1
Design & performance of a hydraulic micro-turbine with counter-rotating runners D. Biner*, V. Hasmatuchi*, F. Avellan#, C. Münch-Alligné*
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 HESSO 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 C Cm Cu d D E g H I M nED N p P P Q Q ED r t
Area Absolute flow velocity Meridional absolute flow velocity component Peripheral absolute flow velocity component Local blade thickness Outer runner diameter Specific energy Gravity Head Momentum flow Torque Speed factor Runner rotational speed Static pressure Hydraulic power Mechanical power Discharge Discharge factor Radial position Maximal blade thickness
m m·s m·s m·s m m J · kg m·s bar kg · m · s N·m min Pa W W m ·s m 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 978-1-4673-7172-8/15/$31.00 ©2015 IEEE
U w W zS Z α β η θ ρ ω
Peripheral flow velocity Width of turbine stage Relative flow velocity Number of stages Altitude Ratio between rotational speed of runners Relative flow angle Hydraulic efficiency Blade wrap angle Water density Angular speed
m·s m m·s m ° ° kg · m rad · s
High pressure reference section Low pressure reference section First runner Second runner
II. INTRODUCTION
H
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 counterrotating 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 counterrotating 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.
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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.
Outflow
First runner
Second runner
Inflow
Drinking water reservoir
TABLE I NOMINAL OPERATING CONDITIONS AND REQUIREMENTS Variable
Symbol
Value 8.7 l · s 3 bar 3’000 min 0.050 m 0.040 m 2.61 kW 2.09 kW ≥ 80 %
Q ∆p N r r Ph Pm η
Discharge Pressure drop Runner rotational speed Runner outer radius Runner inner radius Hydraulic power Mechanical power Hydraulic efficiency
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.
Consumption area Altitude difference
Release valve
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 of counter-rotating runner pairs can be number of stages 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.
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
I
I
2 I
I
I
(2)
(3)
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 calculated as:
of this energy conversion is (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).
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)
y = f(x): circumferential position [m]
0
cotan
(11)
cotan
(12) /
0/0
x: axial position [m] Fig. 4. Definition of the skeleton-line
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
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
1 3
1
(13)
The final blade contour is given by the thickness distribution of a standard NACA 4–digit hydrofoil. The local
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. /
rows are interpolated to close the 3D spline contours and used to define the blade surface. Define Parameters
Turbine GUI
Additional Functions
Save Parameters Draw 3D Display Performance Calculations
/2
SkeletonLine Calculations
/ /2
Profile Calculations
Save Runner Geometry
/
Fig. 5. Upper and lower profile contours. 3D splines
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.
CAD API
TABLE II SETUP PARAMETERS OF THE DESIGN SOFTWARE Shroud radius Hub radius Runner separation Blade clearance gap Blade thickness Minimal edge radius Wrap angle type Profile type Pressure drop Discharge Design efficiency Runner width Rotational speed Number of blades Runner width Rotational speed Number of blades
50 40 10 0.2 4.5 0.5 Fixed NACA-XXX4 3 8.7 85 Runner A 15 3000 5 Runner B 20 3000 7
mm mm mm mm mm mm
bar l·s %
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.
mm min
mm min
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
TABLE III NUMERICAL SCHEME Simulation type Spatial scheme Turbulence Model Residual Target
Steady 2nd order specified blend factor:1 SST 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.
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B. Computational domains and spatial discreetization The computational domains are illustrated in Fig. 7. Stator 1 domain consists of the inlet pipe with thee static hub that houses the electrical generator of the first rrunner. Then, the Rotor 1 and 2 domains include respectivelyy the inlet and the outlet runners of the micro-turbine. In both R Rotor domains, the hub and shroud are assumed to rotate with thee blades and there is no gap between the tip of the blades and thhe shroud. Finally, the Stator 2 domain consists of the outlet piipe with the static hub region that houses this time the electricaal generator of the second runner.
Interface (GGI) scheme. Finally, a smooth no-slip wall condition is used for all the solid stattic or rotating surfaces. TABLE V BOUNDARY CONDIITIONS Surface
Boundary con ndition
Inlet Outlet Interfaces Solid walls
Q = 6.96 / 7.8 85 / 8.7 / 9.57 [l∙s-1] 0 [Pa] averag ge static pressure Frozen-rotor Smooth no-sllip wall
Stator 1
Rotorr 2
Fig. 8. Domains inteerfaces.
Rotor 1
Stato or 2
D. Numerical simulation results The following results are baased on four numerical simulation setups with different values v for the discharge. Indeed, the values correspond resspectively to 80%, 90%, 100% and 110% of the nominaal discharge. The runner rotational speeds (3000 rpm), corresp ponding to the one defined in the design, has been kept consstant for all cases. Fig. 9 shows the field of the relative velociity streamlines through the whole micro-turbine at the nominal operating point (Q = 8.7 l∙s-1).
Fig. 7. Computational domains.
The global mesh information is provided in Table IV. The adapted unstructured mesh has been generateed with the Ansys ICEM commercial software for each domaain using mostly tetrahedral cells.
Fig. 9. Field of streamlines as an illustrativ ve result of the flow simulation
TABLE IV SUMMARY OF SPATIAL DISCRETIZATIION Domain
Type
Nodes
Elements
Rotor1 Rotor2 Stator 1 Stator 2
Rotating Rotating Stationary Stationary
546’864 571’559 475’084 557’981
1’458’718 1’491’680 1’193’490 1’415’510
2’151’488
5’559’398
Full domain
C. Boundary conditions The detailed boundary conditions of the whole computational domain, for both the stationaryy and the rotating parts, are provided in Table V. At the inlet off the Stator 1, four different constant discharge values, correesponding to the investigated operating conditions, are impossed. At the outlet of the Stator 2 domain, 0 [Pa] relative averaage static pressure condition is selected. The interfaces between the static and rotating domains (see Fig. 8) are treated wiith a Frozen-rotor condition, the connection being ensured by the General Grid
TABLE VI SIMULATION RESULTS
% 78.86 83.08 83.14 81.29
∆
∆
∆
1.50 2.13 2.84 3.63
0.30 0.63 0.96 1.33
1.81 2.76 3.80 4.96
· 7.0 7.9 8.7 9.6
1'210 2'142 3'311 4'762
713 1'084 1'516 2'013
240 0 695 5 1'23 37 1'85 58
Table VI gives an overview of the most important he Best Efficiency Point numerical results. As expected, th (BEP) is found at 8.7 · dischaarge, which corresponds to the design parameters. One may statte here that the first runner recovers more mechanical power thaan the second one. Indeed, the main part of the static pressure drop d is created by the first runner. A maximum efficiency of 83.14% 8 is reached for the nominal discharge, the mechanical power p being 2’753 W.
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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.
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 HESSO Valais/Wallis, a hydraulic test rig was installed to perform
+C +Z
+Y
+X
d
+B
c
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
a b
Fig. 11. Generated tool path for the roughing operation (left) and solid simulation of the 5-axis milling (right).
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
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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)
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 1 1.3 2 2.5 3
3.95 5.63 6.77 7.9 9.22 9.8
1 1 1.18 1 1 1.18
1010 1493 1749 2003 2499 2257
50 51.5 50.5 52.8 52.9 53
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 H = 0.5 / 1.3 / 2 / 2.5 / 3 [bar]
250 500 750 1000 1250 1500 1750 2000 2250 2500 2750 3000 3250
α
50
125 175
150
213 325
375 350
250
638 650
625 525
350
975
700
1050 1625
0.20 0.35
0.50
2857 2500
2308 3077 2647
2500 2338
1950
650
and the
2059 2000
1625
Fig. 14. Hydraulic efficiency as a function of the speed factor for a testing head of 1.3 [bar]. discharge factor
1500
1471
1913
1375
1429
1538
1500
1300
875 550
882
1488
1125
500 769
1000 1063
875
450
294 500
3235 3000
2763 0.65
0.85
1.00
1.18
1.54
2.00
2.86
To characterize hydraulic machines, dimensionless values are often used to enable the comparison between different operating points or different model scales. For the microturbine, the discharge factor and the speed factor are
Fig. 15. Relative hydraulic efficiency as a function of the discharge head .
and the
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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.
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 and can be formulated as: angle
VII. DISCUSSION OF THE RESULTS
TABLE IX SUMMARY OF DESIGN, CFD AND TEST RESULTS Value [%] ∆ [bar] [ls-1]
Design
CFD BEP
Tests BEP
85
83
~53
3
3.8
3
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.
(18) The mentioned relationship is represented in Fig. 16. At the reaches a value of 90°. To point of highest instability 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.
Inlet flow angle of the second runner [°]
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.
Outlet flow angle of the first runner
°
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 Number of Elements Number of Nodes Viscosity Model Computing time on standard PC
Non structured, triangular 46’129 25’225 k-epsilon (2eqn) ~70sec
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executing a large number of simulations inn a short time to obtain an iterative optimization process of thee blade geometry. The main parameters of the employed 2D D flow simulations are given in Table X. An example of the rresult is shown in within the runners Fig. 17. The evolution of the static pressure w is represented by a contour plot along with thhe direction of the relative flow velocity indicated by a vector field. Finally, the global performance can be approximatelyy determined by interpolating the simulation results of seveeral 2D domains between the inner and the outer radiuses.
IX. CONCLUSSION Runners of a multi-stage miccro-turbine with counterrotating runners for drinking waater networks have been designed based on a simplified onee dimensional flow model. By using 3D numerical flow simulaations, the performance of the designed runner geometry has beeen analyzed. The required hydraulic efficiency (> 0.8) has been successfully verified. hined on a 5-axis milling Runner prototypes have been mach center and finally tested on a hyd draulic test rig, where the turbine characteristics have been measured. Due to unconsidered mechanical losses, the measured hydraulic mulated one. To cope with efficiency is found lower than the sim this, improvement of the prototype bearings and measurements of mechanical losses are going to be performed in a future step. To conclude, firstt insights of the hydraulic design optimization is presented d using blade cascade simulations. X. REFERENC CES
...
...
Fig. 17. Pressure contour and vector field of relative veloocity as a result of the 2D blade cascade simulation.
Fig. 18. Quasi-three-dimensional analysis of the turbinne performance, used for iterative optimization of the runner geometry. Contours of the static pressure are presented for different radial ppositions.
[1] C. Münch-Alligné, S. Richard, B. Meiier, V. Hasmatuchi, F. Avellan, “Numerical simulations of a counter rotaating micro turbine”, Advances in Hydroinformatics, P. Gourbesville et al. (eds.), Springer Hydrogeology, p 363-373, 2014 [2] V. Hasmatuchi, C. Münch, S. Gabathuleer, S. Chevailler, and F. Avellan, “New Counter-Rotating Micro-Hydro Turbine for Drinking Water Systems”, Hidroenergia 2014, Istanbul, Turkey, T 2014. [3] D. Melly, R. Horta, C. Münch, H. Biner, S. Chevailler, “Development of ng Micro-hydro Turbine” XXI a PM-Generator for a Counter-Rotatin International Conference on Electrical Machines, M Berlin, Germany, 2014. [4] V.Hasmatuchi, F. Botero, S. Gabathuleer and C. Münch, “Design and Control of a New Test Rig for Small Power Turbomachines”, Hydro 2014, Cernobbio, Italy, 2014. ma, “Design optimization of axial [5] G. Peng, S. Cao, M. Ishizuka, S. Hayam flow hydraulic turbine runner”, Intern national Journal for Numerical Methods in Fluids, pp. 517-531, June 200 02. [6] P. Drtina, M. Sallaberger, “Hydraulic tu urbines-basic principles and statof-the-art computational fluid dynamics applications, Proceedings of the Institution of Mechanical Engineers, Part P C: Journal of Mechanical Engineering Science, vol. 213 no. 1 85-102, 1999 dbuch 2009/2010, München, Carl [7] H. B. Kief, H. A. Roschiwal, CNC Hand Hanser Verlag, 2009.
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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.
