Abstract-- The largely unexploited potential of small scale
energy hydropower remains crucial the development of new technologies to harvest the hydraulic energy on existing facilities. In this framework, several projects have been set up by the HES-SO Valais//Wallis and the EPFL Laboratory for Hydraulic Machines. One of the developed technologies is a new hydraulic micro- turbine, for recovering the energy lost in release valves of water supply networks. One stage of the micro-turbine consists of two axial counter-rotating runners. This paper deals with the hydraulic design process of the runners for a given site, including numerical flow simulations, fabrication and performance measurements of the micro-turbine. An overview of theoretical basics, simulation settings and assumptions, simulation results and test results is given. In the last part, the design optimization process is discussed.
Index Terms-- Hydraulic design, 5 axis machining, numerical flow simulation, performance measurements, optimization process
I. NOMENCLATURE A Area m C Absolute flow velocity m · s Cm Meridional absolute flow velocity component m · s Cu Peripheral absolute flow velocity component m · s d Local blade thickness m D Outer runner diameter m E Specific energy J · kg g Gravity m · s H Head bar I Momentum flow kg · m · s M Torque N · m nED Speed factor N Runner rotational speed min p Static pressure Pa P Hydraulic power W P Mechanical power W Q Discharge m · s QED Discharge factor r Radial position m t Maximal blade thickness m
This initial research work was supported by The Ark Foundation in the
framework of the HydroVS project. The project is now included in the SCCER Supply of Electricity and supported by the Swiss Commission for Technology and Innovation as part of the DuoTurbo project number 17197.1 PFEN IW.
* University of Applied Sciences HES-SO Valais//Wallis, Route du Rawyl 47, 1950 Sion, Switzerland. (Email: [email protected]).
# École Polytechnique Fédérale de Lausanne, Laboratory for Hydraulic Machines, avenue de cour 33bis, 1007 Lausanne, Switzerland
U Peripheral flow velocity m · s w Width of turbine stage m W Relative flow velocity m · s zS Number of stages Z Altitude m α Ratio between rotational speed of runners β Relative flow angle ° η Hydraulic efficiency θ Blade wrap angle ° ρ Water density kg · mω Angular speed rad · s
High pressure reference section Low pressure reference section First runner Second runner
II. INTRODUCTION YDROPOWER, small and large, remains the most important source of renewable energy for electrical
power production providing more than 15% of the world’s electricity mix. Small scale energy hydropower is distinct from traditional hydropower by generating less than 10 MW per site: the term mini-hydro is generally used below 2MW, micro-hydro below 500kW and pico-hydro below 10kW. In Switzerland, 56.6% of the electricity is provided by Hydropower, 5.7% coming from small hydro. Indeed, there are more than 1’300 small-scale hydropower plants in operation, with an installed capacity of approximately 860MW and an output of 3’770 GWh per year. In the post-Fukushima era, Switzerland has decided to renounce to its nuclear energy power stations and to accelerate the transition to a sustainable energy future based on carbon-free renewable electricity sources.
In this framework, a research project to develop new technologies to harvest hydraulic energy on existing facilities has been set up by the HES-SO Valais//Wallis and the EPFL Laboratory for Hydraulic Machines. One of the developed technologies is a new axial micro-turbine with counter-rotating runners for drinking water networks, which will cover a part of the hydraulic energy potential that must be exploited till 2050, see Fig.1. Indeed, the micro-turbine will recover the energy lost in release valves of water supply networks. One stage of this new micro-turbine consists of two axial counter-rotating runners, [1] & [2], whereby each runner drives an electrical generator, [3]. The compact axial design enables an in-line installation on existing facilities for low investment costs.
Design & performance of a hydraulic micro-turbine with counter-rotating runners
D. Biner*, V. Hasmatuchi*, F. Avellan#, C. Münch-Alligné*
H
2
In this paper, the process to develop the runners of the
micro-turbine for a given site is presented. The theoretical aspects of the hydraulic design are first introduced, as well as the design software to generate the runner geometry. Then the numerical flow simulations used to evaluate the performance of the hydraulic design of the micro-turbine are exposed. In the third part, the manufacturing process of the runners is described. Then the performance measurements of the runners on the hydraulic test rig of the HES-SO Valais//Wallis are introduced. In the last part, first insights of the optimization process based on the CFD results are discussed.
Fig. 1. Schematic representation of the counter-rotating micro-turbine
III. THE HYDRAULIC DESIGN
A. Technical specifications The counter-rotating micro-turbine belongs to the type of
reaction turbines such as Kaplan or Francis turbines. Considering the possibility of stacking several stages of counter-rotating pairs of runners in series, the micro-turbine can match quite wide range of pressure drop values ∆ in drinking water systems. For a given rated pressure drop, the number of stages of counter-rotating runner pairs can be selected to define the turbine stage specific energy: 1 · ∆p
(1)
The degree of freedom to adapt the operating point to the discharge fluctuations is the rotational speed of the runners, regulated by the electrical generators. Nominal operating conditions and most important requirements for the actual design are given in Table I. The pressure drop ∆p refers to one stage of the micro-turbine.
TABLE I NOMINAL OPERATING CONDITIONS AND REQUIREMENTS
Variable Symbol Value
Discharge Q 8.7 l · s Pressure drop ∆p 3 bar Runner rotational speed N 3’000 min Runner outer radius r 0.050 m Runner inner radius r 0.040 m Hydraulic power Ph 2.61 kW Mechanical power Pm 2.09 kW Hydraulic efficiency η ≥ 80 %
B. Design method The micro-turbine is a multi-stage axial machine with two
counter-rotating runners per stage placed in series. Considering the available hydraulic energy of a site, the maximal mechanical energy transferred by each runner can be determined. Assuming that the flow passing through the micro-turbine remains on a constant radius cylindrical surface, the Euler equation applied to a given streamline yields the relation between the hydraulic specific energy transferred to the runners and the balance of angular momentum which depends on the flow direction and velocity (Fig. 2). Moreover, assuming that the operating medium behaves like a perfect fluid, Euler equation can be considered as a one dimensional model of the fluid dynamics within the turbine to describe the runner geometry at the initial design phase.
Fig. 2. Model of the energy conversion in an axial turbine stage
C. Basic equations The transformed hydraulic energy inside a hydraulic
machine can be expressed using the specific hydraulic energy (2). Indeed, the specific energy results from the balance of
the static pressure, the kinetic energy and the potential energy of the operating medium between the turbine’s inlet and outlet. In this case, the potential and kinetic energies are the same at the inlet and the outlet. Consequently, the specific energy depends only on the difference of the static pressure between the inlet and outlet of the micro-turbine (3).
I I I I2 I I (2)
I I (3)
Inflow
Outflow First runner
Second runner
Consumption area
Altitude difference
Drinking water reservoir
Release valve
3
By taking into account the discharge , the hydraulic power can be expressed as: (4)
As mentioned in part B, the transferred mechanical energy is based on the conservation of the angular momentum of the flow creating a torque around the runner axis. The Euler equation (5) allows calculating the theoretically transferred specific energy using the peripheral flow velocity and the peripheral absolute flow velocity component at the inlet and the outlet sections of the runner. This theory is actually valid for a runner with an infinite number of turbine blades with an infinitely small thickness and for a totally inviscid fluid.
(5)
Finally, the efficiency of this energy conversion is calculated as: (6)
D. Velocity triangles The vector of the absolute flow velocity is the sum of the
peripheral flow velocity vector and the relative velocity vector (7). The geometrical representation of those vectors results in a typical velocity triangle, which is defined both at the inlet and at the outlet of each runner (see Fig. 3).
Fig. 3. Velocity triangles at the inlet and the outlet of the runners
At the inlet and the outlet of the micro-turbine the absolute flow velocity is parallel to the pipeline axis, the component is equal to zero. The meridional flow velocity component
is constant at all locations, due to the discharge conservation and the fact that the flow section between the hub and the shroud is constant. The peripheral flow velocity is given by the product between the radial position and the rotational speed of the runner. Using the Euler’s turbine equation, the component can thus be calculated. Consequently, all velocity triangles are defined and the relative flow angles can be determined. The latest provides actually the orientation of the blades.
(7)
E. Runner geometry To define the runner geometry, the blade angle is assumed
to be the relative flow angle . In reality, the flow direction slightly differs from the blade orientation, due to the limitation of the blade number and the profile effects. The definition of the blade geometry is done at an unwrapped cylindrical surface for a given radial position, supposing that there is no radial flow component. The skeleton-line is the basis of the blade profile and is determined by the width of the turbine stage w, the blade wrap angle θ, the radial position r and the relative flow angle at the leading and trailing edges, as shown in Fig. 4. The skeleton-line is defined by a polynomial of third order (8), ensuring thus a smooth flow deflection.
(8)
The coefficients … of the polynomial are determined using the following boundary conditions: 0 0 (9)
(10) 0 cotan (11) cotan (12)
Fig. 4. Definition of the skeleton-line
For a given stage width, the blade wrap angle is still a free variable. The blade wrap angle has to be optimized to get the smallest possible length of the skeleton-line: · 1 13 1 (13)
The final blade contour is given by the thickness distribution of a standard NACA 4–digit hydrofoil. The local
0/0
/
x: axial position [m]
y =
f(x)
: ci
rcum
fere
ntia
l pos
ition
[m]
4
thickness of the blade is given by a specific equation, depending on the type of the NACA profile, the maximal thickness , the blade width and the location on the skeleton-line (14). For the current case, the maximal thickness is fixed at 40% from the leading edge, with the value of the maximal thickness depending on the mechanical strength. , , (14)
As shown in Fig. 5, the upper and the lower profile contours can be finally described by a vertical /2 offset from the skeleton-line.
Fig. 5. Upper and lower profile contours.
F. The design software A Matlab Graphical User Interface (GUI) has been
implemented to design the runner geometry. Once the parameters defined, the software calculates the blade shape for each specified radial position between the hub and the shroud of the runner. Table II shows all the imposed design parameters and the resulting values for the actual design.
TABLE II SETUP PARAMETERS OF THE DESIGN SOFTWARE
Shroud radius 50 mm Hub radius 40 mm Runner separation 10 mm Blade clearance gap 0.2 mm Blade thickness 4.5 mm Minimal edge radius 0.5 mm Wrap angle type Fixed Profile type NACA-XXX4 Pressure drop 3 bar Discharge 8.7 l · s Design efficiency 85 %
Runner A Runner width 15 mm Rotational speed 3000 min Number of blades 5
Runner B Runner width 20 mm Rotational speed 3000 min Number of blades 7
The structure of the design software is given in Fig. 6.
Accordingly, the upper and lower profile contours are calculated and saved as 3D point rows for several radial positions. These data are then exported to the CAD (Computer-Aided Design) software using the API (Application Programming Interface). Finally, the 3D point
rows are interpolated to close the 3D spline contours and used to define the blade surface.
Fig. 6. Structure of the design software and the connection to the CAD software.
IV. FLUID SIMULATION
A. Numerical setup Numerical flow simulations are today an indispensable tool
for the development of turbine design and the evaluation of hydraulic performance. Development costs for expensive experimental explorations can be saved and very detailed analysis results can be obtained.
TABLE III NUMERICAL SCHEME
Simulation type Steady Spatial scheme 2nd order specified blend factor:1 Turbulence Model SST Residual Target RMSmax < 10-12
The performance of the designed runners has been analyzed
using 3D flow simulations of the full water passage of the micro-turbine. The numerical parameters of the simulation are summarized in Table III. The steady numerical simulations have been performed with the commercial software ANSYS CFX 15.0, based on the finite volume method.
/2 /
/
/ /2
Define Parameters
Skeleton-Line
Calculations
Profile Calculations
Performance Calculations
Draw 3D Display
Save Parameters
Turbine GUI
Additional Functions
Save Runner Geometry
CAD API3D splines
B. Computational domains and spatial discreThe computational domains are illustrated
domain consists of the inlet pipe with thehouses the electrical generator of the first rRotor 1 and 2 domains include respectivelyoutlet runners of the micro-turbine. In both Rhub and shroud are assumed to rotate with theis no gap between the tip of the blades and ththe Stator 2 domain consists of the outlet pihub region that houses this time the electricasecond runner.
Fig. 7. Computational domains.
The global mesh information is provided adapted unstructured mesh has been generateICEM commercial software for each domatetrahedral cells.
C. Boundary conditions The detailed boundary conditions
computational domain, for both the stationaryparts, are provided in Table V. At the inlet ofdifferent constant discharge values, correinvestigated operating conditions, are imposof the Stator 2 domain, 0 [Pa] relative averacondition is selected. The interfaces betwerotating domains (see Fig. 8) are treated wicondition, the connection being ensured by
Stator 1
Stato
RotorRotor 1
etization in Fig. 7. Stator 1 e static hub that runner. Then, the
y the inlet and the Rotor domains, the e blades and there
he shroud. Finally, ipe with the static al generator of the
in Table IV. The ed with the Ansys ain using mostly
ION
Elements
1’458’718 1’491’680 1’193’490 1’415’510
5’559’398
of the whole y and the rotating f the Stator 1, four esponding to the sed. At the outlet age static pressure en the static and ith a Frozen-rotor the General Grid
Interface (GGI) scheme. Finally, condition is used for all the solid stat
D. Numerical simulation results The following results are ba
simulation setups with different vIndeed, the values correspond res100% and 110% of the nominarotational speeds (3000 rpm), correspin the design, has been kept consshows the field of the relative velociwhole micro-turbine at the n(Q = 8.7 l∙s-1).
numerical results. As expected, th(BEP) is found at 8.7 · dischathe design parameters. One may statrecovers more mechanical power thathe main part of the static pressure drunner. A maximum efficiency of 8nominal discharge, the mechanical p
or 2
r 2
5
a smooth no-slip wall tic or rotating surfaces.
ITIONS
ndition
85 / 8.7 / 9.57 [l∙s-1] ge static pressure
lip wall
erfaces.
ased on four numerical values for the discharge. spectively to 80%, 90%, al discharge. The runner ponding to the one defined
stant for all cases. Fig. 9 ity streamlines through the ominal operating point
arge, which corresponds to te here that the first runner an the second one. Indeed, drop is created by the first 83.14% is reached for the
power being 2’753 W.
6
V. MACHINING
A. Machining process To finalize the development of the turbine geometry, the validation of the hydraulic performance must be done by model testing. In this purpose, the machining of the runner prototypes is mandatory. The designed runners have been manufactured in the mechanical workshop of the HES-SO Valais//Wallis. First, the basic axisymmetric body of the runners is made by a turning operation. The relatively complex blade geometry is then realized by a 5-axis milling operation [7], using the Deckel Maho DMU50 eVolution 5-axis machining center (see Fig. 10). The translational movements on the X, Y and Z axes are executed by the milling tool, whereby the rotational movements around the B and C axes are executed by the machine table. Since the actual configuration of the machine (adapted to mill molds for plastic or powder injection moldings) does not allow machining the whole runner in one single step. To cope with this, the turbine runners are machined blade by blade, using a dividing head. Anyway, if the method is reasonable for prototyping, a different machine type would be recommended for production of series.
Finally, the chosen material for the runners is brass, due to its good corrosion resistance to water and its excellent machinability. Moreover, its mechanical stress has been validated by FEM simulations.
Fig. 10. Schematic representation of the DMU 50 eVolution 5-axis machining center.
B. Tool path generation The tool paths for the 5-axis machining have been generated
with the AlphaCAM, a Computer Aided Manufacturing commercial software. The CAD model serves as a basis for
Fig. 11. Generated tool path for the roughing operation (left) and solid
simulation of the 5-axis milling (right).
the tool path generation. A wrapped pocketing operation is used for the roughing operation by an ø4mm end mill tool, as shown in Fig. 11. The finishing is realized by a surface parallel peripheral milling as well as a wrapped pocketing operation to complete the hub surface by an ø3mm spherical cutter.
C. Machining results The machining takes approximately four hours per runner
and a satisfying quality is obtained. The 5-axis milling process as well as the final result is illustrated in Fig. 12.
Fig. 12. 5-axis milling (left) and final result of the machining (right).
VI. PERFORMANCE MEASUREMENTS
A. The hydraulic test rig To validate the simulation process of the hydraulic
turbomachines and to complete their development, model tests are still essential. At the University of Applied Sciences HES-SO Valais/Wallis, a hydraulic test rig was installed to perform
Fig. 13. Hydraulic test rig of the HES-SO Valais//Wallis with the installed micro-turbine prototype, a) Main reservoir b) Centrifugal pumps c)
Pressurized reservoir d) Testing model
+B
+C
+Z
+Y
+X
a
b
cd
7
experimental tests of small scale turbines, pumps or other hydraulic components, see [4]. The test rig is built on two floors and supplied with fresh water from a main reservoir, see Fig.13. Three recirculating multistage centrifugal pumps connected in parallel supply the water circuit with hydraulic power. A pressurized reservoir allows simulating different implantation levels of the model. The actual prototype of the micro-turbine has been installed on the test rig and allows testing the characteristic of different runner geometries. The electrical generators placed inside the hubs allow the regulation of the rotational speed of each runner [3]. To measure the mechanical power, each runner axis is equipped with a torque sensor.
B. Testing method An advantage of experimental tests is the possibility to
obtain a large number of measurement points over the whole operating range of a hydraulic machine in relatively short time. To retrieve the characteristic curves of the micro-turbine by fluid simulation would require a substantial computing time: more than 16 hours per operating point if the whole water passage is considered.
To create the characteristic curves of the micro-turbine, the degree of freedom of the turbine regulation, α (15), describing the ratio between the absolute rotational speeds of the two counter-rotating runners, is introduced. The discharge, the head, the rotational speed and the torque of each runner are measured for each operating point.
(15)
Finally, the hydraulic performance of a large number of operating points has been measured at different constant operating heads. Indeed, different values of the rotational speed ratio (see Table VII) have been systematically considered over the whole possible range of the runner rotational speed.
TABLE VII
RUNNERS ROTATIONAL SPEEDS OF THE MEASURED OPERATING POINTS
To characterize hydraulic machines, dimensionless values
are often used to enable the comparison between different operating points or different model scales. For the micro-turbine, the discharge factor and the speed factor are
used to create the characteristic curves (16), (17). Those values refer to the external runner diameter and the rotational speed of the second runner.
(16)
(17)
C. Test results The main experimental results are given in Table VIII. For
every constant head measurements, the resulting BEP is given. Theoretically, the maximal hydraulic efficiency does not depend on the head, indeed there is only small difference of the efficiency between the different heads. Due to the mechanical losses of the runner bearings, discharge losses and turbulences in the blade clearance gap, the desired efficiency cannot be reached in the experimental tests with the actual configuration of the prototype.
TABLE VIII SUMMARY OF TEST RESULTS
H [bar] Q [l·s-1] α [-] NB [min-1] ηh [%]
0.5 3.95 1 1010 50 1 5.63 1 1493 51.5
1.3 6.77 1.18 1749 50.5 2 7.9 1 2003 52.8
2.5 9.22 1 2499 52.9 3 9.8 1.18 2257 53
Fig. 14. Hydraulic efficiency as a function of the speed factor and the discharge factor for a testing head of 1.3 [bar].
Fig. 15. Relative hydraulic efficiency as a function of the discharge and the head .
8
In Fig. 14 the characteristic for a constant head of 1.3 [bar] is provided. The black lines indicate the ratio between the rotational speeds of the runners. As predicted, the best efficiency point was found close to 1. The diagram is based on dimensionless values, normalized to the specific energy and implicitly to the testing head. It actually allows comparing characteristics between different heads.
Another way to present the characteristic of the turbine is to express the efficiency in relation to the discharge and the operating head. This representation is given in Fig. 15.
VII. DISCUSSION OF THE RESULTS The nominal design parameters along with the obtained
results of the numerical simulation and of the experimental tests at the BEP (in the investigated operating range) are given in Table IX.
TABLE IX SUMMARY OF DESIGN, CFD AND TEST RESULTS
Value Design CFD BEP Tests BEP
[%] 85 83 ~53 ∆ [bar] 3 3.8 3
[ls-1] 8.7 8.7 9.8
[-] 1 1 1.18
[min-1] 3000 3000 2257 / [-] 1 1.22 1.5
The efficiency found by flow simulation is satisfying the
requirements. The measured efficiency is much lower due to the mechanical losses and the flow effects in the blade clearance gap, which are not considered at the simulation level. The assumption that there’s no gap between the runners and the shroud does not represent the physical behavior on the test rig. To obtain more comparable conditions between the simulation and the experimental tests, the gap of about 0.2mm has to be eliminated using a fix attached external ring. Tests with the mentioned configuration are planned to be performed. Further the mechanical losses of the bearings are not precisely known, which does not allow making a statement about the real hydraulic efficiency.
VIII. FIRST INSIGHTS OF THE OPTIMIZATION PROCESS The optimization of a turbomachine is a complex procedure
whereby a large number of parameters are interacting. The optimization is generally based on results of numerical simulation as well as empirical formulae.
A. General optimization tendency A challenging problematic in the hydraulic design of this
type of axial turbines is the relatively low discharge in contrast with a relatively high desired operating head. The higher the operation head for a given discharge, the more inconvenient are the flow conditions within the runners. Moreover, for the relatively low mechanical powers, the bearings friction losses become more and more important and limit the maximal efficiency. In other words, the obtained hydraulic efficiency of the micro-turbine sets the limitation of the maximal mechanical power that can be transmitted by one pair of runners.
B. Flow stability at the runner interface The particular turbine configuration of two counter-rotating
runners can create an effect of instability on the flow direction at the interface between runners. A small deviation of the relative flow angle at the outlet of the first runner can affect negatively the flow direction at the inlet of the second runner. The absolute velocity vector at the outlet of the first runner is equal to the absolute velocity vector at the inlet of the second runner. The relation between the relative flow angle and can be formulated as:
(18)
The mentioned relationship is represented in Fig. 16. At the point of highest instability reaches a value of 90°. To guarantee a correct flow angle at the inlet of the second runner and to undesired flow separation, the operating point must be outside the shaded region.
Fig. 16. Dependence between the relative flow angles at the runner interface for 3.08 · and 3000 . The shaded zone is
characterized by flow instability.
C. 2D blade cascade simulation As already mentioned, the design process is based on a one
dimensional model: all blade profile effects are not taken into account. Due to this simplification, the real flow angle cannot be precisely predicted. For small relative flow angles at the runner outlet, the absolute velocity component becomes highly sensitive to small deviations of , and consequently the transmitted hydraulic power can significantly differ from the desired value. A fast method to identify blade profile effects is to use two dimensional blade cascade simulation. The two dimensional model is used to optimize the flow conditions on a defined radial position. Executing numerical analysis on several 2D domains, this method may be considered as a quasi-three-dimensional (Q3D) method [5]. The main advantage comes from the fact that this method allows
TABLE X MAIN PARAMETERS OF THE 2D FLOW SIMULATION
Mesh Type Non structured, triangular Number of Elements 46’129 Number of Nodes 25’225 Viscosity Model k-epsilon (2eqn) Computing time on standard PC ~70sec
Inle
t flo
w a
ngle
of t
he
seco
nd ru
nner
[°]
Outlet flow angle of the first runner °
executing a large number of simulations inobtain an iterative optimization process of the
The main parameters of the employed 2Dare given in Table X. An example of the rFig. 17. The evolution of the static pressure wis represented by a contour plot along with threlative flow velocity indicated by a vector global performance can be approximatelyinterpolating the simulation results of sevebetween the inner and the outer radiuses
Fig. 17. Pressure contour and vector field of relative velo2D blade cascade simulation.
Fig. 18. Quasi-three-dimensional analysis of the turbinfor iterative optimization of the runner geometry. Co
pressure are presented for different radial p
n a short time to e blade geometry.
D flow simulations result is shown in within the runners he direction of the field. Finally, the
y determined by eral 2D domains .
ocity as a result of the
ne performance, used ntours of the static
positions.
IX. CONCLUS
Runners of a multi-stage microtating runners for drinking wadesigned based on a simplified oneBy using 3D numerical flow simulathe designed runner geometry has behydraulic efficiency (> 0.8) has bRunner prototypes have been machcenter and finally tested on a hydturbine characteristics have beunconsidered mechanical losses, efficiency is found lower than the simthis, improvement of the pmeasurements of mechanical losses in a future step. To conclude, firstdesign optimization is presentedsimulations.
X. REFERENC
[1] C. Münch-Alligné, S. Richard, B. Mei“Numerical simulations of a counter rotaHydroinformatics, P. Gourbesville et al. 363-373, 2014
[2] V. Hasmatuchi, C. Münch, S. Gabathule“New Counter-Rotating Micro-Hydro Systems”, Hidroenergia 2014, Istanbul, T
[3] D. Melly, R. Horta, C. Münch, H. Binera PM-Generator for a Counter-RotatinInternational Conference on Electrical M
[4] V.Hasmatuchi, F. Botero, S. GabathuleControl of a New Test Rig for Small 2014, Cernobbio, Italy, 2014.
[5] G. Peng, S. Cao, M. Ishizuka, S. Hayamflow hydraulic turbine runner”, InternMethods in Fluids, pp. 517-531, June 200
[6] P. Drtina, M. Sallaberger, “Hydraulic tuof-the-art computational fluid dynamics Institution of Mechanical Engineers, PEngineering Science, vol. 213 no. 1 85-
[7] H. B. Kief, H. A. Roschiwal, CNC HandHanser Verlag, 2009.
. . .
. . .
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SION cro-turbine with counter-ater networks have been e dimensional flow model. ations, the performance of een analyzed. The required een successfully verified. hined on a 5-axis milling
draulic test rig, where the en measured. Due to the measured hydraulic
mulated one. To cope with rototype bearings and are going to be performed
t insights of the hydraulic d using blade cascade
CES ier, V. Hasmatuchi, F. Avellan, ating micro turbine”, Advances in (eds.), Springer Hydrogeology, p
er, S. Chevailler, and F. Avellan, Turbine for Drinking Water
Turkey, 2014. , S. Chevailler, “Development of ng Micro-hydro Turbine” XXI
Machines, Berlin, Germany, 2014. er and C. Münch, “Design and Power Turbomachines”, Hydro
ma, “Design optimization of axial national Journal for Numerical 02. urbines-basic principles and stat-applications, Proceedings of the
Part C: Journal of Mechanical 102, 1999 dbuch 2009/2010, München, Carl
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XI. BIOGRAPHIES D. Biner graduated in 2014 the Bachelor of Science in Systems Engineering, Design & Materials specialization, at the University of Applied Sciences Western Switzerland, HES-SO Valais//Wallis in Sion. Since September 2014, he is scientific assistant at a 50% basis in the hydraulic energy research team of Prof. Münch at the HES-SO Valais//Wallis, besides he’s going through the MSE master’s degree studies in industrial technologies at the HES-SO. He is working on experimental projects in hydraulic turbo machinery. His main research interests are the hydraulic design, the performance measurements, flow simulations and the optimisation of small scale hydro machinery. Vlad Hasmatuchi graduated in 2007 at the Faculty of Mechanical Engineering, Hydraulic Machinery Branch from “Politehnica” University of Timisoara, Romania. In the same year, Vlad Hasmatuchi joined the Laboratory for Hydraulic Machines from the École Polytechnique Fédérale de Lausanne (EPFL), Switzerland, to achieve a doctoral work in the field of hydraulic turbomachinery. In 2012 he got his Doctoral Degree in Engineering from the EPFL. Since 2012 he is Senior Research Assistant in the hydraulic energy research team at the HES-SO Valais//Wallis, School of Engineering in Sion, Switzerland. He is in charge mainly of experimental investigations, as well as of numerical simulations. His main research interests are the hydrodynamics of turbines, pumps and pump-turbines, including design and evaluation of hydraulic performance. Prof. François Avellan graduated in Hydraulic Engineering at the INPG Ecole Nationale Supérieure d'Hydraulique, Grenoble France in 1977, and, in 1980, got his Doctoral Degree in Engineering from the University of Aix-Marseille II, France, at IMST, the Institut de mécanique statistique de la turbulence, CNRS Associate Laboratory. In 1980, he is joining the EPFL Laboratory of Fluid Mechanics as Research Associate and, in 1984; he is appointed Senior Research Associate at the newly created EPFL Institute of Hydraulic Machines and Fluid Mechanics for leading the Research Group in Cavitation. Since 1994, he is Director of the EPFL Laboratory for Hydraulic Machines and he was appointed Ordinary Professor in 2003. His main research interests are the hydrodynamics of turbines, pumps and pump-turbines, including cavitation, hydro-acoustics, design and evaluation of the performance of hydraulic machines trough both experimental investigations and numerical simulations. From 2002 to 2012, Prof. F. Avellan was the Chairman of the IAHR Section on Hydraulic Machinery and Systems. Honorary Doctorate of the Polytechnic University of Bucharest, Romania, in October 2003, Prof. François Avellan has been awarded "Grand Prix d'Hydrotechnique 2010" by the Société Hydrotechnique de France. Cécile Münch-Alligné obtained an engineering degree from INPG, École Nationale Supérieure d'Hydraulique, Grenoble France ENSHMG, department of Numerical and Modelling of Fluids and Solids in 2002. Then, she got a grant from the CNRS and the CNES to start a Ph.D. thesis on large eddy simulations of compressible turbulent flows. She defended her doctoral degree in 2005 at the INPG. From 2006 to 2010, she worked as a research associate in the Laboratory of Hydraulics Machines at EPFL on flow numerical simulations in hydraulic turbines. Since 2010, she is professor at the HES-SO Valais//Wallis, School of Engineering in Sion, Switzerland. She is head of a new hydraulic research team specialized in small hydro applications. Her main research interests are small hydro, hydraulic turbomachinery, numerical simulations, performance measurements, turbulence and fluid-structure interactions.
