
fluence of Pressure in Runner Chamber on Performance
Estimation of Micro-Head Cross Flow Hydraulic Turbine,”
Proceedings of Renewable Energy 2010, Yokohama, 31
May-2 June 2010, CD-ROM: O-Sh-4-5.
[5] Turbomachinery Society of Japan, “Hydro Turbine,” Ja-
pan Industrial Publishing Co. Ltd., Tokyo, 1992, pp. 65-
67. (in Japanese)
[6] S. Croquer, M. Asuaje, J. de Andrade, F. Jeanty
Clarembaux, “Use of CFD Tools in Internal Deflector
Design for Cross Flow Turbine Efficiency Improvement,”
Proceedings of ASME 2012 Fluids Engineering Summer
Meeting, Rio Grande, 10 July 2012, CD-ROM: FEDSM
2012-72017.
[7] H. Olgun, “Effect of Interior Guide in Cross Flow Tur-
bine Runner on Turbine Performance,” International Jour-
nal of Energy Researches, Vol. 24, No. 11, 2000, pp. 953-
964.
doi:10.1002/1099-114X(200009)24:11<953::AID-ER634
and J.
>3.0.CO;2-3
o. Ltd., Tokyo, 1992, p. 78.
92, p. 64.
e Korean Society of Marine Engi-
[8] T. Toyokura and T. Kanemoto, “Cross Flow Turbine,”
Energy and Resource, Vol. 4, No. 3 1983, pp. 234-239.
(in Japanese)
[9] Turbomachinery Society of Japan, “Hydro Turbine,” Ja-
pan Industrial Publishing C
(in Japanese)
[10] Turbomachinery Society of Japan, “Hydro Turbine,”
Japan Industrial Publishing Co. Ltd., Tokyo, 19
(in Japanese)
[11] K. Kokubu, K. Yamasaki, H. Hond and T. Kanemoto,
“Effect of Inner Guide on Performances of Cross Flow
Turbine,” Proceedings of 26th IAHR Symposium on
Hydraulic Machinery and System, Beijing, 19-23 August
2012, CD-ROM: IAHRXXVI-198.
[12] K. Kokubu, T. Kanemoto, S. Son and Y. Choi, “Perform-
ance Improvement of a Micro Eco Cross-Flow Hydro
Turbine,” Journal of th
neering, Vol. 36, No. 7, 2012, pp. 902-909.
doi:10.5916/jkosme.2012.36.7.902