ACTIVITY REPORT 2016 – 2017 · with PIV or LDA is our daily business at the highest level. For valve research we dispose of appropriate research loops conforming to the relevant

ACTIVITY REPORT2016 – 2017

Graz University of TechnologyInstitute of Hydraulic Fluid Machinery (HFM)Kopernikusgasse 24 / 48010 Graz, Austria

T +43 316 873 7571F +43 316 873 [email protected]

Graz University of Technology


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ACTIVITY REPORT

2016 – 2017

II

Publisher:

Institute of Hydraulic Fluid Machinery


Kopernikusgasse 24/4

8010 Graz

Austria

T +43 316 873 7571

F +43 316 873 7577

[email protected]

www.hfm.tugraz.at

Head of the Institute: o. Univ. Prof. Dipl.-Ing. Dr. techn. Helmut Jaberg

Vice Head of the Institute: Assoc. Prof. Dipl.-Ing. Dr. techn. Helmut Benigni

Editorial:

o. Univ. Prof. Dipl.-Ing. Dr. techn. Helmut Jaberg

Cover picture: PIV Measurement

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Helmut Jaberg o.Univ.-Prof. Dipl.-Ing. Dr.techn.

Head of the Institute of Hydraulic Fluid Machinery (HFM)


Foreword

Practice-oriented basic research and teaching Institute of Hydraulic Fluid Machinery Graz University of Technology As mechanical engineers, of course, the machines are very important to us. But most important are people: PhD students and undergraduate students of the institute have achieved considerable leadership positions in the industry for hydraulic machinery in Germany, Austria and Switzerland. Everything indicates that this great trend continues. Practically all graduates continue to work in the field of hydraulic turbomachinery with its many facets: With the manufacturers of turbines, pumps and valves, with the operators with their versatile tasks and also with the planners. Fortunately, a number of associates could be held at the institute - nota bene also after their dissertation - so that we can be proud of a core team of highly experienced engineers placing their skills in numerical analysis, laboratory and field measurement at the disposal of hydro power, pumps and valve industry and providing solutions to most challenging problem statements by practice-oriented research.

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At the source a boy was sitting: Likewise we are happy to attract students to work at the institute, for diploma theses and as doctoral students - almost always working together with the Industry. As in previous years, this activity report for the years 2016/2017 outlines the various activities and innovations at the Institute of Hydraulic Turbomachinery at the University of Technology Graz. We claim to have achieved the Champions League level of European research and development facilities for hydraulic machinery and systems. Now we want to win the Champions League! For this purpose we have taken significant steps in the two years of this activity report. We have significantly expanded our IEC test rig to three parallel test set-ups for all conceivable turbine and pump types: Model tests with Pelton, Francis, Kaplan, pit and cross flow turbines have already been IEC-tested, improvements explored and directly been verified. We have very extensively invested in new floating bearings for all machinery types so that we now can measure torques much more accurately and reliably without having to calibrate bearing and sealing effects. Flow visualization with high-speed cameras (also in the field!) with PIV or LDA is our daily business at the highest level. For valve research we dispose of appropriate research loops conforming to the relevant standards so that - as one of very few institutions - we can determine their cavitation characteristics by means of acoustic measurements. We have also upgraded our system measurement technology so that we now have all the IEC measuring techniques available and can also investigate large scale plants – what we did in fact in the last two years: thermodynamic measurements with ultra-modern measuring technology (record: 0.3 per thousand repetition exactness!), Acoustic Doppler measurements (with the only mobile four-path system, ever!) and the index measurement. And our newest baby was also used for the first time: the Gibson method for flowrate measurement by means of a pressure surge. Our research is practice oriented. In numerical analysis, we dispose of no less than 10 commercial ANSYS licenses and claim to investigate stationary and transient flow phenomena with the highest quality: Our analyses and predictions have NEVER been of the mark and have always agreed with measurements. With our own program development for hydro power we have decisively optimised the commercial Flowmaster license and can conduct reliable research on power plants and pipe-

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lines. Significant hydro power plant improvements have been investigated and have been made possible in case of refurbishment, notably secure operation while increasing the annual production in all operating modes. An especially excellent reputation the institute has won by research on hydro power plant operation with the hydraulic short circuit that needs to be run because of the increasingly important grid stabilization and because we are able to detect these transient phenomena in the machine and the lines numerically on our computers as well as experimentally in the field. In all modesty we may report that our activities have been perceived in an international frame, we also give it a little shove and make our research results accessible to the scientific community in relevant journals and at global conferences, very often with an own booth on connected exhibitions: On IAHR in Grenoble (and most recently in Kyoto), A & M University Houston, VDMA Euro Pump conference (where with five contributions we were the most frequent presenter of all), Hydrovision Denver, Hydro Power Conference in Montreux, Seville (recently in Gdansk), Marrakech and in Da Nang we exhibited as well as on regional conferences, among others Renexpo in Salzburg. Also in other contexts, we pass on our knowledge: The distant learning university programme “hydropower”, leading to M.Eng. or to the Academic Hydropower Engineer, was started for the third time, the next course in German language is planned for the coming spring 2019. There are already applications from employees from premium companies. We would like to take this opportunity to thank our many speakers from industry (mainly from operators) who share their knowledge and experience so generously like a good colleague. One sees: Capital region is and will remain central Europe: Austria, Germany, Italy, Switzerland. This is not only because of the spatial proximity but also due to the fact that in exactly these countries work is performed at the world's leading technology for hydraulic machinery and systems. As before, we are proud to contribute important and always practice-relevant work. Here the success story of recent years could be successfully continued during the last two years: This activity report does show it. Institute management and staff are very proud as virtually all our research work is in use in the industry - be it in the form of turbines and pumps with improved efficiency or cavitation behaviour, or be it with particularly low-vibration operation

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(e.g. Francis part load behaviour and axial force) or when solving actual problems in large scale systems, also with the most diverse valves. A rolling stone gathers no moss. This activity report is an incentive for us to become even better in the future: To be able to work on even more complex questions, especially on the urgent topics of reliability and flexibility. To offer better and better support to operators and manufacturers in their daily business. And our students and our PhD candidates better and better future opportunities. Above all, we would like to thank our team with all internal and external employees and speakers for their cooperation – full of energy. Graz, November 2018

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Table of contents 1. Staff ................................................................................................................................ 1

1.1 Management of the Institute .......................................................................................... 1

1.2 Academic Staff ............................................................................................................... 2

1.3 External Lecturer ........................................................................................................... 3

1.4 Technical staff ................................................................................................................ 6

1.5 Administration ............................................................................................................... 7

2. Laboratory and test rig .................................................................................................. 8

2.1 Equipment measurement techniques ........................................................................... 8

2.2 Micro test rig for teaching and testing measurement ................................................. 9

3. Research Projects.........................................................................................................10

3.1 “On-site measurements - efficiency”...........................................................................10

3.1.1 Introduction...................................................................................................................10

3.1.2 Current meter method ..................................................................................................13

3.1.3 Volumetric method .......................................................................................................15

3.1.4 Acoustic Doppler method for discharge measurement .............................................16

3.1.5 Winter-Kennedy method, index method ......................................................................19

3.1.6 Iductive flow measurement (electromagnetic) ............................................................20

3.1.7 Thermodynamic method ..............................................................................................21

3.1.8 Pressure-time method (Gibson)...................................................................................25

3.1.9 Conclusion ....................................................................................................................27

3.2 “On-site measurements – Sound and vibrationmeasurements” ...............................28

3.3 Research on low head turbines – Kaplan PIT turbine ................................................31

3.3.1 Background...................................................................................................................31

3.3.2 Model turbine and test rig installation .........................................................................33

3.3.2.1 Pressure measurement ................................................................................................33

3.3.2.2 Velocity measurement ..................................................................................................33

3.3.3 CFD Calculation ............................................................................................................34

3.3.4 Cavitation measurement versus numerical simulation ..............................................35

3.3.5 Results of the pressure pulsation measurements ......................................................36

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3.3.6 Results velocity measurements ...................................................................................37

3.3.7 Summary .......................................................................................................................38

3.4 An analysis of the impact of draft tube modifications on the performance of a

Kaplan turbine by means of Cumputational Fluid Dynamics .....................................39

3.5 Benchmark study of automotive oil pumps by means of experimental

investigations ................................................................................................................43

3.6 Research on mechanical seals with a custom made endurance test system ...........47

3.7 Research on high head process pumps – simulation and optimisation of side

channel pumps .............................................................................................................50

3.8 Research on canned motor pumps – simulation, optimisation and experimental

validation .......................................................................................................................53

3.9 Research on valves – combining measurement and simulation to get essential

information on valve behavior .....................................................................................56

3.10 Cooling jacket for electric mobility ..............................................................................60

3.11 Research on energy optimisation in passengers cars – simulation of a gear box

oil pump.........................................................................................................................63

3.12 Research on safe operation of hydraulic systems by means of sophisticated

transient analysis..........................................................................................................65

3.13 Multiphase Simulation of the hydraulic behaviour of a weir ......................................69

3.14 Increasing power output and flexibility of existing high head power plants with

help of waterhammer simulations ...............................................................................71

3.15 Improvement of intake structures with numerical simulation....................................74

3.16 Optimisation of Pelton distributors with the help of 3D CFD .....................................77

3.17 Investigation of hydraulic losses at the connection of a surge tank and penstock

for different load cases with the help of CFD ..............................................................79

3.18 Research project Francis Turbine Virtual + ................................................................81

3.18.1 The model test ..............................................................................................................82

3.18.2 Numerical simulation....................................................................................................85

3.19 Numerical investigation and imporvement of cross-flow-turbines (high head

application) ...................................................................................................................90

3.20 Test rig for guide vane seals ........................................................................................93

3.21 Experimental and numerical investigation of a single blade impeller pump ............95

3.22 Investigation of a 2-nozzle Pelton turbine ...................................................................98

3.23 3D-scanning and reverse engineering of hydraulic sufaces .................................... 102

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3.24 Investigations of the intake structure of the power plant GKI/Inn ........................... 105

3.25 Analysis of the leakage behaviour of Francis turbines and its impact on the

hydraulic efficiency - A validation of an analytical model based on CFD-results .. 109

4. Graduation work ......................................................................................................... 112

4.1 Masterthesis ................................................................................................................ 112

5 Continuing education ................................................................................................. 118

5.1 University Programme in Hydropower ...................................................................... 118

5.1.1 Master of Engineering (MEng) – Hydropower (5 Semester) ..................................... 120

5.1.2 Akademische Wasserkraftingenieurin/Wasserkraftingenieur.................................. 120

5.2 pump.ing – Vocational training for pump experts .................................................... 122

5.3 Practitioners’ conferences ......................................................................................... 125

6 Publications ................................................................................................................ 127

6.1 Articels in journals ...................................................................................................... 127

6.1.1 2017 ............................................................................................................................. 127

6.1.2 2016 ............................................................................................................................. 127

6.2 Contributions at conferences .................................................................................... 128

6.2.1 2017 ............................................................................................................................. 128

6.2.2 2016 ............................................................................................................................. 129

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1. Staff

1.1 Management of the Institute Helmut Jaberg O. Univ.-Prof. Dipl.-Ing. Dr. techn. Head of the Institute Professor at the Graz University of Technology (since 1995) as well as managing director and partner of an international consulting company. After having studied aerospace engineering in Stuttgart, Southampton and Munich, he worked as design engineer at the German turbine manufacturer MTU München GmbH. After the accomplishment of his doctoral thesis at the Stuttgart University he worked as head of the ‘Department of central research and development’ at the KSB AG and, from 1988 onwards, as Head of development in KSB’s German-French business unit ‘Building services engineering’. Helmut Benigni Assoc. Prof. Dipl.-Ing. Dr.-techn. Vice Head of the Institute Studied mechanical engineering at the Graz University of Technology, specialisation in numerical simulation, PhD thesis in optimisation of hydraulic machines. In post-doctoral position responsible for hydraulic machine simulations by means of CFD methods, development of different hydraulic designs and of machine configurations, habilitation in hydraulic fluid machinery. Vice-head of the Institute of Hydraulic Fluidmachinery, Graz University of Technology.

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1.2 Academic Staff

Christian Bodner

Dipl.-Ing.

Mark Guggenberger

Dipl.-Ing.

Stefan Höller-Litzlhammer

Dipl.-Ing.

Philipp Günther Kandutsch

Dipl.-Ing.

Stefan Leithner

Dipl.-Ing.

Markus Mosshammer

Dipl.-Ing. Dr.-techn.

Jürgen Schiffer-Rosenberger

Dipl.-Ing. Dr.-tech.

Florian Senn

Dipl.-Ing.(FH)

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1.3 External Lecturer

� Wolfgang Bach Dipl.-Ing. fo. GE Energy Lecture: Converter technology

� Thomas Beyer

Dipl.-Ing. Vatenvall Wasserkraft GmbH Lecture: Operational safety, events of malfunction and damage

� Andreas Blauhut

Mag. Verbund VHP Lecture: Planning and licensing of hydropower plants

� Bruno Buchmayr

Univ.-Prof. Dr. Montanuniversität Leoben Lecture: Materials for mechanical engineering

� Siegfried Demel

Dr. fo. Verbund VHP Lecture: Operational management of hydropower stations

� Lothar Fickert Prof. Dr. TU Graz Lecture: Basic of electrical systems

� Marco Fiegl

Dipl.-Ing. Verbund VHP Lecture: Construction management

� Erwin Franz

Dipl.-Ing. Hydro Power Consultant Lecture: Machinery components

� Rudolf Fritsch

Dipl.-Ing. ZT Fritsch GmbH Lecture: Small hydro power

� Oliver Haupt

Dipl.-Ing. KfW Lecture: Concept an feasibility studies

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� Günther Heigerth Prof. Dr. TU Graz Lecture: Concepts of power plant surge tanks

� Georg Klapper

Ing. Künz Lecture: Hydraulic steel structure and moveable machinery

� Helmut Knoblauch

Dr. TU Graz Lecture: Structural civil components, Structures and sealing systems

� Wolfgang Kofler

Dipl.-Ing. TIWAG Lecture: Energy transport and switching mechanism

� Hansjörg Köfler

Prof. Dr. TU Graz Lecture: Basic of electrical systems

� Werner Ladstätter Dipl.-Ing. Andritz Hydro GmbH Lecture: Basic of electrical systems

� Florian Landstorfer Dipl.-Ing. Verbund VHP Lecture: Site safeguarding and monitoring

� Peter Maydl

Prof. Dr. TU Graz Lecture: Construction material

� Josef Mayrhuber

Dipl.-Ing. Dr. Verbund VHP Lecture: Materials for mechanical engineering

� Peter Meusburger

Dipl.-Ing. Dr. EW-Frastanz Lecture: Operational safety, events of malfunction and damage

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� Paul Oberleitner Dipl.-Ing. Zivilingenieur für Bauwesen Lecture: Concept and feasibility studies

� Gerhart Penninger

Dipl.-Ing. Dr. Verbund VHP Lecture: Operational safety, events of malfunction and damage, power house and cavern, storage and pumped storage power plant

� Milan Pudar Dipl.-Ing. Dr. Magna Graz Lecture: Joining technology, welding

� Christian Reszler Dipl.-Ing. Dr. JR AquaConSol GmbH Lecture: Joining technology, welding

� Edgar Röck

Dipl.-Ing. TIWAG Lecture: Energy economics

� Christian Schenk Dipl.-Ing. Dr. TIWAG Lecture: Machine dynamics and bearings

� Rudolf Thalhammer

Dipl.-Ing. Dr. ABB Lecture: Control technology for hydraulic machinery

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1.4 Technical staff

Alfred Krisper

Electrical engineering, Prototyping

Kurt Mass

Ing. Electrical and electronic engineering, IT

Josef Papst

Mechanical manufacturing

Erwin Pischler

Mechanical manufacturing

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1.5 Administration

Martina Halser

Mag. Secretariat

Barbara Towpyha

Secretariat

Melanie Köll

Secretariat

Jeannine Pöschl

Secretariat

Andreas Stachel

MSc Marketing & Sales

Margot Jaberg

Mag. Management University Programme in Hydropower

Karin Hermann

Mag. Management Pump Engineer

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2. Laboratory and test rig

2.1 Equipment measurement techniques

Velocity field measurement

• Particle Image velocity (PIV)

• Laser Doppler Anemometry (LDA)

• 5 hole probe

• Hot wire anemometry

• High speed camera

Flow rate measurement

• Differential pressure method (Orifice plates or Venturi tubes)

• Magnetic inductive flow measurement device

• Acoustic Doppler Method, transit time difference method

• Current meter acc. KIEC 60041

Pressure measurement

• Static pressure sensor

• Dynamic pressure sensor

• Highly dynamic pressure Sensor

Vibration measurement

• Vibration acceleration

• Vibration velocities

• Vibration displacement

Force – Torque measurement

• Load cell

• strain gauges

• different torque measurement flanges and –shaft

Temperature measurement

• Thermocouple

• PT 100

• IR- thermometer

• Seabird sensors, highest accuracy

Data acquisition and post processing software

• LabView®, Matlab®, Diadem®

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2.2 Micro test rig for teaching and testing measurement equipment

To get the best overview of the functionality of the 4-quadrant main test rig, the institute of hydraulic fluid machinery owns a minimized test rig with the same measurement equipment installed as at the main test rig. This test rig is used for lectures on hydraulic fluid machinery measurement techniques and shows students the behaviour of pumps together with the pipework. Since every equipment is installed in a small area the students are able to compare analogue measurement techniques with digital ones and can discuss the advantages and disadvantages of both. An asynchronous motor drives the installed pump at variable speed. The flowrate can be measured by a magnetic inductive flow meter or by a Venturi nozzle. For measuring the pumps head or the losses of a specified pipe section, water column gauges or inductive pressure transmitters can be used.

Figure 1: Micro test rig with the necessary hardware and measurement equipment installed as in the main test rig

Since the design of the test rig is very simple, the piping can be easily changed and allows to measure the losses of different pipe fittings or valves. It is also the best way to observe and measure cavitation effects caused by the flow pattern inside the pump impeller. Furthermore it is possible to measure the influence of cavitation on the valve- or pipe-fitting-characteristics itself.

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3. Research Projects

3.1 “On-site measurements - efficiency” 3.1.1 Introduction

Turbines in hydropower plants achieve peak efficiencies of far more than 90%, in larger power plants more than 95% depending on the speed and the type of machine. The research topic “On-site measurements” verify the quality of hydraulic turbine performance as well as the measurement technique for specific turbine installations. Based on model tests or numerical simulations efficiency of a prototype machine has to be guaranteed and measured with the best measurement technique available. However, in order to be able to answer what the best measurement technique is, not only the methodology must be known in detail as the limitations are of even more interest. And this does not only depend from the specific installation in a power plant. In the scope of refurbishment, expansion or revision of hydropower plants, the determination of the actual efficiency curve also provides a fundamental contribution to the analysis of the optimisation potential during the planning phase and to the evaluation of technical measures implemented. In addition, to achieve the maximum efficiency at the design point, nowadays modern hydropower plants have to meet high efficiency requirements, particularly in part-load operation. These operation points are taken into account by different prioritisations of the operating points guaranteed in the respective operating range. During commissioning or during the acceptance test of a new installation, the following questions arise, among others: Does the newly constructed hydropower plant reach the projected efficiency levels? Are the guaranteed efficiency values met? Is there any cavitation occurring in operating points off the best efficiency point? Do hydraulic forces cause vibrations and noise? What are the pipeline or waterway losses? A system measurement provides answers and reliable data, and through expert interpretation various questions can be clarified. Especially losses in pipelines

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are often underestimated and lead to operation of turbines beyond the optimum flow rate range. If refurbishment is performed for a hydropower plant, usually determination of the absolute efficiency value is not the main target. The focus is rather set on a performance increase or a change of the relative efficiency. Therefore, it is essential to determine the actual state before modifications are realised in order to allow for a realistic estimation of the increase in efficiency. The choice of the appropriate efficiency measurement method does not only depend on the expected measurement uncertainties and tolerances set in a contract, it is also a question of the effort for such a measuring campaign and furthermore of technical challenges e.g. limitations imposed by plant operating conditions as draining of the system or constant load operation. Applicable regulations are the IEC standards 60041: "Field acceptance tests to determine the hydraulic performance of hydraulic turbines, storage pumps and pump turbines" and 62006: "Hydraulic machines – acceptance tests of small hydroelectric installations". With regard to hydraulic turbomachines, modern measuring technology offers different methods for efficiency measurements at a turbine unit. Especially through the development of real-time capable data acquisition systems, it is possible to perform high-precision measurements rather quickly, whereby minimal downtimes can be achieved. The prerequisite for carrying out efficiency measurements are existing measuring connections or suitable measuring sections, which, however, can easily be provided in the planning phase. Depending on the accuracy, purpose and applicability required, different measuring methods are used. Individual framework conditions, such as characteristics of the measuring range or the measuring instrument itself, determine the choice of measuring method for the particular requirement. The following Table provides a comparison of the typical measurement uncertainties of the different measurement methods under standard boundary conditions. Additional absolute flow rate measurement methods are weirs, differential pressure measurements with Orifice plates, nozzles, Venturi tubes or tracer methods. Relative flow rate measurements can also be realised with a needle stroke on Pelton turbines or with accoustic Doppler methods with one or two paths. The methods can basically be divided into relative and absolute measuring methods, whereby only the Winter-Kennedy method represents a relative method which does not provide any information about the absolute

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Table 1: Overview of measurement methods

Method

IEC

Ab

so

lute

or

Exp

ecte

d

syste

mati

c

err

or

≈fη

Comment

Measurement

campaign carried out

or analysis/witness of

measurement

campaign

Current meter

60041 Abs

.

1,3-

2,5%

More precise in

closed conduits

HPP Meitingen (2016)

HPP Hoosic River

(2017)

Volumetric

method

60041 Abs

. 1,5% For artificial

reservoirs

PSP Langenprozelten

(2017)

Acoustic Doppler

flow

measurement

(60041)

62006

Abs

.

1,0-

2,5%

Acc. to standard at

least 4 paths or

other mutual

agreement. Better

accuracy if pipe

inside mounted

sensors.

HPP Prančeviči, Francis

(2017)

HPP Codalonga, Pelton

(2017)

HPP Suman, Kaplan

(2017)

HPP Schiester, Crossfl.

(2016)

Winter-Kennedy

method

60041 Rel

.

1,0-5,0

%

Only for relative flow

measurement HPP Reisach (2016)

Inductive flow

measurement

device

(electromagnetic)

62006 Abs

. 1,5%

For small pipe dia.,

only meas.

campaign < DN500.

Permanently

installed up to DN

2000

HPP Jerzens (2017)

Thermodynamic

efficiency

measurement 60041 Abs

.

0,6-

1,3%

Minimum head

required 100 m PSP Kaunertal (2018)

Pressure time

method (Gibson) 60041 Abs

.

1,5-

2,3%

More precise for

regular pipe

diameters

Lab, Master thesis

(2017)

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efficiency level without an additional calibration. In addition, the measuring methods are subdivided into conventional measuring methods (volume flow measuring method) and thermodynamic efficiency measurements. 3.1.2 Current meter method Volumetric flow measurements with current meters enable determination of the absolute value of the volumetric flow by measuring the velocity field in a cross-section. This method is used e.g. in closed cross-sections, under pressure and in open channels. For this purpose, current meters are positioned rectangular to the evaluation plane of the bulk flow rate. The velocity component w, which is rectangular to the cross section, is decisive for the flow rate. It is important to ensure that neither the measuring device (current meter and frame) influences the flow, nor that the blockage effect by the measuring probes exceeds a limit value. In most cases, the cross-section is scanned according to a defined grid and the individual speeds in the grid points are measured gradually in time. During the individual measurement as well as during the scanning of the entire speed field the flow should be as constant as possible (constant operation point) in case that not all measurement points can be measured at once. The number of impeller revolutions per unit of time allows for the local flow velocity to be determined. In most cases flow velocities close to walls cannot be measured and therefore a wall treatment with a suitable power law has to be applied. The cross-section must be resolved at least by 25 measurement points. On the basis of numerical weighting factors as well as the exact knowledge of the cross-sectional area, the flow is determined by integration. This method can only be used for artificial channels. Efficiency measurements with current meters are mainly used for low-head systems. A measuring setup developed at the Institute enables the synchronous scanning by up to 16 current meters as well as the visualization of the velocity distribution during the measurement campaign. This way inhomogeneous flow distribution are detected during the measurement and thus costly re-measurements can be avoided. As an example, the power plant HPP Meitingen can be mentioned. There, before

Figure 2: Current meter, new installation on-site measurement, witness test

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and after the refurbishment, an on-site measurement of a unit by means of a flow measurement with current meters was carried out, which confirmed the previously forecast improvement in efficiency. In this case, a horizontal displacement of the current meters was chosen, because the horizontal position could easily be adjusted with the trash rack cleaning machine, where the stable support was mounted.

Figure 4: Left – installed current meter, right – Hoosic River

Figure 3: Current meter, new installation on-site measurement, witness test

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As another example the Hydro Power Plant HPP Hoosic River could be mentioned, which is located in Pownal, a town 225 km from Boston in the Southwestern part of Vermont (USA). The commissioning of the power plant took place in the second half of 2017. There is one Kaplan turbine operated in the HPP to gain energy from the Hoosic River water. The turbine with a concrete intake is positioned at the end of a steel penstock with a length of about L=28m and a nominal diameter of D=2.43m. The field measurement was conducted according to the standards IEC 60041, ASME PTC 18-2011 and ISO 3354. 3.1.3 Volumetric method The pump storage hydro power plant PSP Langenprozelten, for example, was measured after refurbishment with the volumetric method in accordance to IEC 60041. The volume change was recalculated for a measurement period during which a suitable level change in the artificial upper and lower reservoirs, as well as a known correlation between the level height and the surface area are caused, which is shown in the following Figure. All data measured was recorded synchronously in order to determine the flow rate change and head change with constant power control during the measuring period. The change of the specific hydraulic energy due to the level difference was 0.32% (∆Eh <1% according to the standard). The measurements were carried out within a level range of the reservoirs in which the surface areas of the reservoirs have a nearly linear function (see following Figure). For the start and end values of the level heights, the temporary level change was approximated by a regression line. Based on this measurement principle, the level at the beginning and at the end of the measurement was determined at given times.

Figure 5: Upper and lower reservoirs

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3.1.4 Acoustic Doppler method for discharge measurement A very flexible method for determining the flow rate, without any effect on the flowing medium itself, is the acoustic flow measurement by means of the transit time difference method. In a non-flowing fluid the pressure changes of a sound signal propagate in all directions with the speed of sound. In moving liquids, however, the relative velocity of the sound changes with the fluid velocity. If the flow direction of the medium w is in the same direction as the sound waves, the velocities are superimposed to a + w. For a propagation against the direction of flow, the resulting velocity is a - w. The resulting transit time difference is a measure of the mean flow velocity wmean. So, the calculation of the flow rate is easily possible with a known area A for the cross-section of the pipe.

Figure 6: Upper and lower reservoirs and area versus level for upper and lower reservoirs

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For this purpose, a sound pulse is sent obliquely to the flow direction through the measuring cross-section (The sound impulse passes the medium with an angle

β). The sound impulse passes the complete measurement speed profile. The mean reproductive rates are

� � �� ∙ �� ∙ ��

��

with the runtimes t1 and t2. However, since even small temperature fluctuations change the speed of sound

considerably (change of sound velocity a ≅ 3.7m/s for water, for a temperature

difference of ∆Θ = 1°C), one can subtract the change of the sound velocity. The result is an equation which is independent of the speed of sound:

��

��

This can be achieved by simultaneously measuring the transit times t1 and t2. In addition to this measurement principle (transit time difference method), there are a number of others like the phase difference method, the Doppler effect method or the signal correlation method. The advantage of the Acoustic Doppler measurement method is the flexible determination of the flow rate even with large pipe diameters, however there are limiting factors such as low flow velocities and inhomogeneous velocity distribution in the cross-section of the pipe. An exact positioning of the sensor together with a comparably long inlet distance is a must. Also, knowledge on the wall thickness and the coating when using clamp-on systems is necessary. In the latest IEC Standard 60041 this method is listed in the appendix and is therefore not intended to be applied as the primary method. An absolute efficiency measurement is only possible if the volume flow determination is carried out with four paths. This measuring method is frequently used in practice. Clamp-on sensors, which are mounted magnetically on the outer wall of a steel pipe or by means of a tension belt on a GRP pipe, have the advantage that no pipe modifications have to be undertaken. Plant operation remains undisturbed by this non-invasive measuring method. To reduce measurement uncertainties, additional numerical analyses of the velocity profile in the measuring cross-

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section can optimise the integration of the volume flow. The following pictures show clamp-on acoustic flow measurements of different turbines types and different types of penstocks.

Figure 7: HPP Prančeviči, Francis turbine, steel pipe, D = 1.60 m

Figure 8: HPP Codalonga, Pelton turbine, steel pipe, D = 0.40 m

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3.1.5 Winter-Kennedy method, index method

The index method is an indirect flow rate measurement technique commonly used on Francis turbines, Francis pump turbines and Kaplan turbines. This easy-to-implement method is based on a differential pressure measurement and two pressure measuring points and allows for the determination of the relative change of the flow. The relationship between the measured pressure difference of both measuring points and the flow can be determined with an absolute volumetric flow measuring method and subsequently the index measuring device can be calibrated proportionally to this relationship. The differential pressure depending on the flow rate is described by the Bernoulli equation and the following correlation results (according to the standard IEC 60041)

� � ∙ ∆�� with n=0.48…0.52

The arrangement of the pressure measuring points depends on the type of machine to be measured. Regardless of the arrangement made, it is important to ensure that the arising differential pressure is sufficiently high. The relative measurement method is also referred to as the Winter-Kennedy method and is used for double-controlled machine sets to determine the optimal relationship between the runner and the guide vane positions. As already mentioned, in order to be able to carry out acceptance measurements with the help of the index method, prior calibration of the measuring system using an absolute measuring

Figure 9: HPP Suman, Kaplan turbine, GRP pipe, D = 2.363 m

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method is required. General limits are the differential pressure which must be higher than 40 mbar and the condition of existing (old) measuring taps. In practice, a guarantee agreement between operator and turbine manufacturer may include terms regarding the relative efficiency difference before and after the modernisation of facilities. The pump storage hydro power plant PSP Reisach can be mentioned as an example, where the index test was carried out by the turbine manufacturer. The measurements were witnessed and the results were analysed together with the plant operator.

3.1.6 Inductive flow measurement (electromagnetic)

Inductive flow measurement devices use a magnetic field applied to the metering tube, which results in a potential difference proportional to the flow velocity perpendicular to the flux lines. The potential difference is sensed by electrodes aligned perpendicular to the flow and a magnetic field is applied. The liner is non-conducting, however the fluid must be conducting – which is true for the water which feeds the turbine. The magnetic field applied is pulsed, which allows the device to cancel out the effect of stray fields in the piping system. For the HPP Jerzens measurements with a permanently installed inductive flow measurement

Figure 10: HPP Reisach

21

device were analysed and compared to the results of an acoustic Doppler flow measurement performed during an on-site measurement campaign.

Figure 11: HPP Jerzens, installed situation of DN 1400 inductive flow measurement device

3.1.7 Thermodynamic method Efficiency measurements are realised by means of the direct method according to IEC standards. The thermodynamic measurement of the efficiency level is based on the characteristic parameter that all hydraulic losses inevitably lead to an increase of the fluid temperature. Thus, by means of registering the intensive state values pressure and temperature at the inlet and the outlet of the machine, the efficiency level can be determined. In contrast to the volumetric flow measuring methods, the hydraulic power is not related to an external power (generator shaft power) in the thermodynamic method, but to the total energy change in the fluid. The hydraulic efficiency results from the ratio of the flow containing the losses versus the lossless flow.

22

IEC 60041 defines the efficiency of turbines as

! "#"! $ ∆%!%# ∙ "#

With the thermodynamic method the correction term for the hydraulic performance ∆Ph is not necessary – as there is no need for flow rate data to determine the efficiency – and can thus be set to zero. The variables required for the determination of the efficiency are depicted in the following Figure. The various variables are broken down in detail.

Figure 12: Schematic display of the thermodynamic measurement variables

23

According to IEC 60041 the hydraulic performance Ph is defined as

%! &̅ ∙ ( ∙ )* ∙ �+-./

with

&̅ ∙ ( "! �01 � �21*̅ � 3012 � 3212

2 � &̅ ∙ )501 � 521+- 67&/

Within the scope of the thermodynamic method this hydraulic performance is related to the mechanical energy, the latter being defined as:

"# �8)�00 � �20+ � �9̅)T00 � T20+ � 3002 � 32022 � &)500 � 520+ � ;"#- 67&/

The efficiency measurement description will be realised by means of the direct method according to IEC standards. In order to determine the specific hydraulic energy Eh, pressure p10 has to be measured. This is realised by means of a relative pressure sensor. By definition pressure p20 for Pelton turbines is equal to the atmospheric pressure pat. The velocity v10 is calculated based on the iteratively determined flow rate and the given cross-section. The exit velocity v20 is zero by definition. Height levels z10 and z20 have to be disclosed. The specific mechanical energy Em can, inter alia, be determined by measuring pressure p11. According to the standard IEC 60041 this measurement has to be made with the same relative pressure sensor as the measurement of p10 whereas the height level of the pressure sensor installed has to be taken into account as well. Pressure p21 is being measured by means of a level gauge in the downstream channel in the case of an (open) tail water channel. Temperature T11 is determined in the isolated measurement chamber. For the downstream temperature measurement (the IEC standard requires six measurement locations at open channels: T21_1 to T21_6) a special structure must be installed. Velocity v11 is determined by means of the sampling quantity measured and the given cross-section whereas velocity v21 is determined via the iteratively calculated flow rate and the given cross-section. Height level z21 is stored for every single sensor. Term δEm is the total of all correction factors.

24

The calculation of the hydraulic efficiency by means of the thermodynamic method is influenced by three thermodynamic characteristics of water – the density ρ, the specific isobaric thermal capacity cp and the isothermal coefficient a. The formulations of these equations defined by Herbst and Rögener as well as by the International Association for the Properties of Water and Steam are stored in the measurement and analysis software. One characteristic feature of the direct measurement method is the minimum decompression of the fluid between pressure pipe and measurement chamber which results from the flow restriction in the sampling probe and pressure losses in existing pipes (length < 0.2 m) between sampling probe and measurement chamber. This method places highest demand on the accuracy of temperature measurements. In the IEC standards the field of application of the thermodynamic efficiency measurement is limited by a minimum drop or delivery height of 100 meters. The experience and the metrological equipment of the Institute of Hydraulic Fluidmachinery show that measurements are possible from a drop height of 80 m or more. The measuring method can be used for all common types of turbines and pumps. In addition to high accuracy, the biggest advantage of this measurement method is that determination and knowledge of the flow are not essential. High accuracies up to 0.5% are possible with the thermodynamic method and therefore no flow measurement is required. However, the technology is very sensitive to sunlight. A constant temperature of the feed water is absolutely necessary (permitted gradient: 5 mK / minute). The thermodynamic measurement campaign at the HPP Kaunertal before and after the refurbishment can be mentioned as an example for the successful application of this methodology.

25

Figure 13: HPP Kaunertal

3.1.8 Pressure-time method (Gibson) The pressure-time method, formerly also known as Gibson method, is an easy to implement absolute volume flow measurement method. The measuring method requires piped hydroelectric power plants. At least two pressure measuring points at a distance of several meters at the pressure pipeline are necessary. A measurement of a system with free upper surface of the water is possible as well. With a fast regulating device or a shut-off device, a pressure surge is consciously caused and then its pressure profile is detected whereas different measuring positions are analysed.

<= �> ∙ ? ∙@ A∆B�)�+ � ∆BC)�+ � ∆BD)�+EC� � <F

�F

�=

The propagation velocity of the pressure surge in the pipeline is a measure of the flow velocity and thus of the volume flow. If suitable possibilities to access penstocks are provided, the volume flow can be precisely determined up to a deviation of only 2-3% by this method with minimal effort. The pressure profile is recorded in the form of a differential measurement in two measuring cross-sections, integrated over the duration of the momentum change period. With these results the discharge underlying the change in momentum is

26

calculated. The change in momentum is primarily generated by a continuous closing of the shut-off device of a hydraulic machine.

Among the difficulties to be controlled are: the exact determination of the penstock factor for measuring sections, especially if they deviate from the straight tube; the selection of the end of the integration range during measurements in pipelines due to the decay of the pressure surge; the time-synchronous and self-sufficient acquisition and storage of measurement data in both measurement levels; the usage of acceleration-compensated absolute pressure transducers to compensate for system-specific vibrations.

Figure 14: Pressure time method, principle sketch, water hammer measurement

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Figure 15: Pressure time method, lab test (source: Master thesis J. Klausner)

3.1.9 Conclusion When it comes to the application of the most suitable measurement method for a specific task, a certain amount of experience, a precise knowledge and conscientious application of the standard as well as the deliberate non-standard procedure are essential. Therefore, the application of different measurement techniques must be scrupulously questioned and thoroughly analysed in order to answer the many questions that arise during system measurements. Particular attention has to be paid to the research and testing of new measurement techniques or application instructions. At the same time, such measurements provide the possibility to check hydraulic designs and to obtain proof for the compliance with performance guarantees, and generally increase the transparency between the contracting parties by involving independent experts.

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3.2 “On-site measurements – Sound and vibrationmeasurements”

In addition to efficiency measurements, noise and vibration measurements are another possibility to check machine performance. Meeting of all guarantees given is often linked to compliance with noise and vibration limits. Nowadays, it is customary for refurbishment projects to expand the operating range of the units to deep part-load operation, and if vibrations are below the limits specified in ISO 10816-5, there is nothing to prevent operation within these operation ranges. The following two examples show refurbishment projects where measurements were performed for the operator to check the quality of the refurbishment.

Figure 16: Left – before the refurbishment, right – after the refurbishment, plant data

The hydro power plant HPP Pontives is located in South Tyrol's Val Gardena. The penstock is made of steel and has a nominal diameter of DN1200 and a length of L = 150 m. The penstock connects to the power station with a horizontal Francis turbine. The maximum flowrate is currently limited to Q = 2.0 m³/s. The mechanical and electrical equipment was refurbished in 2016. Before and after the refurbishment, an on-site system measurement campaign according to IEC 60041 was carried out. In addition, vibration measurements according to ISO 10816-5 and sound power measurements according to ISO 3744 were performed.

29

During the measurement four sensors were used simultaneously on three planes. The measurement points were at different positions (1: bearings A X-plane, 2: bearings A Y-plane, 3: bearings A Z-plane, 4: turbine cover, 5: bearings B X-plane, 6: bearings B Y-plane). The following Figure shows the measured vibration velocities for the different operating points for the measurements performed before and after the refurbishment. The vibrations occurring at the bearings as well as at the turbine cover have decreased drastically and are now well below the standard value of 1.6 mm/s for new machines, which is rated in the standard ISO 10816-5.

Figure 17: Vibration velocity HPP Pontives, before and after the refurbishment

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As another example for noise measurement, HPP Langenprozelten can be mentioned. The position of the microphone is marked in the following Figure. The maximum noise level measured was 108 dB(A) for the operation mode of 80 MW at the draft tube cone at unit 1.

Figure 18: Position of the microphone, noise level measurement unit M1 and M2

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3.3 Research on low head turbines – Kaplan PIT turbine 3.3.1 Background The horizontal axis 3-blade runner with a nominal specific speed in the best efficiency point of nq = 230 rpm has such a flat characteristic that operation with nq = 260 rpm practically provides the same value as peak efficiency (see Figure 19 (a) for the hill chart). As the built-in situation is very variable and as the encasing of the turbine in concrete is not guaranteed, the structure itself must be stiff and able to resist all possible load cases. Pressure pulsations during operation are of special interest. Using the software package ANSYS CFX steady-state and unsteady simulation methods were tested in a preliminary study in order to define the most promising approach for a reliable efficiency prediction. In the next step a statistical evaluation of the pressure distribution on the runner blades was performed. By

using the so-called “Histogram Method” the Thoma number σ was calculated for

various operation points of the turbine. The results gave a first impression of the expected cavitation performance. Additionally, two-phase simulations were performed to determine the extent of cavitation and its impact on the turbine performance. Furthermore, measurement results obtained from an experimental investigation on a model test rig served as the basis for the evaluation of the results achieved by means of Computational Fluid Dynamics (CFD). Pressure pulsation measurements as well as velocity field measurements at different sections were used for a further cross-check between experimental and numerical results. Fig 19 provides an overview of the installed turbine together with the main components. The flow enters the model section through the inlet tank, which represents the head water of the turbine (see Fig 19(b) and Fig 19(c)) The pit is mounted on the turbine frame and includes the torque measurement shaft and a bevel gear box to transfer the shaft power to the generator during the model test instead of a belt drive for the prototype installation. The flow leaves the pit to the guide vane section in downstream direction. The 13 guide vanes are connected by a guide vane pitch adjustment device and can be adjusted by a hand wheel during operation of the model. Between the gearbox and the runner of the turbine the torque measurement flange was placed, see Fig 19(d). Therefore, the measured torque only includes

32

the runner blade torque and the friction torque of the two sealing rings as well as the two bearings of the turbine shaft. The friction can be measured separately. The flow then passes the runner blades, which can be adjusted when the runner is dismounted. The connection of the runner root to the hub was pinned in order to guarantee the correct position and then fixed with a screw. For cavitation measurements the runner shroud was made out of Plexiglas (see Fig 19(c) for the Plexiglas shroud). This allows for the observation of cavitation at the area around the runner and at the beginning of the draft tube. For flow fields with high Sigma values and for pressure fluctuation measurements this was changed to a housing made of steel. The flow leaves the turbine section towards the outlet tank, which represents the tail water downstream of the turbine and is shown in Fig 19(b).

Figure 19: Hilchart, meridional view and model test

33

3.3.2 Model turbine and test rig installation 3.3.2.1 Pressure measurement For the pressure pulsation measurements in the rotating system 16 pressure transmitters were placed on the runner blades, 8 at the pressure side of the blade and 8 at the suction side of the blade (Fig 20). As of the thickness and for easier cable management, two blades were equipped with pressure sensors. The holes and the channels for the cable management were milled and sealed. The sensors of each blade side were equipped with a 32-pin plug.

Figure 20: Pressure pulsation measurement, position of sensors

3.3.2.2 Velocity measurement

To determine the velocity fields of the turbine a 3D-directional probe, also known as 5-hole probe, was used. With this 3D-direction probe it is possible to calculate the absolute velocity and the orientation at specific points inside the turbine flow by measuring the pressures of five holes at the probe tip (nulling mode). Nine different positions (Fig. 21 left) were measured.

34

Figure 21: Velocity measurement, position of measurement sections

3.3.3 CFD Calculation The unit was split into components (= domains) for the purposes of the CFD-calculation. The calculation started with the pressure vessel in flow direction. This intake part was connected to the pit region, which was connected to the guide vane domain downstream(Fig. 12). The guide vane mesh passage was copied into the model 13 times and connected by means of a 1:1 interface. The mesh for the passage of the runner domain itself was also generated and connected by means of a 1:1 grid interface between the 3 runner blades. The draft tube domain was then connected by the same method as the guide vane to rotor interface. Downstream of the draft tube an additional component, the tail water tank, was connected to the draft tube. The function of this component is the representation of the tail water, which had been chosen to avoid the setting of boundary conditions directly at the draft tube outlet as that would influence the draft tube simulation and prescribe the flow situation. The stand-alone generated meshes were combined to a complete unit for different guide vane positions. Boundary conditions were set for inlet and outlet with a pressure boundary condition and thus the flow rate resulted.

35

Figure 22:CFD Model

3.3.4 Cavitation measurement versus numerical simulation For one operation point of the start geometry cavitation curves were measured on the test rig and numerically simulated. The shape of the measured and

calculated η−/σ−curve is very similar – especially in the region of the steep drop

of the turbine efficiency. The left side shows a contour plot of the numerically calculated vapour volume fraction on the runner blade and on the right side a photographic documentation of the cavitation conditions occurring at the model

test rig is shown for various σ-values.

At comparatively low σ-values of 1.15 to 1.7 the experiment and simulation show quite similar cavitation zones on the blade surface. Apart from that, significant differences between the cavitation detected in the experiment and in the

simulation can be found at higher σ-values. During the CFD-simulation incipient

cavitation was detected at approx. σ = 4 while the σi-value found in the course

of the experiment accounts for approx. σ = 5.5. However, the practical usability

of the σi-value is relatively low since a certain amount of cavitation is typically allowed even for the normal operation of Kaplan turbines due to economic reasons.

36

Figure 23: Pressure pulsations: left – test rig, right – CFD calculation

3.3.5 Results of the pressure pulsation measurements A high pass filter with 2 Hz was used for the experimental data and normalised with the rotational speed of the turbine. Afterwards, the signals were analysed by using an FFT algorithm. The draft tube pressure sensors signal amplitudes are very low compared to the amplitudes of the runner blade sensors. The basic frequency of the guide vane passing can only be detected at sensor positions 1 to 3 and position 6.

Figure 24: Numerical results – draft tube sensors

37

The pressure pulsations resulting from the runner rotation itself (Fig. 24 for the draft tube sensors in detail), the pit, the guide vane and the double guide vane influence could be detected for all sensors as an example in the CFD calculation. Additionally, the runner blade and double runner blade could be detected in the draft tube. All the CFD signals were normalised with the maximum amplitude of all sensors during the last 10 revolutions of 10008 Pa. For sensor 3 on the pressure side (PS3) of the runner blade, which is on a large diameter at the leading edge and a peak, it is remarkable that it could be detected in the experimental as well as in the CFD investigation for the guide vane and double guide vane frequency.

3.3.6 Results velocity measurements The velocity profiles were measured at different operation points and different measuring positions. Fig. 9 shows the velocity profiles at the measuring positions 1, 2 and 3 (in front of the guide vane) at the best efficiency point. The absolute values of the velocity at positions 2 and 3 are almost constant over the displacement. The velocity at position 1 slightly differs from position 2 and position 3 – this is caused by the turbine-pit influence (see also small picture in Fig. 20). The flow direction for position 1 and 3 changes over the displacement, which is a result of the crossing of the displacement line and the leading edge of the guide vane in flow direction. The displacement line of position 2 is located between the guide vanes, and the angle does not change along the displacement. Fig. 25 also shows the velocity profiles at the measuring positions 4, 5 and 6 (between the guide vane blades and the runner blades). The velocity profiles of the different positions are equal, which implies homogenous flow conditions behind the guide vane blades and no deviation along the blade channel width (between the two guide vanes). Only a small deviation could be found at 33% channel height. Finally, Fig 20. shows the velocity profiles at the measuring positions 7, 8 and 9 (draft tube inlet) for the same operation point. The absolute velocity is almost constant over the displacement and only a very small amount of residual swirl behind the runner could be detected.

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Figure 25: Velocity measurement

For the numerical simulation the monitor points were set during the pre-processing phase at the same locations where the experimental investigation was carried out. As far as the transient simulations is concerned, the last 5.000 time steps – of more than 12000 time steps for one simulation – were averaged (0.3 degree revolution per time step). 10 monitor points were set for each section, therefore the region close to the wall was roughly captured. Measurements within the experimental setup closer to the wall or into the middle of the draft tube were not possible due to the limited length of the very thin 5-hole probe.

3.3.7 Summary To investigate the hydraulic system in detail, pressure pulsations and cavitation behaviour were investigated and velocity field measurements were realised. The pressure pulsations were within the expected range and a comparison could be realised between numerical simulation and experiment. In the case of developed cavitation a satisfactory compliance of the experimentally determined and numerically calculated results was achieved

39

3.4 An analysis of the impact of draft tube modifications on the performance of a Kaplan turbine by means of Cumputational Fluid Dynamics

Due to the low electricity prices in central Europe cost optimizations related to all parts of a new hydro power plant have become increasingly important. Thus, the layout of the power plant has to be simplified and the dimensions of the single components of the turbine type have to be reasonably reduced. In case of a run-of-river hydropower plant using a vertical axis Kaplan turbine one of the cost drivers are the excavation works. Thus, a decisive factor for the reduction of construction costs is the minimisation of the construction depth of the elbow-type draft tube.

This background information was the starting point of a study dealing with a new run-of-river hydropower plant in Austria. The power plant will be equipped with two vertical axis Kaplan turbines with a runner diameter of D = 3.15 m. The net turbine head accounts for Hnet = 9.00 m and the maximum discharge per turbine is Qmax = 57.5 m³/s. Figure 20 gives an impression of the general layout of the power plant and its location in the river.

In course of the planning

phase it turned out that the

original turbine concept with a

model tested vertical Kaplan

turbine does not meet the

initially planned construction

cost target. Figure 26 shows a

sectional view of the original

Kaplan turbine configuration

with some basic dimensions

related to the construction

depth. The presented

dimensions show that the

excavation depth at the

bottom of the elbow T as well as at the outlet of the draft tube (see ∆z1 and ∆z2)

had to be reduced by significantly more than 1 m.

Figure 26: General layout of the power plant

40

One of the proposed cost saving opportunities was to change the draft tube design in order to reduce the excavation costs. However, it is well known that the construction depth of an elbow-type draft tube may have a big impact on the turbine performance. To ensure that the intended design modifications of the draft tube do not have a negative impact on the overall turbine performance Computational Fluid Dynamics (CFD) was applied.

After it was proven that there is a good agreement of the numerically calculated efficiency and the model test efficiency for the original turbine configuration various alternative draft tube designs were investigated in order to find out their impact on the turbine performance. Finally it was possible to find a new draft tube design representing a compromise of reduced construction costs and acceptable turbine efficiency.

An overview of the three different basic draft tube designs analysed in course of a preliminary study is given with Fig. 28. The draft tube geometry V1 refers to the original turbine configuration which underwent a model test. Geometry V2 represents the draft tube of an existing Kaplan turbine which was built around the year 1980. The specific speed of this turbine is quite close to the one of the turbine investigated in course of the present study. It was chosen because of the fact that the construction depth agrees well with the dimensions presented in Fig. 27. Finally, geometry V3 was created by using an in-house-developed design tool for elbow type draft tubes.

Figure 27: Sectional view of the Kaplan turbine with some basic dimensions related to the construction depth

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Figure 28: Superposition of the draft tube geometries V1, V2 and V3

In order to validate the different draft tube designs the turbine efficiency curve was calculated within an operational range of Q = 27 - 63 m³/s by applying numerical simulation. A comparison of the resulting efficiency curves is depicted in Fig. 29.

Figure 29: Comparison of the efficiency curves calculated with the draft tube designs V1 to V3 and V1-mod

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Due to the outstanding efficiency reached with the original draft tube design V1 it was chosen for all further considerations. The plan was to keep the basic idea behind the original draft tube design and to modify the existing geometry by scaling and cutting of parts like the entrance cone, the elbow and the discharge part in a way to reach the dimensional specifications given in Fig 26. After passing a few optimization loops the final draft tube version V1-mod (see Fig. 30) was created. The efficiency curve calculated by using the new draft tube geometry V1-mod is presented in Fig. 28 as well. Although the excavation depth was significantly reduced it turns out that the efficiency curve is only slightly below the reference curve and thus met the expectations of the future operator of the power plant in the end.

Figure 30: Comparison of the initial and modified draft tube

geometry in side view

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3.5 Benchmark study of automotive oil pumps by means of experimental investigations

In the course of a benchmark study 4 different types of automotive oil pumps were investigated on a custom designed test rig. For each of the pumps a measurement of the performance hill chart was carried out at fixed temperature values of T = 30°C, T = 60°C and T = 80°C. The test rig which was specially designed for this purpose allows an adjustment of a differential pressure in the

range of ∆p = 1 – 5 bar. By a speed variation between n = 250 rpm to n = 4000 rpm a flow rate of up to Q = 40 l/min could be reached. A schematic drawing of the test rig is depicted in Figure 31. A photographic documentation of the assembled test rig with one of the investigated pumps is shown in Figure 32.

The investigated pumps differ from each other in many aspects – they were produced by different manufacturers and work with different pumping principles. While two of the pumps work as conventional vane pump, one pump works as external gear pump and one as a pendulum pump. A basic overview of the investigated oil pumps is given with Figure 33. However, what all of the pumps have in common is the possibility of a displacement control via a regulation of the pump stroke. In the presented study the stroke adjustment mechanism was deactivated and only the full stroke and half stroke state was investigated.

Figure 31: Schematic drawing of the custom designed test rig

M

Temperature T

Torque M,

Speed n

Differential pressure ∆p

Pressure pulsations pS Pressure pulsations pD

Connection for secondary

cooling circuit

Pressure control valve

Pressure

relief valve

Filling

Manual deaeration

(with additional opportunity for pressurization)

Collecting basin for

external leakage

Oil tank

44

Figure 32: Test rig with one of the investigated pumps

Pendulum pump (Manufacturer A)

Vane pump (Manufacturer A)

External gear pump (Manufacturer B)

Vane pump (Manufacturer C)

Figure 33: Overview of the investigated pumps

45

At the example of the pendulum pump with fixed half stroke the efficiency hill chart measured at a constant temperature of T = 80°C is depicted in Figure 34. The corresponding pump characteristics measured at variable speed are shown in Figure 35. In this case a total pump efficiency of around 66 % was reached at

a differential pressure of ∆p = 4.25 bar and at a discharge of Q = 15 l/min.

Figure 34: Efficiency hillchart of the pendulum pump measured with half stroke at T = 80°C

With all 4 oil pumps mounted on the test rig the pump characteristics were measured for up to 3 different fixed temperature values and for the fixed full stroke and half stroke state. A summary of the measurement results is presented in Figure 36. A comparison of the results shows that the maximum pump efficiency ranges between 32 % and 74 %. While the highest efficiency level was reached with the pendulum pump, the lowest efficiency level was found in course of the measurement of the external gear pump. Additionally it turned out that the full stroke operation is more efficient than the half stroke operation. Furthermore it was found that an increased oil temperature results in a shift of the best efficiency point to a lower pressure level and a higher discharge.

46

Figure 35. Corresponding pump characteristics at const. speed measured with half stroke at T = 80°C

Figure 36: Summary of measurement result

47

3.6 Research on mechanical seals with a custom made endurance test system

In general, mechanical seals can be used for any kind of pump. Due their costs, they are mainly used in more demanding situations – especially in powerplants. So most of the cooling pumps, condensate pumps or boiler feed pumps are equipped with mechanical seals. Boiler feed pumps in particular are the most critical application for mechanical seals, as beside the high pressure and temperature, the used medium is deionised water (DI). It is vital, that the conductivity of the water is below 0,2 µS/cm to ensure a protecting surface on the used steel parts to prevent corrosion.

Figure 37: Gliding faces of the SiC paring at the begin, in the middle and at the end of the experiment

48

Nowadays, most of the mechanical seals use SiC as the main sliding material. This works quite well in most applications, but for demanding situations like boiler feed pumps, those seals are damaged within a few 100 hours of operation as shown in Figure 37. Therefore it is vital to find materials which are capable of longer lifetimes. For this purpose the sliding materials can be coated with a very thin layer of artificial diamonds.

Conditions at Test Rig

Common Conditions in Powerplants

Pressure [bar] 35 15-30

Sliding Velocity [m/s] 60 40-60

Water Temperature [°C] 60 ≤ 60

Conductivity VE-water [µS/cm] < 0,1 < 2

Figure 38: Conditions for the mechanical seal at the test rig and in real operation

This research project proves the higher lifetime of different diamond layer heights. For that, different material pairings were tested for more than 13 months. The results of one material pairing are shown in Figure 39. The overall uptime of the test rig was higher than 81%, which means a constant load for more than 7700 hours, although there was a damage of the main test rig pump which forced a downtime of almost 20 days. During the monthly inspections, the surfaces were tested in an electron microscope to get a visual impression of the surfaces.

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Figure 39: Chronological sequence of the 14 months lasting test

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3.7 Research on high head process pumps – simulation and optimisation of side channel pumps

Providing high pressure at relatively low flow rates, side channel pumps are often used when choosing between a displacement pump and a centrifugal pump. Beside the main advantages of an outstanding suction performance and the capability of pumping fluids with high gas loads, the drawback of pumps with such low specific speeds is a quite poor efficiency. In the past, when efficiency was often negligible, this fact was simply accepted. Nowadays, companies are pushed by law to increase the efficiency of their products, e.g. by the Energy Efficiency Directive of the European Union or the Paris Agreement. As pumps almost require 20% of the world electric power consumption, they offer a huge potential of saving energy. Though many manufacturers have recognized this trend and improved their main models, some exotics like side channel pumps were neglected in the past. This research project includes a detailed analysis of a typical 1-stage side channel pump with an additional radial suction impeller by means of CFD simulations to find the main losses and show optimization potential. Beside the challenge of increasing the efficiency, it is vital to mention, that the overall dimensions, the head curve and the suction behaviour have to remain identical to guarantee compatibility of the pump as spare part. In a first step, the model was successively generated and it was obvious, already in a very early stage that it has to contain all details including like suction impeller (360° model), main stage (360° model) as well as the pressure housing and all narrow gaps – see Figure 35 – to provide useful simulation results. Numerical simulations were carried out with the commercial CFD package ANSYS CFX 17.1 in stationary and transient way with scale resolving turbulence models. Mainly structured grids for the impeller, side channel, suction piece and gaps were used. The final model of the existing pump consists of 15 mio. nodes.

51

Figure 40: CFD-model of the analysed original side channel pump with radial suction runner (yellow) and impeller (blue)

The CFD-simulations were validated with model tests, and the behaviour of the main components (in stream wise direction) was analysed in detail. Beside loss analyses of the mentioned components, particular attention was paid to the behaviour of the fluid in the inlet and outlet port and the pressure generation in the side channel – see Figure 41. For the optimization process it was necessary to generate a reduced numerical model to analyse the effects of more than 300 geometry variations. After the effects of single and multiple geometry variations were analysed, the findings were combined to establish the desired objectives and validated again with the full numerical model. The numerical simulations predicted a relative efficiency increase at BEP and part load >30% with respect to all given limitations like identical head curve, suction behaviour and dimensions. Finally, those numerically predicted numbers were validated as the optimized model was manufactured by rapid prototyping and tested on the test rig in the laboratory. The excellent accordance between CFD and measurement is shown in Figure 42.

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Figure 41: Development of head from inlet port via side channel to outlet port with distribution in each blade channel

Figure 42: Comparison of original and optimized pump

180 195 210 225 240 255 2700 45 90 135 180 225 270 315 360

Angle Angle

Inle

t p

ort

Sid

e c

ha

nn

el

Ou

tle

t p

ort

He

ad

He

ad

0%

5%

10%

15%

20%

25%

30%

35%

40%

45%

50%

0

5

10

15

20

25

30

35

40

45

50

0 2 4 6 8

Pu

mp

he

ad

Pu

mp

effic

ien

cy

OPT trans. - Head

OPT - measurement - Head

ORIGINAL trans. - Head

ORIGINAL - measurement - Head

OPT trans. - eff

OPT - measurement - eff

ORIGINAL trans. -eff

ORIGINAL - measurement - eff

Pump discharge

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3.8 Research on canned motor pumps – simulation, optimisation and experimental validation

As the investigated pump is designed for high head and small flowrates, like many process pumps, the specific speed is relatively low and therefore the hydraulic efficiency of the pump is rather moderate due to physical limitations. For an outstanding safety of a hermetic pump concept like a canned motor or a magnetic coupling, a higher energy consumption of the drive is required. These two effects were simply accepted in the past. Nowadays, companies are pushed by law to increase the efficiency of their products like the Energy Efficiency Directive of the European Union or the Paris Agreement. As pumps almost require 20% of the world electric power consumption, they offer a huge potential of saving energy. This research project includes the detailed analysis of a multi-stage canned motor pump with an additional radial suction impeller by means of CFD simulations to find the main losses and show optimisation potential of the hydraulic parts. In a first step, the model was successively generated and it was obvious, already in a very early stage that it has to contain all details including impellers, guide vanes, return vanes, gaps and pressure relief holes as well as the hydraulically coupled rotor of the drive – shown in Figure 6 – to provide useful simulation results. Numerical simulations were carried out with the commercial CFD package ANSYS CFX 17.1 in stationary and transient way with scale resolving turbulence models. Mainly structured grids for the impeller, gaps and pressure relief holes were used with a final model of the existing pump consisting of more than 35 mio. nodes. The behaviour of the main components was analysed in detail and additionally the CFD-simulations were validated with model tests.

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Figure 43: Cross section of a 3-stage canned motor pump with modelled components and flow paths

Figure 44: Different loss analyses for the investigated original 3-stage canned motor pump

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Based on the loss analysis of the existing pump as shown in Figure 44, a 2-stage optimisation procedure was performed by combining manual engineering with automated optimisation of a reduced model of the fully parameterised return channel. After the effects of more than 300 geometry variations were analysed, the findings were combined to establish the desired objectives and validated again with the full numerical model. The numerical simulations predict a huge efficiency increase at BEP and part load with respect to all given limitations like identical head curve, suction behaviour and dimensions.

Finally, the optimised model was manufactured by rapid prototyping and tested in the laboratory. The results in Figure 45 show a satisfying correlation between the numerically predicted and the measured results on the test rig and prove the high quality of the executed numerical simulations.

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3.9 Research on valves – combining measurement and simulation to get essential information on valve behavior

Valves and shut-off devices are system-relevant and important for various industries. Depending on the risk potential, the verification of valve-typical parameters such as flow coefficient (cv, kv), adjustment torques, forces or closing times is of importance. For some industries (heavy process engineering, refineries, automotive), 1:1 experimental investigations are necessary, for other industries, experiments on a scaled model or CFD calculations are suitable (hydro, power plant technology). Some sectors (process engineering, water supply) work with enormous safety margins, other ones try to comply with the specifications as accurately as possible and minimise component weight and space as far as possible (automotive). All of these examples have one thing in common: They require either a test environment that meets the specific requirements, standards and provide flexibility in a tailor-made manner or they are obtained with numerical methods. However, relying on simulation results implies some difficulties: First, the trustworthiness and accuracy of the simulation. There are many CFD programs available and every program allows for completely different modelling types (turbulence model, mesh type, stationary or transient analysis, etc.). Second, every simulation is only as reliable as the available data is. So, if there are deviations in the geometry, e.g. the casted body, the simulation will never be able to predict the characteristic values correctly. Third, for characteristic values like cv calculation is much easier than for cavitation or acoustic data. Comparing the effort, time and cost required, CFD is usually superior to measurement. But when it comes to accuracy, measurements are still unbeaten – given that the measurement equipment is appropriate. Combining those two approaches is the key factor for successful development and analysis. The following examples, all carried out at the Institute of Hydraulic Fluid Machinery in recent years, show the combination of both ways. This helps in “calibrating” the numerical simulations for a more precise calculation in the future to fasten the development and analyse process on one hand and to minimize the test effort on the other hand.

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� Expansion valve (automotive)

In automotive industry, research on new products requires a high reliability

already in an early development stage. Therefore numerical simulations are

often the method of choice. In this presented topic, an expansion valve for a

cooling circuit had to be developed. Besides providing a certain flow

characteristic within a limiting range – which is already a challenging task – it is

vital to mention that an expansion valve is operating in a 2-phase regime.

Depending on opening and pressure, the distribution between liquid and vapor

varies as shown in Figure 46.

In Figure 47, the results of the numerical simulation in comparison to the

specified values are presented and they show a quite good correlation within the

complete range.

Figure 46: log p-h diagram for coolant and investigated cycle

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Figure 47: Comparison of measurement and CFD for investigated expansion valve

� Control valves (process industry)

Designing control valves for certain applications in the process industry nowadays relies heavily on experience. In a research project with one of the biggest valve manufacturers, the Institute for Hydraulic Fluidmachinery made measurements and CFD simulations (see Figure 48) to investigate the occurring flow phenomena. With this knowledge it is now possible to reliably design control valves, allowing for much lower tolerances than some standards require.

� Butterfly valve (hydropower Häusling) The original flap in model size is examined on a test bench and subjected to a comprehensive test program, whereby measurements are made under different cavitation conditions. As those tests are quite complex it was interesting to find out whether the flow coefficient can also be determined under heavy cavitation with means of CFD.

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Figure 48: Tested control valve and comparison of measured (blue) and CFD data

As the results in Figure 49 show, it is possible to simulate the real behavior under heavy cavitation with a sufficiently high accuracy which is of high importance during the design process of certain valves.

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3.10 Cooling jacket for electric mobility

Based on prototype measurements, a research project between an international automotive supplier and the Institute for Hydraulic Fluidmachinery was initiated to improve the accuracy and reliability of the numerically driven design process for cooling jackets for electric drives. As time is a decisive competitive advantage, it is vital to deliver “better” results within a shorter period. In this research project, the focus was mainly on the occurring pressure losses and the thermal behaviour of various cooling jacket geometries. In order to verify the influence of the quality of the computational mesh on the results, detailed studies were carried out. Coarse, unstructured grids with low boundary layer resolution were started and the quality was successively increased. In addition to an increased accuracy through finer resolution of the unstructured grids, the superiority of structured meshes as shown in Figure 50 could also be demonstrated.

Figure 50: Numerical mesh of the flow channels inside the cooling jacket

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Geometries of cooling jackets are often dominated by the manufacturing process and the available space inside engine compartment. Though this allows mostly not for the optimum performance, engineers have to make the best out of it. In a first step, the influence of the manufacturing process was investigated, which affects mainly the gap height between the cooling jacket and the housing and in a further way also the surface roughness of the processed parts. In Figure 51, the pressure drop between inlet and outlet is shown and it is visible that an increased gap leads to a lower pressure drop. This means that most of the cooling fluid bypasses the narrow channels and goes directly from inlet to outlet which leads to an inferior cooling performance. In addition to the very precise calculation of the pressure loss, the numerical simulation also enables the finding of hot spots for the reliable placement of power electronics and critical components, which under no circumstances may overheat.

Figure 51: Dependency of pressure drop on gap width between cooling jacket and housing

On the basis of simulation results, an optimization of the entire cooling jacket can take place, whereby reliable results are already available at a very early development time. The accuracy of the calculations generally results in a

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shortened development time because inappropriate geometries can be identified quickly and unnecessary iteration loops can be avoided. After thorough analysis of a geometry using stationary simulations, transient calculations from load changes to combined load cycles are possible.

Figure 52: Temperature distribution of a cooling jacket – simulation results

Furthermore, the results provide an important contribution to the calculation of the entire powertrain or subsequently of the vehicle through the use of 1D programs for system analysis. The current state of development already provides important information for the future design of heat sinks. With the gained knowledge about the pressure loss calculation and the thermal analysis by coupled fluid-solid-simulations the development process can be made significantly more efficient and effective.

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3.11 Research on energy optimisation in passengers cars – simulation of a gear box oil pump

To reduce CO2 emissions and fuel consumption, some transmission manufacturers take advantage of dry sump lubrication. So a pump is designed to lubricate the wheelsets and bearings in the gearbox. Preferably, it pumps the medium to a higher level into a container inside the gearbox. From there it passes through outlet openings to reduce the friction between the flanks of the gears and cools them. Through this dry sump principle drag and hydraulic tow losses are prevented, which in conventional dive lubrication occur, and this ultimately leads to reduced fuel consumption. However, a cost and efficiency optimum is only achieved if the pump is designed for the mandatory operating points which are achieved by optimizing the hydraulic shape on one hand and on the other hand by providing a speed controlled drive.

Figure 53: Hydraulically optimized components

The main challenges in the design of this pump were the highly fluctuating viscosity of the oil in the relevant temperature range and the limited assembly space. As the pump should be as simple as possible, inflow, volute and outflow

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are combined in one single part (shown in Figure 53) which required a lot of proper expertise. Due to the installation on the bottom of the gearbox the inflow region is slightly submerged which helps on one hand to ensure a stable suction behaviour but on the other hand air sucking vortices have to be eliminated. The design of the impeller blades, which is the main task when a pump is hydraulically developed, had to be adjusted to the variable speed of the drive. Therefore an optimization task in a very early phase was necessary to define the design (unshrouded impeller) and the dimensions (hub and shroud diameter) of the pump. In a further step all gaps between stationary and rotating parts were analysed and optimized to ensure a maximum electric efficiency of the drive and to keep the friction losses as small as possible. The CFD results were finally validated in model tests and have shown a good correlation – both hydraulic performance and friction losses.

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3.12 Research on safe operation of hydraulic systems by means of sophisticated transient analysis

Transient flow phenomena will occur in every hydraulic system and must already be taken into account in the layout and design process. If the resulting loads from unsteady fluid dynamics are not known, damage to the system and the associated hazards and downtimes are often unavoidable.

Figure 54: Numerical model of a cooling system for a process plant

In the planning phase of hydraulic systems (including redesign as well as adaptation / upgrading of process plants, cooling systems, waste water or drinking water systems) their components are usually designed or dimensioned only for stationary operation. Unsteady flow phenomena are often mistakenly neglected since their impact is not known. But transient flow processes can lead to significantly higher structural loads on the system than those in stationary operation. Already a simple start-up or shutdown of a hydraulic system induces transient flow processes. However, exceptional or accidental load cases may not be eliminated in the operation of such a system. Therefore those events (e.g. a sudden pump trip due to an electrical defect or the closing of a control valve)

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Figure 55: Simulation results pump trip– above: initial system layout – below: secured system layout must be subjected to a precise investigation by means transient simulation or waterhammer analysis. With sophisticated numerical models it is possible to accurately predict the resulting loads like positive and negative pressure and / or cavitation). If those loads are found to be unacceptable any precautionary measures planned in advance and corresponding further steps can be taken.

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Basically, pressure surges increase with increasing length of unsecured pipeline and with increasing the speed of changing the flow rate.

In the course of the research project not only the behaviour of different counter measures is investigated. Also the correct position in the system and its design and size is optimized by means of numerical methods. A comparison of the numerical results between an unsecured initial system layout and an optimized system is shown above. As one can easily identify not only the pressure surge can be decreased tremendously by applying an optimized waterhammer protection. Additionally the calculated pressure transients in case of a sudden pump trip are comparatively smooth for the secured system layout. Hence lower pressure loads are to be expected for the optimized system yielding in benefits in endurance of all pressurized components of a plant.

Figure 56: Comparison of measurement and simulation

In order to confirm the simulation results and to verify any necessary assumptions in the course of the numerical modelling, field measurements must always be carried out during commissioning. It must always be ensured that the plant performance and thus the plant load are gradually increased in terms of

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these tests. In order to accurately measure the occurring pressure transients, high-resolution pressure transducers are required. Additionally the pressure measuring taps as well as their location must fulfil particular requirements depending on the system to be investigated to ensure a high accuracy and reliable measurement results. At the Institute of Hydraulic Fluid Machinery both, the measurement equipment and especially the particular know-how to gain high quality measurement results is available. The comparison of measurement data and simulation result shows that the rapid pressure drop during pump trip cannot be detected precisely with the low res pressure signal of the control system. Additionally the minimum negative pressure of the control system measurement is not reproduced exactly. However, the measurement results of the high-resolution pressure sensors almost exactly match the simulation results in this area. Hence the quality and reliability of the numerical results could be confirmed. Transient flow phenomena lead to very high plant loads which may exceed the design pressure if these are not known. A comparison between simulation and system measurement shows how precisely these loads can be predicted with numerical methods. The prerequisite for this, however, is a conscientious procedure in data processing as well as the availability of corresponding data or information. If these analyses are already carried out in the planning phase, appropriate remedial measures can usually be implemented at relatively low cost. Adaptations during operation of a plant are usually associated with high time and effort at correspondingly high costs (also resulting from standstill). In addition, subsequent counter measures are usually limited in their application.

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3.13 Multiphase Simulation of the hydraulic behaviour of a weir

Weirs are often used in run-off river hydro power plants to guarantee a save flood water flow. In history the weir performance for a certain hydraulic geometry was usually measured in terms of a model test in a hydraulic laboratory. In the course of the present research project the weir performance was investigated only by means of 3D-CFD Simulations. This method presents a novel approach in analysing the hydraulic behaviour of civil works. By means of multiphase simulations not only the hydraulic characteristics could be investigated. Also the mechanical loads on the steel works could be analysed. This information is crucial to design the drive mechanism in order to operate the weir also in the case of flood water. Additionally the tailwater flow field especially in the stilling basin as well as the aeration behaviour could be analysed.

Figure 57: Left - CAD model of the weir situated at the power plant; right – CFD model

At a first step the geometry of the radial gate combined with a flap had to be prepared for the numerical simulation and a mesh was generated. At the beginning a quasi 2D approach was chosen for numerical modelling of the real problem. The results of these simulations in terms of the hydraulic behaviour by means of flow rate versus liquid level were then compared to available literature data. Additionally transient simulations were carried out in order to investigate possible vibration excitations on the mechanical structure triggered by the fluid flow. Therefore a very fine numerical mesh had to be created in order to resolve the vortex shedding at the trailing edge of the flap.

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Figure 58: Flow field in a symmetry cross section from different simulations

The final simulations were carried out on a full 3D model capturing constructive details like deflectors and also gaps between steelworks and civil works. These analyses were conducted for several positions for both the radial gate as well as the flap for different liquid levels in the headwater.

Figure 59: Numerical results for different positions

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3.14 Increasing power output and flexibility of existing high head power plants with help of waterhammer simulations

In a changing energy market with decreasing and non-predictable energy prices, energy suppliers try to sustain their financial return with their existing hydro power plants (HPP). On the other hand, grid service tasks like quick reacting power generation are needed to compensate the stochastic generation from wind and solar. A chance to increase the annual turnover is found not only in increasing the power output but also the availability and flexibility of an existing power plant.

Figure 60: Examples of complex surge tank designs

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As many plants operated today were constructed decades ago, the original plant design was based on operating conditions that were completely different to the ones actually required. To prove the ability for higher flexibility in power generation, detailed investigations of the waterway and especially the surge tank limitations are necessary. Since the main constructive infrastructure can hardly be modified, the possibilities of additional power generation are mainly limited to an increase of flow rate. But the higher the demand for an increasing power output, the higher are the limitations in operational flexibility. An overflow of the surge tank must be prevented in any possible operational or exceptional scenario of the power plant. On the other hand, a ventilation of the headrace system through an empty surge tank must not occur during operation. Additionally the maximum penstock pressure must not exceed the already existing limits.

Figure 61: Comparison of simulation results and field measurements

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Flexible operation induces a highly transient fluid flow in the head race and

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refurbishment to increase the power output and/or the flexibility, a reliable

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Modern techniques in transient system simulation provide the only feasible possibility to accurately calculate the occurring waterhammer and mass oscillations. Since most alpine high head power plants have a unique surge tank concept their physical behaviour cannot be simulated accurately with standard models supplied by commercial software-packages. Thus detailed and exact tailor made numerical models of complex prototype power plant components like turbines, emergency closing valves and especially surge tanks are necessary. To validate the results of the tailor made numerical models a comparison of the numerical result with on-site measurements was carried out at the beginning of the numerical simulations. In that course a validation of the numerical model, especially for the surge tank behaviour, as well as the losses in the power water way, was possible. With the help of transient 1-D numerical simulations and a validation with on-site measurements surge tank oscillations with increased power output can be calculated precisely. It could be shown that a given hydraulic infrastructure is able to allow for greater flexibility and higher power with only minor additional limits in the permissible operating range.

Figure 62: Numerical results for different surge tank models

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3.15 Improvement of intake structures with numerical simulation

The design of inlet and outlet areas of hydro power plants significantly influences the overall cost of a plant not only in the design and construction phases but also during operation. Due to an incorrectly shaped inlet geometry, vortices and flow separation occurred at several river power plants. These phenomena can cause flow problems and losses at the turbines and must not be neglected, especially for low head turbines with relatively high specific speed turbines. Sources of such inhomogeneous inlet flow distribution could be a wrong design of dividing piers, gravel steps and other points of discontinuity in the flow guiding walls and river bed. In the course of this research project, samples of water way designs and their optimisations, whose target is to avoid wrong inflow and outflow designs, are analysed together with cost-optimised design. One example is the new MKF hydro power plant. The weir turbine is located at the intake structure of an existing headrace channel for a diversion power plant. The new MKF power plant is located on the inner curvature of the river bend opposite the headrace channel for the existing power plant as shown below.

Figure 63: Numerical model, grid and evaluation planes for the presented research project

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The turbine’s direction of rotation must be investigated as it has an enormous influence on the civil engineering construction cost due to power transmission lines located in the air space above the excavation zone. The investigation was based on qualitative flow evaluations at different planes. The evaluation levels 1, 2 and 3 each represent horizontal sections through the flow field at different height levels. Additional evaluation criteria represent the ratios of maximum and minimum occurring flow velocity to average flow velocity for the evaluation of a uniform flow towards the turbine. The comparison for the left- and right-hand turbine configuration is shown for a specified load case.

Figure 64: Streamlines for left-hand (top) and right-hand (bottom) turbine configuration

It clearly shows that the right-hand variant for evaluation levels 1, 2 and 3 has a more inhomogeneous flow distribution in the MKF bay compared to the left-hand turbine. This statement is also confirmed by the evaluation of the maximum and minimum relative flow velocity at the spiral inlet – there are increased secondary flows at the spiral inlet when using a right-hand machine. Based on these test results, the installation of left-hand machines is recommended, the construction of which is also cheaper.

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The simulations with the initial intake geometry showed a non-optimal orientation of the supporting pillars in the inflow region of the power plant inlet for the left-hand rotating turbine. All tested load cases showed that the pillars should be rotated by up to 15° counter clockwise. In addition, the analysis of the flow field in the horizontal evaluation levels showed optimisation potential in the hydraulic design of the existing wall between the MKF bay and the right bottom outlet. The flow situation at the water surface shows that a dead water zone is to be expected in the inlet area in the bay power plant. An optimisation of the hydraulic geometry (enlargement of the radius) of the right orographic boundary in the inlet area of the MKF bay has been proposed in order to avoid this unfavourable flow situation together with an optimized shape of the gravel step. The hydraulic contour of the existing wall between MKF bay and right to bottom outlet was modified to achieve the desired effect of no detectable level reduction. The water level with the optimized geometry is now constant along this wall as visualised in the figure below, whereby the optimised contour is shown at the bottom left.

Figure 65: Water level at wall, distribution of intake water

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3.16 Optimisation of Pelton distributors with the help of 3D CFD

In the course of the present research project a numerical optimization of a distribution pipe of a Pelton-turbine is carried out. It is based on an existing distribution pipe of a vertical Pelton-turbine with six nozzles. In a first step the flow field in the distributer of the existing configuration is analysed and possible losses and their origin are examined. Additionally the generated jet is analysed in qualitative manner. Thus an influence of secondary flows on the jet shape and the jet / bucket interaction subsequently can be identified.

Figure 66: Left - CAD model of the Pelton distributor; right – numerical mesh

In the course of the optimisation the focus was on the implementation of standardized components and fittings to minimize the production costs of the final solution. The numerical analysis of the hydraulic design of the reference geometry revealed several potential of optimization. It turned out that the shape of the jet is strongly influenced by shape of the junctions. In addition at the locations of these junctions a source of losses was found, too.

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Figure 67: Results for the reference geometry; left – losses, right – velocity distribution

The optimized design is a development concerning standardized components serving in multiple constructive functions to minimize the costs. The optimization of the hydraulic shape is based on the results of the numerical simulation of the reference geometry. Finally, an optimized design could be investigated by implementing a research strategy not only focusing on an optimal hydraulic design but also on production costs.

Figure 68: Comparison of streamlines and jet - reference geometry (left) and the optimized

shape (right)

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3.17 Investigation of hydraulic losses at the connection of a surge tank and penstock for different load cases with the help of CFD

With highly volatile loads in electric grids from renewables like wind and solar grid service tasks become more and more important. This offers a new possibility for existing storage power plants to generate additional revenue. To address the demands of the electrical grids rapidly changing and fluctuating power output has to be guaranteed. But fast and multiple switching of power means rapidly changing flow rates which will cause sever hydraulic transients in the power plant. To prove the ability for higher flexibility in power generation, detailed investigations of the waterway, especially surge tank limitations, are necessary. Especially but not only in the surge tanks of pumped storage power plants complex flow regimes and load cases at the surge tank base point are possible.

Figure 69: Numerical mesh of the surge tank base point

With different states of flow by means of diverging or combining flow at the surge

tank base point also the behaviour of hydraulic losses at this specific location

may change. Since these base point losses account for the damping effect of the

mass oscillations in the headrace system their exact knowledge is of major

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importance. In the course of the presented research project these losses were

calculated with the help of numerical methods for every possible load case.

Figure 70: Investigated operating modes (excerpt)

The investigations showed a strong influence of the calculated loss coefficient at the surge tank base point as a function of the flow regime at that location. This is of major importance since usual approaches so far assumed a constant loss coefficient for every state of flow.

Figure 71: Streamline visualization at mode 3

Additional in-formation was gained by analysing the flow field in the different operation modes. For example the maximum velocity magnitudes in near wall regions or unfavourable load cases for the tunnel-lining can be detected.

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3.18 Research project Francis Turbine Virtual + As electricity market parameters are undergoing major changes, cost optimisation regarding the planning and realisation as well as the revitalisation of hydro power plants becomes mandatory – especially as far as turbine design, durability forecast and component quality management are concerned – aiming at an increase of the annual output and/or the return of investment.

At present, 100% reliability of statements on the efficiency, the annual output, cavitation and vibrations of Francis turbines can only be provided by means of model tests. However, model test are complex, time-consuming and expensive, and only pay off for turbine performances of approximately 10 MW and more. Due to limited project budgets, time and cost-consuming model test are merely realised in exceptional cases. Whenever planning of a plant is solely based on considerably less expensive CFD (Computational Fluid Dynamics) calculations, numerous assumptions are made that lead to differing data on the annual output, the influences on the durability and the behaviour of systems (vibrations, cavitation) and to non-reliable water hammer calculations. This strategy finally bears unrateable technical and economic risks. However, future developments and optimisations of Francis turbines will increasingly be based on CFD calculations (without model tests) in order to save costs.

Whereas prediction accuracy of CFD simulations for the optimum operation range of a turbine is excellent, calculations for part load and overload are still subject to numerous uncertainties as of occurring flow separations and cavitation. Their simulation setup has to meet highest requirements regarding model design and modelling of turbulences. In addition, usually within the scope of CFD calculations the runner side gaps are not taken into account thus allowing for a considerable reduction of the already extensive simulation efforts. However, this approach requires simplified analytic calculation tools, the application of which often leads to misinterpretations resulting in necessary optimisations.

Water hammer calculations are also only possible based on turbine data defined by means of 4-quadrant characteristics diagrams whereas CFD calculation of these – so far – provide merely very inaccurate figures. As of these uncertainties plant components have to be oversized, thus causing unjustified increases of construction costs.

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The target of this research project is the development of a “Virtual Test Rig” for Francis turbines. Based on experimental tests realised in parallel on a model test rig, BEST PRACTICE standard procedures for the reliable calculation of the performance and the characteristics of Francis turbines throughout the ENTIRE operating range (part load, optimum, overload) will be defined, which are not existent in this form yet. As a result, the accuracy of CFD calculations will be increased and thus water hammer calculation based on CFD will become possible. Additionally, reliable assessments of the potential of refurbishment projects as far as possible efficiency levels and the amelioration of part load behaviour by means of new hydraulic systems can be provided.

The subject of the investigations is a Francis turbine which was originally designed at the Institute of Hydraulic Fluidmachinery at Graz University of Technology. The model turbine is equipped with 20 guide vanes and 13 runner blades. At the runner outlet the diameter accounts for D = 260 mm. In order to carry out a model test according to IEC60193 [1] a turbine speed of n = 1180 rpm was chosen to reach a Re-Number of at least 4.2*106. The turbine speed was kept constant over the entire operational range investigated on the test rig. The best efficiency point was measured at a flow rate of QBEP = 0.230 m³/s and a head of HBEP = 13.1 m. The specific speed accounts for nq = 82.2 rpm.

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BEPt

qH

Qnn

3.18.1 The model test The experimental investigations were carried out on a closed-loop 4-quadrant test rig (see Fig. 72) which was originally designed according to the requirements defined by the IEC60193 standard. A cross section of the vertical axis model turbine is depicted in Fig. 73. As shown in the figure the turbine is equipped with an elbow-type draft tube. The draft tube cone was manufactured from clear Plexiglas® to enable an observation of the cavitation effects occurring in the runner and in the inlet of the draft tube.

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Figure 72: Test rig configuration

Figure 73. Cross section of the Francis turbine model

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The basic measurement results obtained by the experimental investigation are summarized in the hill chart depicted in Fig. 74. On the one hand, the chart

contains the typical ISO-lines representing the hydraulic turbine efficiency ηhydr. plotted against the flow rate Q and the head H. On the other hand dashed

parabola representing operation points with constant positions α of the guide vanes are shown. For the hill chart measurements the absolute pressure in the system was adjusted in a way to reach a turbine operation without any appearance of cavitation over the whole operational range. The friction losses occurring in the sealing and bearing system of the turbine were measured with no-load tests and the results were used for the correction of the measured mechanical power output of the turbine. However, the losses in the runner side spaces, which are comparably high in this case (see later on), are still included in the displayed turbine efficiency.

Figure 74: Hillchart of the model turbine measured at n = 1180 rpm (Re = 4.2*106)

The six operation points marked with a filled circle located at deep and medium

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part load and full load were picked for detailed investigations like cavitation measurements and the measurement of pressure pulsations in the draft tube cone. In the course of the cavitation measurements pictures were taken with a camera by using a stroboscope with an appropriate flash frequency. The photos

of the cavitation state at a head of H = 13.1 m and a σ-value of around 0.3 are presented in Fig. 75 for full load operation, Fig. 76 for medium part load and Fig. 77 for deep part load. Especially the picture of the cavitating part load vortex occurring at medium part load and the cavitating inter-blade vortices found at deep part load were compared to the results of the numerical simulation.

Figure 75: Straight cavitating full load rope at H = 13.1 m and GV37.5° ( ~ 100% load)

Figure 76: Helical cavitating part load vortex at H = 13.1 m

and GV22.5° ( ~ 60% load)

Figure 77: Cavitating inter-blade vortices at H = 13.1 m

and GV15° ( ~ 30% load)

3.18.2 Numerical simulation To perform the numerical simulation – including the generation of meshes, the creation of the simulation setup, the solving process and the evaluation of results – the commercial software package ANSYS-CFX V17.1 was used. CFX uses a finite volume based discretization scheme up to second-order accuracy and is commonly used among others for the simulation of hydraulic fluidmachinery. The meshes were assembled to a full 360° CFD-model using ANSYS-CFX-Pre V17.1. For the investigation of the model turbine single-phase and steady-state as well as unsteady simulations were performed at several operation points ranging from part load to full load.

For the majority of the performed simulations the runner side spaces were neglected which dramatically reduces the computational effort. However, to identify the losses occurring in this region the flow in the runner side spaces on hub and shroud side was modelled for numerical simulations performed at

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selected operations points (see Fig. 78). It was found that the implementation of the runner side spaces results in a significant reduction of the turbine efficiency, which is presented with the loss composition depicted in Fig. 79. According to the losses due to leakage and disc friction the efficiency reduction accounts for around 3.5 percentage points at part load, for around 1.75 percentage points at BEP and for around 1.40 percentage points at full load. Especially for the comparison of the results of measurements and CFD this finding is highly relevant.

Figure 78: CFD-model of the Francis turbine with implemented runner side spaces

If the losses occurring in the runner side spaces are subtracted from the CFD-results calculated with the steady state 360° CFD-model, the solid red curve presented in Fig. 80 is obtained. The dashed efficiency curve represents the results of the experimental investigation. At part load and close to the BEP it turns out that there is a good agreement of the calculated and measured efficiency. However, especially at high load operation there is a significant gap between the two curves. Thus, unsteady CFD-simulations were performed and the calculated results are presented with coloured diamonds. By using unsteady numerical simulations the flow patterns occurring in the turbine are captured more realistically. As a consequence there is a much better agreement of the results of measurement and simulation over the full operational range.

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Figure 79: Loss composition calculated by means of CFD at H = 13.1 m at 3 operation points

Figure 80: Comparison of the turbine efficiency measured on the test rig I

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By using unsteady numerical simulations for the investigation of the turbine performance at part load operation it is also possible to visualize the occurring cavitation phenomena in the runner and draft tube. Although the simulation was performed using a single-phase model, which does not take into account the phase-transition from liquid to vapour, it is possible to display the cavitating vortex rope at medium part load (see Fig. 81) and the inter-blade cavitation vortices at deep part load (see Fig. 82) by plotting ISO-surfaces with the lowest pressure values calculated within the flow field. It turns out that there is an excellent agreement of the cavitation zones calculated by means of CFD and detected in course of the experimental investigation (see Fig. 76 and Fig. 77).

With the objective to reach an efficiency curve shifted to higher discharge with slightly higher peak efficiency and improved cavitation behaviour a new runner as well as new guide vanes were designed. A comparison of the original an optimized runner design and the pressure distribution calculated at the best efficiency point is presented in Fig. 83 and Fig. 84.

At the end of 2018 the new Francis runner will replace the existing one on the test rig. With the new design an improvement of the efficiency, performance under cavitation and performance at medium and deep part load is expected.

Figure 81: CFD-visualisation of the cavitating vortex rope at medium part load operation (GV22.5°)

Figure 82: CFD-visualisation of the inter-blade cavitation vortices at deep part load operation (GV15.0°)

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Figure 83: Pressure distribution at BEP on original runner design

Figure 84: Pressure distribution at BEP on optimized runner design

Finally a water hammer calculation using the measured AND calculated turbine hill charts will be performed. Especially at rather small opening of the guide vanes and in the hill chart region close to the runaway curve there will be some deviations between the experimentally determined and numerically calculated results. At the example of an existing power plant the impact of these deviations on the resulting water hammer and runaway speed will be investigated in detail.

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3.19 Numerical investigation and imporvement of cross-flow-turbines (high head application)

As small hydro power becomes more and more important, the investigation of cross-flow turbines became more and more important. Numerical Investigations of Cross-Flow turbines are not state of the art yet, because of the challenging two phase calculations with air and water inside the turbine. Based on the development of a crossflow turbine in the past (numerical as well as model and prototype measurements), more investigations were performed on different appliances (different Q, H, n - combinations).

Figure 85: Different investigated cross-flow turbine projects

One focus of these applications are high head applications. Caused by higher waterjet velocities at high head applications, the flow pattern inside the turbine changes and therefore the efficiency of the turbine could be negatively influence. Furthermore, cavitation may occur and this could lead to problems for the runner blades. At a given turbine application (given head, runner speed and discharge) the influence of the runner blade shape was investigated and improved to prevent bad flow pattern and therefore a minor turbine efficiency. Beside the negative influence on the efficiency, the aim was to evaluate the chance of cavitation caused by the runner blades.

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Figure 86: Flow- and pressure fields of the CFD-calculation

The actual shape of the runner blades was investigated systematically, which means different blade parameters were changed independently and then the effect on the turbine characteristics were calculated. Some of the parameters of the investigations were:

� Inlet- and outlet-angle � Blade thickness � Leading- and trailing edge shape � Blade thickness and different thickness-distributions along the blade

length � Blade length (caused by different diameter ratio of the runner) � Number of blades

As the waterjet hits the runner blades of the turbine twice (entering and leaving the runner) the optimization of one turbine stage of course influences the second turbine stage. For this reason detailed analyses of the flow pattern (e.g. velocity-fields, pressure-fields,i) were performed. The following picture for example shows the torque of one single runner blade during one revolution of the runner. Out of the transferred torque, it is possible to calculate the accumulated torque and finally the torque distribution between the first and the second turbine stage (figure below).

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During the investigation, the maximum turbine efficiency could be significantly

increased by optimizing only the runner blade shape. The following diagram

shows the turbine efficiency of the original runner blade design (red) compared

to the final version of the improved runner blade design (blue). Especial at part

load the turbine efficiency of the new runner blade design is superior.

Figure 88: Proved turbine efficiency of the new turbine blade design

Figure 87: Torque analysis of a single blade during one revolution

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3.20 Test rig for guide vane seals Nowadays fluctuations at the energy market getting higher and higher, a lot of water power plants are now used for the FCR operation mode (Frequency Containment Reserve). This means that there are ten times more motion cycles on the guide vanes then the mFRR operation mode (manual Frequency Restoration Reserve). After the first renew seal at the guide vane shafts they began to leak within 4-5 years and this leads to the question which sealing system is the best to use in the environment with sand containing water. For the investigations, the sealing system with the original diameter of 200 mm was used and the original dry sliding bearings were combined to a test rig with two seals inside. To shorten the needed time the original motion cycle from one power plant was analysed, idealized and then shortened to a “time lapse”-mode for the operation of the test rig. The test rig guide-vane shaft is moved by a synchronous motor and a gearbox in between. The pressure inside the sealing chamber (in nature the space for the guide vane) is measure to ensure constant conditions. In addition, the required driving torque and the leakage rate are logged at each second. The first experiment started with clean water and the results showed that there is no leakage after 11 weeks respectively 8.3 years of operation. For the next step, some changes in the design of the test rig were done for the use of water containing sand (slurry). The sand was taken from the Danube for the best comparison between power plant and test rig. After 7 weeks, respectively 5 years of operation, the sealing began to leak. As a second result, the required driving torque started to rise very quickly at the time the first leakage occurred. That means the sand between the sealing lip and the guide vane shaft increase the friction torque up to 150 %. The investigations are still running with a numerous of new combinations of sealing systems to find the best that would last more than 4-5 years and is therefore a reliable solution for the use in water power plants.

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Figure 89: Test rig for guide vane seals, to use a mixture of water and sand for testing

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3.21 Experimental and numerical investigation of a single blade impeller pump

Based on the past projects of single-blade impeller pumps further investigations were performed. These further investigations for example includes different pump / impeller sizes as well as the difference between open and closed impeller versions. During these further investigations, two different sizes of single blade impeller pumps were investigated to understand the flow behaviour of those impeller types und further on to improve the runner blade shape to reach a maximum pump efficiency. All CFD-calculations were verified by experimental measurements. Figure beneath are showing the 3D-model and the test rig design for measuring the pump characteristics. To determine the hydraulic efficiency a torque measuring flange were placed between the pump and the electric motor. All measurements were performed regarding the DIN EN ISO 9906. Beside the stationary pressure measurements (determining the Head), high frequency pressure transmitters were used to measure the highly transient behaviour of the pressure at the pressure- and the suction side of the pump.

Figure 90: 3D-construction and pump test rig

Because of the unsteady flow field inside single blade impeller pumps and the complex rear and front side of the housings with cutting edges, two different calculation models were used. One simple model without the front and rear housing (reduced calculation time) and a full model of the pump were generated. The following figure shows the mesh of the closed impeller design with included

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front and rear side of the housing (full pump model). For all CFD-calculation ANSYS CFD v17.1 were used. Because of the unsteady behaviour of the pump, only transient calculations lead to realistic pump characteristics.

Figure 91: CFD-mesh of the full pump model with the closed impeller

Figure 92: Pressure pulsations along the front side channel

Beside calculating the pump characteristics, detailed flow analyses of the flow pattern were conducted. For example, the figure below shows the pressure along the front side housing of the closed impeller design. It can be shown that the

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pressure continuously decreases along the front side of the housing from the pressure side (index 1) to the suction side (index 15) of the pump. The small pulsations of pressure line 5 and 7 are based on small fins, which are placed on the hub side of the impeller for balancing the runner. It turned out, that the fins have a positive effect on the pump performance, although additional fins required addition driving-power. As mentioned before single blade impeller pumps are characterized by highly unsteady flow patterns inside the impeller. The figure below shows the vertex structures (Q-criteria) of the closed impeller and the opened impeller with the simplified pump model. It clearly points out, that the quantity of vertex structures in the closed impeller are much lower. This is one of the main reasons, why the closed impeller reaches a higher pump efficiency curve (identical blade design).

Figure 93: comparison of the vortex structure inside the impeller of the closed (left) and opened impeller (right)

Finally the impeller designs of each pump could be improved by new 3D-impeller blade designs. All CFD-calculations (with the full pump model) at the nominal operation range of the runner are showing excellent accordance with the test rig measurements.

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3.22 Investigation of a 2-nozzle Pelton turbine On the 4-quadrant test rig of the Institute a 2-nozzle Pelton turbine was investigated. These investigations include the overall performance measurements (hill chart) as well as the flow visualizations of the jet and the bucket-jet interaction. The model turbine of the test rig was build and manufactured by the company partner. The following pictures shows the turbine model at the test rig.

Figure 94: Pelton turbine model with the option to change the whole turbine housing design (taken with an action camera system)

One focus of the performance measurements was the determination of the best combination of different impellers, housing designs and nozzle shapes. Each impeller had a different shape of the bucket itself and therefore the flow pattern of each impeller differs. Respectively the outflow of the buckets and the following interaction between the housing and the impeller differs.

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Figure 95: Example of the turbine characteristics at different nozzle openings

To better understand the outflow and the interaction one focus point was to visualize the bucket-flow in detail. Therefore after the performance measurements, flow visualization with photos and videos were taken inside as well as outside of the turbine housing. Two different types of high speed cameras (MacroVis and Photron) and a phase-

controlled high-resolution industrial camera (IDS) were used for the

visualizations. The high speed camera, it was possible to took videos with an frame rate of 30.000. So the jet-bucket interaction at the different operation points of the turbine could be captured properly. During and between the hill

chart measurements, additional images were recorded with a so-called "action

camera" to visualize the flow pattern inside the turbine housing. According to which observation of the flow pattern (jet, bucket, overall flow pattern) should be observed, it is essential to choose the right camera system according to framerate, resolution. The investigations using the “action camera” resulted in global views on the outlet jet interacting with the housing. One of the

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main problems of the action camera system was the fact, that there is no option of triggering the camera shutter. So the next step of going into further detail of the flow visualization was to use a Ethernet-driven CMOS-camera with an external trigger and a triggered stroboscope light. With this equipment, it was possible to capture the jet and bucket interaction inside the housing by taking a picture after every revolution of the runner. This resulted in “quasi high speed” videos; they allow only a semi optimal view on the interaction between jet and bucket, because between every single picture of the video is one full rotation of the runner.

The last and most complex option for observing the flow pattern is to use high-speed cameras. With high-speed cameras with up to 675.000 frames per second (fps), it was possible to get real high-speed videos from jet-bucket interaction. The figures on the next page are showing some examples of high-speed videos. Overall, the research project shows that flow visualisation helps to understand the flow pattern of the Pelton turbine and furthermore helps to improve the nozzle and bucket design to improve the turbine characteristics.

Figure 96: Pictures taken by the triggered CMOS-camera systems

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Figure 97: Pictures taken by the high-speed camera (15000 fps) from outside the housing (normal view)

Figure 98: Pictures taken by the high-speed camera (5400 fps) from outside the housing (upper isometric view)

Figure 99: Pictures taken by the high-speed camera (5400 fps) from outside the housing (lower isometric view)

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3.23 3D-scanning and reverse engineering of hydraulic sufaces

When starting a power plant refurbishment project by using CFD-calculations a complete model of the installed hydraulic equipment is needed. Modelling the draft tube and the volute casing of the turbine is usually possible from drawings. To get the real surface of the installed guide vanes and runner blades a 3d-scanning technology is required. The institute of hydraulic fluid machinery has therefore purchased a laser scan arm from FARO. The complete scanning hardware system is portable to be able to use inside an installed turbine. For the scanning process, an additional pc-system is necessary with the software application GEOMAGIC installed. The reachable area is bounded by a sphere with a radius of 90 cm (length of the scan-arm). To scan objects outside of this sphere the scan-arm has to be repositioned. To link the scanned point cloud of two or more “scenes”, reference targets have to be mounted on the object. As targets bearing balls have proved to be the best solution, because they are small and almost perfect spherical. In to “scene” a minimum of 3 identically targets have to be scanned and can then be used as reference between the two-point clouds.

Figure 100: Scanning of an old Kaplan runner blade from the 1960s

As the first result of the scanning process, a point cloud is recorded. For comparing a given CAD surface with the measured data only the point cloud is needed. To get a closed surface from the recorded points, they have to be meshed. The meshing process is also executed in the application GEOMAGIC, the resulting surface consists of triangles, as known from the *.stl-file format.

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Figure 101: Triangulated surface and extracted profiles of a Kaplan runner blade

In further steps, the triangulated surface is then exported to the CAD software package CATIA V5. The aim is to get a clean and smooth parametric surface. For this step, the automatic surface reconstruction tool can be used, but the result does not fulfil the requirements in precision and smoothness for CFD modelling. So usually, the surface reconstruction or reverse engineering has to be done manually. The first step is to assign and centre the scanned surface model to a given coordinate system. Then, the geometry is sliced streamwise into segments and only the extracted profile curves are used for the analysis and reconstruction of the surface. In most cases, the profile curves have to be smoothed to get the best basis for a smooth surface. For the following meshing process for a CFD-analysis the extracted and smoothed profiles are used, but in some cases a surface model is needed for a comparison, e.g. between the old hydraulic contour and the new designed surfaces or for the comparison between all runner blades of one runner.

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Figure 102: Comparison of all five Kaplan runner blades, to measure the misalignment between the blades and the runner hub (left) and the comparison of a casted valve body with the corresponding CAD-model (right)

In the last few years, we experimented with different workflows to optimize the way from the scanned cloud to the CFD analysis. For Kaplan runner blades, Francis runner blades, big storage pump blades and guide vane blades we found a more or less simple and fast way for these steps. We also had the chance to use our workflow on other problems as the reverse engineering of a big 3-way valve. The first target was to compare the given CAD model and the scanned surface. In the next steps, we tried to get a closed surface model from the scanned data to be able to use not the idealized CAD but the real casted geometry for a CFD analysis (not the real surface structure but the influence of the shifting of the casting core during the casting process and of other faults).

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3.24 Investigations of the intake structure of the power plant GKI/Inn

A new diversion-type power plant located at the Swiss-Austrian border at the river Inn is currently under construction. The intake structure located in Ovella contains a 15m high weir channel to collect water for the GKI power plant. The weir was planned as a multi-functional structure and includes the weir valves for the barrage, the inlet channel for the approximately 23 km long gallery to the main stage (powerhouse Prutz / Ried), a small turbine for residual water delivery and a device for barrier-free fish migration. The plant data of the power plant project are summarised in the following table. Table 2: Turbine main data of HPP GKI

Main turbine Residual water turbine

Planned start-up 2018

Type of turbines 2 x vertical Francis 1 x Kaplan S-type

Net head rated 160.7 m 15 m

Flow rate rated 75 m³/s 20 m³/s

Fig. 103 (left) shows a visualisation of the planned reservoir with the intake structure on the orographic right side, which was investigated in the course of a comprehensive CFD study. The flow rate at the intake structure accounts for max. 95 m³/s. The maximum discharge of the main turbines is 75 m³/s. The remaining amount of water is used by the small residual water turbine. Additionally it needs to be mentioned, that – as an additional special feature – the use of a fine screener with a clear width of only 20 mm was prescribed at the transition from the reservoir to the intake structure. In order to obtain a flow velocity of max 0.75 m/s in front of the trash rack and a mean velocity of 1 m/s in the trash rack itself, the dimension of the screener was determined to be: width = 19 m and height = 6.5 m (steel bar width 8 mm).

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The focus of the investigations carried out based on CFD calculations was on the optimisation of the flow guidance in the inlet structure in order to obtain an ideal flow of the calculation field and to improve the geometric connection of the small residual water unit and thus to ensure a swirl free homogenous flow situation at the inflow there. Fig. 103 (right) shows the 3D CFD model used as well as an example result of the calculated phase boundary (= water surface).

Figure 103: Planned hydro power plant intake structure and CFD model

The CFD calculations were all carried out for the maximum flow rate at the intake structure (Q = 95 m³/s) at the design water level (Hmax = 1029.5 m) and the minimum water storage elevation (Hmin = 1025.5 m). The implementation of the fine screener was a specific challenge in the CFD calculation, as it resulted in a particularly complex two-phase CFD model. While a mesh with approximately 5 million nodes was created for the course of the river and the inlet channel, approximately 25 million hexahedral elements had to be used in the area of the fine screening. The CFD model used for the calculation from the storage reservoir to the intake structure is shown in Fig. 104. Velocity vectors were plotted on an evaluation plane placed in the middle of the flow field. The left picture shows the flow situation in the case of the originally planned geometry variant. While the screener field in the middle (pos. 2) is expected to flow well, a false flow appears to occur at the edges (pos. 1 and pos. 3). This leads to pronounced flow separations between the screen bars in these areas (see Fig. 104 – left) and thus the flow is not divided equally across the screener field.

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Furthermore, the inclined computing field of the inflow can cause vibrations on the bars. The justified alignment of inlet structure and screener field therefore appears to be inappropriate.

The flow situation at the inflow is more homogeneous and looks much better in the case of the improved geometry (see Fig. 104 – right). Due to the offset of the plane, where the screener is located in a distance of about 2 m to the rear and the attachment of an elliptical fillet before the area 1 and a circular contour in front of the area 3, a false flow can be prevented (see also Fig. 103 – right) and a homogeneous velocity distribution over the entire calculation field is ensured.

Figure 104: Velocity distribution at different locations

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Further improvement of the originally planned design of the intake structure becomes apparent when analysing the flow situation in a side view. Figure 105 shows a visualisation of streamlines in a vertical central section through the inlet structure. The left-hand illustration shows, that in the original design along the sharp edge at the transition from the flow to the inlet channel (marked red) pronounced detachment zones occur, which on the one hand cause flow losses, on the other hand these zones can lead to oscillation problems in the screener area. Furthermore, the shape of the screener’s beam also does not seem ideal. A fillet around the intake structure and the use of two beams turned into the flow direction significantly improve the flow situation.

Figure 105: Streamlines, connection of headrace channel

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3.25 Analysis of the leakage behaviour of Francis turbines and its impact on the hydraulic efficiency - A validation of an analytical model based on CFD-results

The research project addresses the analysis of the leakage behaviour of a small hydro Francis turbine using an analytical approach which was validated based on the results of Computational Fluid Dynamics (CFD). For a custom-designed Francis turbine with a specific speed of nq = 41.9 rpm the flow chambers resulting from the labyrinth geometry were added to a traditional full CFD model of the turbine and numerical simulations were performed for several operation points ranging from part load (Qmin = 0.5 * Qopt) to over load (Qmax = 1.3 * Qopt). Fig. 106 presents the surface mesh of the Francis runner with the runner side spaces added to the numerical model.

Figure 106: Visualization of the runner mesh with the corresponding labyrinth seals on crown and band side

Consequently, the single losses occurring in the runner seals on crown and band side as well as the pressure distribution within the runner side spaces could be evaluated and compared to the results gained with an analytical approach which was originally developed to calculate the leakage flow of centrifugal pumps. Fig. 107 depicts the loss composition calculated by means of CFD-simulations. If the runner side spaces would have been neglected the orange coloured efficiency curve would have been calculated. The red coloured curve represents the total

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hydraulic turbine efficiency. It turns out that the leakage and disc friction losses occurring in the runner side spaces yield a significantly reduced efficiency. The lower the flow rate the higher are the losses in the labyrinth region on crown and band side.

Figure 107: Hydraulic turbine efficiency IEC and loss composition for various operation points at a constant turbine head of H = 125 m

A comparison of the analytically and numerically calculated pressure distribution in the runner side space on the crown side of the runner is depicted in Fig. 108. The comparison of the pressure distribution achieved with the numerical simulation and the analytical calculation shows that both approaches match well

if the angular velocity of the fluid ωFluid trapped in the runner side spaces is calculated in an appropriate way. Furthermore the analytical model also enables the estimation of the leakage flow and disc friction losses occurring on crown and band side of the runner. Referring to the best efficiency point the results of the analytical calculation were compared with the results of the numerical simulation which is shown in Fig. 109. The achieved results demonstrate that the use of the analytical model enables the calculation of these losses with sufficient accuracy.

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Figure 108: Pressure distribution in the seal region on crown side based on a numerical and analytical approach

The results of the research work contribute to the improvement of the performance prediction of Francis turbines based on combined numerical and analytical calculations.

Figure 109: Overview of the analytical and numerical results for leakage (left) and disc friction (right) - referring to BEP

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4. Graduation work

4.1 Masterthesis Stefan Leithner, BSc: Conceptual Design of an Ultra-Low Specific Speed Pump, 2016 Tutor: Assoc. Prof. Dipl.-Ing. Dr.techn. Helmut Benigni Since the environmental regulations are becoming increasingly strict, the efficiency of pumps is a more and more important selection criterion. However, reliability still remains a key priority for the facility operator. Single stage pumps with a specific speed nq = 40-50 rpm already achieve very high efficiency levels. This does not apply to pumps with a specific speed nq < 10 rpm. As part of this master thesis, a new design for pumps with specific speed 4 ≤ nq ≤ 6 rpm is presented. In the first part of the thesis, different pump types (even of classical centrifugal pumps) with low specific speeds are described. Then, the redesign of an impeller of a pump with volute casing is shown. These pumps are designed to meet the requirements of the API 610 standard [1]. The pump design starts in the classic way one-dimensional and is realized by means of analytical methods. With the help of CFD, the flow inside the pump is simulated, and finally the results of several variants are compared. Philipp Günther Kandutsch, BSc: Conceptual Design of an Ultra-Low Specific Speed Pump, 2016 Tutor: Dipl.-Ing. Stefan Höller-Litzlhammer The aim of this thesis was to work out an algoithmus which can be used to calculate the free surface flow. The developed method of characteristics was programmed using the software MATLAB and comparing the results with analytical test cases. Afterwards the method of caracteristics was implemented in the water chamber model used at the Intitute of Hydraulic Fluid Machinery.

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Patrick Zeno Sterzinger: Flow visualization in a pump-turbine, 2017 Tutor: Dipl.-Ing. Mark Guggenberger Due the increasing need of storage capabilities in the electric sector, pump turbines will become increasingly important. In many cases a Francis pump turbine is used and this is the reason why performance and efficienc are gaining in importance. This study focuses on the unsteady behaviour of Francis pump turbines. Francis pump turbines exhibit a sudden head drop of the pump characteristic at low discharge coefficients. Because the head drop results in a loss of efficiency, and increasing pressure pulsations, it is essential to gain a better understanding of the flow physics leading to this instability. A reduced scale model of a pump turbine was installed on a 4-quadrant test rig of the institute for hydraulics machinery. In order to gain optical access to the high pressure side of the runner the model has been modified while keeping the same hydraulic behaviour.\\ 2d Particle Image Velocimetry (2d PIV) measurements were performed with varying guide vane openings on two different runners. On the pressure side in the vaneless space between impeller and guide vanes the velocity field has been measured in four planes. For each runner between four and six operation points have been investigated. The interaction between runner and guide vanes was analysed using a phase-averaging method at 21 different runner positions. In further consequence the measured and averaged velocity fields enabled the detection and localisation of regions with a inhomogeneous flow distribution and recirculation regions. In addition, measurement data provided can be used to validate CFD simulations in order to improve the accuracy of pump characteristics, particularly for pump turbines. Gerd Siegfried Staudacher, BSc: Experimental investigation of the shaft seals of vertical Kaplan-turbines, 2017 Tutor: Assoc. Prof. Dipl.-Ing. Dr.techn. Helmut Benigni; Dipl.-Ing. Dr.techn. Jürgen Schiffer-Rosenberger For several years now, newly installed and existing Kaplan turbines are equipped with lip seals made of elastomer that act in axial direction of the turbine shaft. The advantages of this kind of sealings are the easy and space saving construction and that they can be operated without sealing water. The desired

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service time of these seals should reach up to six years, however, in some cases, failures occurred significantly earlier, resulting in shorter inspection intervals of the turbines. Power plant operators assume that relatively large axial movements of the turbine shaft might be the reason for the shortened duration. The present thesis aims at examining the influence of the mentioned turbine movements on the wearing behaviour of the sealing. Therefore, several seals with different geometries where tested on a test rig. As results of these tests it can be stated that axial displacement combined with acting water pressure have a major impact on the wear of the seals. Furthermore, a comparison of the different geometries used in power plants and expected to positively influence lifetime, did not result in an improvement of wear behaviour. Based on the experience made and the test results, the author of this thesis found a wear-free type of sealing that worked with the requested boundary conditions on the test rig. Furthermore, the theoretical part presents basics of lubricated rubber friction and rubber wear as well as sealing technology, and a more suitable seal material found by means of literature study. Based on the facts and results presented, a proper working lip seal was installed in a Kaplan turbine of the Aschach power plant on the river Danube in December 2016 and operates successfully. Thomas Reisinger, BSc: Implementation of water lubricated plain bearings in pulp processing machines, 2017 Tutor: Assoc. Prof. Dipl.-Ing. Dr.techn. Helmut Benigni A common characteristic of pulp processing machines is in most cases the cantilevered rotor and in all cases a bearing housing which is locally separated from the flow housing. Therefore, the shaft has to be sealed against the flow housing as well as against the bearing housing. This design is established, to ensure that there can never be a contamination of the procress water with lubrication oil or grease in the case of sealing failure. It leads to a long cantilevering length of the shaft, that provides high bearing loads and high bending moments on the shaft. By bringing together the shaft sealing against the flow housing and the front bearing functionally and locally the cantilevering length of the shaft is decreased tremendously. This requires a bearing concept, that is either lubricated by fresh water or process water. By using this bearing concept, the machines get smaller, lighter and manufacturing costs are reduced. The thesis is based on the Andritz product portfolio and was established in close

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cooperation with the responsible persons there. The water lubricated bearing concept was implemented in various pulp processing machines. Additionally, this bearing concept was compared with the established designs regarding manufacturing costs. Rudolf Peyreder, BSc: Investigation and modification of a calculation method for the determination of dynamic stresses on francis runners, 2017 Tutor: Assoc. Prof. Dipl.-Ing. Dr.techn. Helmut Benigni; Dipl.-Ing. Tomas Ries The present thesis deals with the investigation and modification of a method for the determination of dynamic stresses in Francis turbines. The key aspect of the thesis is the modification of the structural analysis to improve the accuracy of the existing method. As a first step an example is investigated and possible potential for improvement is shown. The com- putation steps carried out are described and considered critically. The investigation shows that a shift of the interface between the flow analysis and the structural analysis leads to an improvement of the transfer. The identified modification is implemented in the existing calculation method, which required a change of the software suite. The modified method is also carried out for the example and both results are compared. The comparison of the calculation methods shows an improvement of the prediction accuracy and enables a correct mapping of the prevailing pressure level in the fluid. In order to be able to estimate the amount of improvement, both methods are compared with an on- site measurement of a prototype runner. For the modified method, a new approach of damping is selected in which the damping effect is modeled by an energy transfer in the fluid. The comparison shows a noticeable improvement in the prediction accuracy as well as the possibility for the introduction of further modifications.

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Florian Martin Höfler, BSc: Model test and flow visualization of a two-jet pelton turbine, 2017 Tutor: Dipl.-Ing. Mark Guggenberger The Institute for Hydraulic Machinery at the University of Technology in Graz has performed measurements and flow visualizations in cooperation with Global

Hydro Energy GmbH, an internationally successful company in the field of small‐scale hydroelectric technology, at a dedicated test rig for a model of a pelton turbine. Various impeller and housing variants were tested on the test rig. Values for pressure, flow, water temperature, torque and rotational speed were recorded according to the standard, evaluated and presented. After the measurements, photos and videos were taken in highspeed from the inside of the turbine

housing. A highspeed camera and a phase‐controlled high‐resolution industrial camera were used for the visualizations. During and between the measurements, additional images were recorded with a so‐called "action camera". At the beginning of this thesis an overview of the basics of a pelton turbine is given and the components of the turbine as well as the structure of the model including the test rig are described in detail. In addition, the used measurement technology is explained and the test rig is described. The results of the realized measurements are summarized and presented in a characteristic diagram. The second part of the thesis describes the flow visualization with different camera models. The visualizations include recordings of the external housing with an action camera, triggered recordings and highspeed recordings from inside the housing. The influence of the housing geometry and the interaction between the bucket and the jet is observed and the possibilities of visualization in the Pelton turbine is explained. The preparation for taking the images and the videos, the used camera technology and the equipment used for illumination and guarding the camera lens coverage from water droplets are also described in course of this thesis.

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Martin Haubenhofer, BSc: Numerical Simulation of a Diagonal Turbine, 2017 Tutor: Assoc. Prof. Dipl.-Ing. Dr.techn. Helmut Benigni The most commonly used turbines in water power plants so far are Francis, Kaplan or Pelton turbines. Nevertheless, in off-design operation of Francis turbines high hydraulic losses and unstable conditions occur due to fixed runner blades. A type of turbine with a similar operation range up to 150 meters drop head is the double-regulated diagonal turbine. So far, this kind of turbine is mostly used as pump turbine in pump storage power plants. Within the scope of this Master´s Thesis runners of a diagonal turbine were designed at the Institute of Hydraulic Fluid Machinery by use of Computational Fluid Dynamics (CFD) for the first time. Later on, performance and behaviour of cavitation were compared with the results of a Francis reference turbine which was carried out at the Institute of Hydraulic Fluid Machinery for an industrial partner. The results clarified a smoother behaviour of efficiency over the discharge and lower hydraulic losses at part load operation in comparison with the Francis reference turbine. The best efficiency at the design discharge of the diagonal turbines was approximately one p ercentage point lower than the maximum efficiency of the Francis reference turbine.

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5 Continuing education

5.1 University Programme in Hydropower In this interdisciplinary programme for further education, TU Graz combines since 2014 expertise from the fields of mechanical engineering, electrical engineering and civil engineering to provide education at the highest level. Environmental protection, sustainability and ecology are included in the programme, as are organisational and economic aspects. Become an expert in the complex field of hydropower! Content The programme contains the following courses on ecological aspects and the three areas of Mechanical Engineering, Electrical Engineering and Civil Engineering:

In the master’s degree programme Master of Engineering (MEng) – Wasserkraft, you also write a master’s thesis.

Mechanical Engineering

� Basic hydraulic principles � Mechanical engineering materials � Machine dynamics � Design details � Pumped-storage plants

Electrical Engineering

� Basics of electrical machines � Energy transport, circuit types � Control technology for

hydraulic machines

Civil Engineering

� Surge tank concept � Dam structures, sealing

systems � Construction management � Hydraulic steel structures

Interdisciplinary

� The energy industry � Operational management,

maintenance � Small-scale hydropower plants � Construction permit planning � Potential and location analyses

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Structure of the Programme The programme follows the Blended Learning approach:

� Online teaching – customised documents ensure fast and successful learning – you learn when and where you wish.

� Compulsory attendance events – Experts impart complex factual knowledge. You work on projects with other participants, putting your knowledge into practice.

• English programme: one week per semester

• German programme: 6 to 8 days per semester – block scheduling on Fridays and Saturdays

� Excursions – by means of practical examples on site e.g. in power

plants, you get a realistic insight into the topics.

At the end of every course, you take an examination.

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Target Groups and Admission Requirements This university programme for further education is targeted at people from the field of hydropower who wish to develop their competence technically and professionally. 5.1.1 Master of Engineering (MEng) – Hydropower (5 Semester) Admission requirements � a completed relevant bachelor’s degree programme (Mechanical

Engineering, Electrical Engineering, Civil Engineering) from a university or a university of applied sciences or

� a completed technical diploma or master’s degree programme from a

university or a university of applied sciences or

� a completed Magister, diploma or master’s degree programme in natural

sciences or economic sciences from a university or a university of applied sciences plus at least 3 years’ experience in the field of hydropower.

5.1.2 Akademische Wasserkraftingenieurin/Akademischer Wasserkraftingenieur (MEng) (4 Semester)

Admission requirements � a completed examination as an industrial foreman or a mechanical

engineering technician according to the Trade, Commerce and Industry Regulations (GewO) and at least 8 years’ relevant professional experience in the area of hydropower or

� graduation (with a secondary school leaving examination – Matura/Abitur) from a Higher Technical Institute (HTL) or a general secondary school (AHS) with relevant professional experience and at least 5 years’ industry practice in hydropower

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Organisation Graz University of Technology Life Long Learning Mandellstraße 13/II 8010 Graz www.LifeLongLearning.tugraz.at Margot JABERG Mag. Tel. +43 316 873-4936 Programme Director Institute of Hydraulic Fluid Machinery Kopernikusgasse 24/IV 8010 Graz www.hfm.tugraz.at Helmut JABERG Univ.-Prof.-Dipl.-Ing. Dr.-techn. Tel. +43 316 873-7570

Figure 110: University Programme Hydropower - Kick-off March 2015

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5.2 pump.ing – Vocational training for pump experts

pump.ing. is an extra-occupational distance learning study programme for pump-specific expertise regarding pump technology, operational issues and systems. The study course lasts 400 hours and is divided into the chapters "Pump basics", "Pump unit", "Pumps in systems" and "Fields of application (optional subjects)". Out of eight arbitrary fields of application offered (process engineering, refineries, power plants, water, sewage, food and sterile industry, paper industry and vacuum technology) two more subjects have to be chosen according to individual interest or industry. Via the TU Graz teaching and learning platform TUGonline all students have access to the study material and are totally free with regard to their self-study time management. In addition, they have to participate in eight attendance events of two days each (Friday and Saturday) at various locations in Austria and Germany. The attendance event "Pump assembly" takes place at the laboratory of the Institute of Hydraulic Fluid Machinery at Graz University of Technology. The duration of the course is about 14 months. The acquired knowledge on the optimum design and operation of pumps and systems and the ability to conduct energy counselling are documented by the certificates "Certified energy consultant for pumps and systems" and “Pump Engineer”.

Practical training by renowned pump specialists: The German study course for pump engineers starts every year in July. In 2018 already the 14th study course has been launched. In 2005 the idea of setting up a special study programme for pump specialists arose, on the one hand, on the initiative of Professor Helmut Jaberg, Head of the Institute of Hydraulic Fluid Machinery at Graz University of Technology. On the other hand, its setup was based on considerations by the pump section of the German Association of the chemical industry VCI with its then chairman, Dr. Friedrich Wilhelm Hennecke, longstanding head of BASF's Pump Centre. At the annual Practitioner’s Conference on pumps – "Pumpen in der Verfahrens-, Abwasser und Kraftwerkstechnik" – held in Graz, the participants were asked about the project to install the study programme “Pump engineer” respectively “Pump technician” (depending on the level of professional qualification a participant presents at the beginning of the programme). The proposal was received enthusiastically. Manufacturers were hooked and operators could finally acquire specialist knowledge about pumps. Moreover, plant operators were very pleased as they could finally count on qualified pump salesmen. Fortunately, participants from both the manufacturers’ and the operators’ side attend the study course for pump

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engineers in large numbers. Prof. Helmut Jaberg and Dr. Friedrich Wilhelm Hennecke also brought Professor Paul-Uwe Thamsen from the Technical University of Berlin and Dr. Walter Schicketanz, then head of a planning unit at BASF, on board. The Pumpenfachingenieur GesmbH was founded and a high-ranking management board, in which the pump section of the the VCI is still prominently represented, was installed and has since then defined the teaching content. The first study course was launched in 2005 with around 30 lecturers – all of them leading pump experts with well-founded and relevant industrial experience. More than 350 pump experts from the top industry representatives have since been exclusively trained as pump engineers and pump technicians pump.ing. is highly appreciated by the industry. The practical relevance and the high standard of the teaching content probably contribute to this. Ambitious goals and challenges – also of international significance – are a welcome incentive for us, and of course we aim at keeping the programme at the latest state of the art.

Figure 111: pump.ing – theory and practice in the laboratory: pump assembly

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pump.ing. – International study course in English Since 2012 already five international study courses for pump engineers have been launched. The idea of offering the course for an international audience is based on the demand of globally operating companies wishing to employ qualified pump experts worldwide. Companies such as Netzsch, Sulzer, Andritz AG, Shell Petroleum, Grundfos, Egger Turo Pumps are just some of our international customers. The English-language programme corresponds to the syllabus of the German study course for pump engineers. However, there are only two attendance events of two weeks each. Managing Director pump.ing: O.Univ.-Prof. DI Dr. Helmut JABERG Tel.: +43 316 873 7570 [email protected] Management: Mag. Karin Hermann Tel.: +43 316 873 8079 [email protected]

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5.3 Practitioners’ conferences

Practitioners’ conferences are an integral part of the Institute's calendar of events. The attractiveness of the concept developed by Prof. Jaberg in the last 20 years is unbroken, as each conference attracts up to 150 participants to visit Graz. The mission statement is "by practitioners for practitioners" and means that renowned lecturers from the industry will take part in the programme – manufacturers and planners as well as operators. The fact that lectures have to be rejected every year because the agenda is already overflowing proves the importance of the conference in the German-speaking world as a forum for exchange between experts in the field of hydraulic fluid machinery.

What distinguishes practitioners’ conferences from similar events? First of all, the speakers come from the front row of companies, the presentations are given by leading engineers and the focus is on technical topics. The generous time frame allows to go into depth during the lectures and not just touch on challenging points. Another advantage is that the conference is held in German. And this will not change, despite a growing number of visitors and speakers from non-German speaking countries: An open dialogue is important to us, and the lively discussions after the lectures confirm that this path is well accepted. Every year, representatives of operators find the time to report on their experiences and problems. After all, the attending audience of experts could often provide suggestions for essential solutions to solve one or the other operating problem. An additional plus to be mentioned is the framework of the conferences – as a trip to Graz should be worth it in every respect! “Regular guests" never leave

Figure 112: Practitioners‘ conference

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Graz without a bottle of pumpkin seed oil, and the atmospheric evenings in the historic buildings of Styria and the city of Graz offer plenty of time to pursue ideas which emerged during the day and make new contacts.

The conference "Pumps in process engineering" organised by practitioners provides a unique platform for practitioners to share experiences, new developments and ideas for coping with the everyday challenges of pumps and pump systems. The Practitioners' Conference Pump especially addresses operators as well as planners and manufacturers of pumps and pump systems.

23rd Practitioners’ Conference Graz

“Pumps in process engineering” May 6 – 8, 2019

Graz, Austria

More information and registration: www.praktiker-konferenz.com

This German-language conference provides the ideal platform to build up networks as well as for the exchange of experience and knowledge by practitioners for practitioners – manufacturers, operators, planners and researchers – focussing on hydro power plants and core technical / economic topics such as turbine efficiency, pressure surge, turbine in- and outtake, pump storage, control units, surge chamber, penstock, service life, refurbishment etc. covering the range from small to large hydro power.

6th Practitioners’ Conference Hydropower

“Practitioners’ Conference Hydropower / Turbines / Systems”

September 10 – 11, 2019 Graz, Austria

More information and registration:

www.wasserkraft-graz.at

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6 Publications

6.1 Articels in journals

6.1.1 2017 Schiffer-Rosenberger, J., Benigni, H. & Jaberg, H.

An analysis of the impact of draft tube modifications on the performance of a Kaplan turbine by means of Computational Fluid Dynamics - in: Proceedings of the Institution of Mechanical Engineers / C. 16 p (22 Jun 2017)

6.1.2 2016 Schiffer-Rosenberger, J., Benigni, H. & Jaberg, H.

Analysis of the Leakage Behavior of Francis Turbines and Its Impact on the Hydraulic Efficiency—A Validation of an Analytical Model Based on Computational Fluid Dynamics Results - in: Journal of fluids engineering. 139, 2, 11 p., 021106-2 (2 Sep 2016)

Jaberg, H.

Innovationspotenzial bei Pumpen - in: Chemie & more. 1, 16, p. 22-27 6 p. (Jan 2016 )

Höller-Litzlhammer, S., Benigni, H. & Jaberg, H.

Investigation of the 4-Quadrant behaviour of a mixed flow diffuser pump with CFD-methods and test rig evaluation - in : IOP Conference Series / Earth and Environmental Science. 49, 10 p. (Nov 2016)

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6.2 Contributions at conferences 6.2.1 2017 Schiffer-Rosenberger, J., Benigni, H., Jaberg, H. & Penninger, G.

A verification of draft tube modifications of a vertical axis Kaplan turbine by means of CFD – in: Proceedings of AFRICA 2017: Water Storage and Hydropower Development for Africa. The International Journal of Hydropower and Dams, 10 p. 08.04 (14 Mar 2017)

Benigni, H., Schiffer-Rosenberger, J., Guggenberger, M. & Jaberg, H.

Erhöhung des Jahresarbeitsvermögens von Kleinwasserkraftanlagen durch gezieltes Refurbishment anhand von Fallbeispielen: Erhöhung des Jahresarbeitsvermögens von Kleinwasserkraftanlagen durch gezieltes Refurbishment anhand von Fallbeispielen – in: 20. Internationales Anwenderforum Kleinwasserkraftwerke. PSE AG (ed.). Brixen, Vol. 20, p. 17-24 8 p. (28 Sep 2017 )

Benigni, H., Mosshammer, M., Jaberg, H., Hölbling, M. & Mayrhuber, J.

Flow analysis and operational optimisation of a two unit bulb turbine hydro power plant with significant power dif-ference – simultaneous CFD calculation of both turbines – in: Hydro 2017: SHAPING THE FUTURE OF HYDROPOWER. Aqua~Media International Ltd., Vol. 2017, p. 3.10 8 p. (9 Oct 2017)

Schiffer-Rosenberger, J., Benigni, H., Mosshammer, M. & Jaberg, H.

Verbesserung der Zu- und Abströmung von Kraftwerksanlagen – in: Tagungsband: 20. Internationales Anwenderforum Kleinwasserkraftwerke 28. – 29. September 2017, Brixen, Italien. PSE AG, p. 105-112 8 p. (28 Sep 2017)

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6.2.2 2016 Mosshammer, M.

100% Flexibilität bei Pumpspeicheranlagen: Hydraulischer Kurzschluss als zusätzlicher Betriebsmodus – Simulation und Validierung – in: 14.Symposium Energieinnovation: Energie für unser Europa. I. F. E. U. E. . I. E. E. (ed.). Verlag der Technischen Universität Graz (1 Mar 2016)

Höller-Litzlhammer, S., Jaberg, H., Benigni, H. & Kim, J. J. Cavitation optimization of a variable pitch mixed flow pump for cooling water by numerical mehtods and test rig verification – in: Pump Users International Forum 2016: Proceedings. Düsseldorf: VDMA, p. 777-789 13 p. 18-2 (14 Sep 2016)

Widmer, C., Spyrou, N., Jaberg, H., Senn, F. & Guggenberger, M. Comprehensive analysis of the saddle curve of pump turbines in pump mode using PIV and CFD – in: IAHR Conference on Hydraulic machines and systems, Grenoble, 4-8 July 2016 : Publication in IOP Conference Series (4 Jul 2016)

Benigni, H., Leithner, S., Schiffer-Rosenberger, J., Jaberg, H. & Höller-Litzlhammer, S.

Development of a novel centrifugal pump with lowest specific speed – in: Pump Users International Forum 2016: Proceedings. Düsseldorf: VDMA, p. 179-192 14 p. 5-3 (14 Sep 2016 )

Guggenberger, M., Senn, F., Jaberg, H., Gehrer, A., Sallaberger, M. & Widmer, C.

Experimental analysis of the flow pattern of a pump turbine model in pump mode – in: IAHR Conference on Hydraulic machines and systems, Grenoble, 4-8 July 2016 : Publication in IOP Conference Series (4 Jul 2016)

Guggenberger, M., Senn, F., Sallaberger, M. J., Gehrer, A. & Widmer, C. High speed visualization of the part load flow of an impeller in pump-mode – in: Pump Users International Forum 2016: Proceedings. Düsseldorf: VDMA, 12 p. 15-2 (13 Sep 2016 )

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Höller-Litzlhammer, S., Jaberg, H., Benigni, H. & Kim, J. J. Low specific speed mixed flow API pump for single and multistage usage - multi-objective design challenge – in: Pump Users International Forum 2016: Proceedings. Düsseldorf: VDMA, p. 167-178 12 p. 5-2 (14 Sep 2016 )

Mosshammer, M.

Numerische Untersuchungen und Optimierung einer mehrstufigen Radialpumpe - Validierung anhand Prüfstandsmessungen – in: Conference Proceedings - 34. CADFEM ANSYS Simulation conference (1 Dec 2016)

Schiffer-Rosenberger, J., Bodner, C. & Jaberg, H.

Performance analysis of a single-blade impeller pump based on unsteady 3D numerical simulation – in: Proceedings of the 3rd International Rotating Equipment Conference. VDMA, p. 193-203 10 p. (14 Sep 2016 )

Schiffer-Rosenberger, J., Benigni, H. & Jaberg, H.

Refurbishment von Bestandsanlagen unter Verwendung der numerischen Strömungssimulation – in: Proceedings des 19. Internationalen Anwenderforums Kleinwasserkraftwerke. p. 50-55 6 p. (22 Sep 2016 )

Widmer, C., Neubauer, R., Jaberg, H., Senn, F. & Guggenberger, M.

Simulation and measurement of unstable pump characteristics: comparison with PIV – in: HYDRO 2016, Montreux: Proceedings. Montreux, 9 p. 34 (12 Oct 2016)

Benigni, H., Montenarie, B., Jaberg, H., Schiffer-Rosenberger, J., Gehrer, A., Grundner, F. & Döltelmayer, R.

Simulation of damage caused by cavitation in non-rotating components of a Kaplan turbine – in: Hydro 2016: Achievements, opportunities and challenges. Aqua~Media International Ltd., 8 p. 4.04 (10 Oct 2016 )

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ACTIVITY REPORT2016 – 2017

Graz University of TechnologyInstitute of Hydraulic Fluid Machinery (HFM)Kopernikusgasse 24 / 48010 Graz, Austria

T +43 316 873 7571F +43 316 873 [email protected]



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ACTIVITY REPORT 2016 – 2017 · with PIV or LDA is our daily business at the highest level. For valve research we dispose of appropriate research loops conforming to the relevant