investigating the dynamic aspects of the...

6th IAHR International Meeting of the Workgroup on Cavitation and Dynamic Problems in Hydraulic Machinery and Systems, September 9-11, 2015, Ljubljana, Slovenia

* Corresponding author: Institute of Hydraulic Fluidmachinery, University of Technology Graz, Graz, Austria, phone: +43 316 8072, email: [email protected]

INVESTIGATING THE DYNAMIC ASPECTS OF THE

TURBINE INSTABILITY OF A PUMP TURBINE MODEL

Mark Guggenberger* Institute of Hydraulic Fluidmachinery, University of Technology Graz

Florian Senn Institute of Hydraulic Fluidmachinery, University of Technology Graz

Jürgen Schiffer Institute of Hydraulic Fluidmachinery, University of Technology Graz

Helmut Jaberg Institute of Hydraulic Fluidmachinery, University of Technology Graz

Manfred Sallaberger Andritz Hydro, Switzerland

Christian Widmer Andritz Hydro, Switzerland

ABSTRACT Pump-storage power plants provide the only valuable contribution to efficient storage of primary energies at large scale. Equipping power plants with reversible pump turbines additionally provides high flexibility to the electricity market. The possibility to operate pump turbines even at off-design conditions as well as the option to change the operation mode fast and frequently are the major advantages of reversible pump turbines. However, the operation of reversible pump turbines in such a dynamic way is also connected to the occurrence of unstable flow phenomena between guide vanes and runner, resulting in a change of sign of the slope of the head curve in turbine brake mode. Consequently, the so called S-shaped characteristics appear, which can seriously impede the start-up process of a power plant. These unstable operation conditions can cause hydraulic system oscillations resulting in large forces on the hydraulic equipment. Numerical simulations of these unstable off-design conditions are challenging due to the complexity flow patterns in this operation mode. However they indicated as the occurrence of turbine instabilities is connected to the appearance of rotor-stator interactions and backflow regions. The present study aims at an improved understanding of the origin of the fluid mechanical mechanisms responsible for the occurrence of unstable behaviour. Therefore, a reduced scale model of a radial pump turbine was integrated into a 4-quadrant test rig based on the IEC60193 standard featuring the possibility of adjusting stable operation even in unstable regions of the characteristics. The guide vane apparatus, the spiral casing outlet, the draft tube inlet and the surroundings of the pump turbine model were equipped with dynamic wall pressure sensors. Thus it was possible to perform measurements of rotor-stator interactions as well as to investigate the development and the spread of flow instabilities. In order to verify the dynamic aspects of the flow instability, detailed analyses in the frequency domain have been carried out for several operation

IAHR WG Meeting on Cavitation and Dynamic Problems in Hydraulic Machinery and Systems, Ljubljana 2015

points spread over the entire turbine quadrant. The results of the dynamic pressure analysis offered a global overview of the characteristics and additionally indicated interesting operation points which were analysed with laser optical measurements. Particle Image Velocimetry (PIV) allowed to measure a 3D velocity field in the vaneless space between runner and guide vanes. To get the time-resolved information of the unstable behaviour of the flow instability a complete blade channel was investigated by measuring 21 rotor-stator positions and combining the results with globally averaged velocity fields. The spectral distribution of characteristic frequencies determined by the dynamic pressure measurements was confirmed by the results of the PIV measurements. A comparison of both measurement methods improved the physical explanation of the fluid mechanical mechanisms leading to the S-shaped characteristics of pump turbines.

KEYWORDS Pump turbine, Off-design operation, Instability, S-shape, Experimental investigation

1. INTRODUCTION

A subsidized production of renewable energies such as wind and solar energy usually happens decoupled from the actual demand. Thus, it becomes necessary to store electrical energy produced by volatile sources. Pump storage power plants as multifunctional power plants should provide high flexibility, stabilization of the grid and enhancement of the electricity supply. They provide the only valuable contribution to the efficient storage of primary energies at large scale. Reversible pump turbines offer fast start-up times and fast switching between pumping and generating mode and therefore have substantial advantages in the case of react on the demands of the electrical grid within a short period of time. An improvement of modern pump turbines is the shift of the limitations in the operating range [1]. During the last decade only few research projects described the S-shaped characteristics. Numerical flow simulations of the S-shaped region of the characteristics predicted the emergence of a vortex formation [4]. This vortex formation spreads against the usual flow direction in turbine mode and against the pressure gradient. Besides the normal discharge through the pump turbine model, a secondary flow can be interpreted as local blockage region in the rotational system of the rotor. The secondary flow has a centrifugal acceleration which can cause backflow [3], [5]. Good correlation was found by comparing the flow survey for operation points in the vicinity of the no-load curve with CFD and PIV, which supports the explanation of the mechanisms that lead to potentially unstable turbine characteristics [2]. Widmer et al. [6] attempt to detect rotating stall established at larger GVO where the vaneless space becomes small and the negative pressure gradient through the guide vanes lacks. The present study presents parts of the experimental investigation of the turbine instability of a reduced scale model of a radial pump turbine. After describing the experimental setup, the measured Kcm-Ku-characteristics as well as the operation points investigated in the course of the pressure and PIV-measurements are presented. Detailed analysis of the dynamic pressure measurements in combination with the PIV-measurements performed describes the development and the spread of flow instabilities as well as their flow character. 2. EXPERIMENTAL SETUP

A model of a low specific speed radial pump-turbine was installed in the closed-circuit 4-quadrant test rig at the Graz University of Technology. The speed-regulated machine consists of a runner with 9 blades and a guide vane apparatus with 20 guide vanes. A 500 kW centrifugal pump provides a maximum head of 90 m and a maximum discharge of 0.75 m³s-1

(Fig. 1 (a)). The test rig and the measurement instruments were based on the IEC60193


standard [9] providing measurement accuracy of ± 0.2% even at off-design operation points with low discharge. The full range of the characteristics was investigated from overload to deep part load, runaway and finally to zero discharge at different guide vane openings (GVOs). Decreasing the discharge at fixed GVOs (6° to 15°), the operation of the pump turbine model became unstable in the region of the runaway curve. It suddenly switched to reverse pumping mode by passing the turbine brake mode in a completely autonomous way. In the reverse pumping quadrant the operation stabilised itself without any change in control. By shutting a plunger valve located in the upstream to an almost closed position stable machine operation was achieved even in the S-shape region of the Kcm-Ku-characteristics. In this way the stability of the pump-turbine operation was improved significantly and the operation points within the S-shape could be investigated. As already mentioned, the turbine instabilities which are connected with the appearance of rotor-stator interactions and a backflow region at the inlet area of the runner were predicted by numerical simulations. To analyse the unstable area of the characteristics for typical amplitudes in the frequency spectra, miniature pressure transducers were flush-mounted in the region of the guide vanes at different circumferential angles and radii (Fig. 1 (b)). Additionally, piezo-resistive pressure transmitters were positioned in the draft tube cone and the piping system surrounding of the model in order to register the influence of the test rig and of the feeding pump separately. Signal recording was done at a sample rate of 10 kHz with an A/D resolution of 24 bits and a measurement time of 25 seconds. The miniature pressure transducers are suitable for measurements of hydrodynamic pressure values from low frequencies (0.1 Hz) up to high frequencies (>10 kHz), thus enabling the investigation of the pressure measurements in frequency spectra for dominant frequencies which correspond to the S-shape phenomena.

Fig. 1 (a) Reduced pump turbine model; (b) sensor arrangement in the guide vane channels

In order to explain the origin of the fluid mechanical mechanisms responsible for the appearance of the unstable behaviour two-dimensional ensemble-averaged Particle Image Velocimetry (PIV) was applied to visualize the velocity field in the vaneless space between impeller and guide vanes. An appropriate design of the parts enclosing the pump turbine runner was necessary to integrate this optical measurement method. The modifications of the pump turbine model were realised without changing the geometry according to the IEC60193 standard [9]. The reduced model finally provided access for the measurement of velocity fields in one radial and three axial normal planes (Fig. 2 (a)). The radial optical access for allows the observation of the velocity field in a plane tangential to the outer diameter of the rotor (D=0.35 m).


Fig. 2 (a) Measurement planes; (b) experimental setup for pressure measurements and PIV

Fig. 2 (b) shows the location of the PIV-laser, the light guide arm and the cylinder lens creating a thin laser light section (thickness around 1 mm) for the illumination of tracer particles carried with the fluid flow. At the top of the spiral casing a traversing unit and the camera are located. For each operation point investigated with PIV-measurements 190 single velocity fields were created for one rotor-/ stator-position. Combined with 21 different rotor-/stator-positions for the investigation of one full blade channel, the PIV-measurements at one single operation point result in 190 x 21 = 3990 velocity fields. This way up to five operating points at different guide vane positions were investigated.

3. RESULTS

An overview of the pump turbine operation behaviour including the “S-shape”-region of the characteristics is shown in Fig. 3. The speed and discharge factors are defined by Eq. (1). Every characteristic curve was measured with a constant specific hydraulic machine energy of 100 J kg-1. Similar to the measurement campaign of the PIV-measurements, the pressure fluctuations were recorded at three different GVOs (6°, 8° and 15°).

(1)

Fig. 3 Operation points investigated in the course of the pressure and PIV-measurements


The runaway curve separates the characteristics in turbine mode and turbine brake mode. The region, where the operation of the pump turbine model became unstable, is depicted with dotted lines. With the help of an automatic pressure control a constant level of absolute pressure has been set. This way cavitation-free operation was ensured. Starting with a GVO of 5°, the Kcm-Ku-characteristics exhibit a positive slope below the runaway curve. A stabilisation procedure was necessary to obtain steady-state conditions even in the unstable branch of the characteristics [7].

a. Pressure measurement results

In order to get a global view of the unstable area, the Kcm-Ku-characteristics are superposed with the standard deviation of the pre-filtered pressure fluctuations of the sensors GVO 2/GVO 3 and PS 3/SS 2 according to Fig. 1. The entire characteristics of three GVO were analysed by means of dynamic pressure sensors. The diameter of the superposed circles shown in the following figures is proportional to the standard deviation of the pressure scaled by ρE. The centre of the circle is given by the adjusted operation point on the characteristic curve. The sensors positioned in the area of the guide vane channels measure much higher pressure pulsations than the sensor positioned on the pressure side of the spiral casing. The sensor positioned between the guide vanes and the runner (GVO 2) detects the most important pressure pulsations – up to 5 times higher than the measurement results of the sensor between guide vanes and stay vanes (see Fig. 4 (a)).

Fig. 4 Pressure fluctuations standard deviation within the guide vane apparatus (a) and in the

surrounding of the pump turbine model (b) By comparing the pressure fluctuations in the local best efficiency points with those in the region of the runaway it can be easily observed that the pressure pulsations massively increase at part load. The highest standard deviation of the pressure fluctuations at 6° and 8° can be found in the area of the S-shaped characteristic. In turbine brake mode the value of the standard deviation at a GVO of 6° is about 7 times higher than in the local optimum, at a GVO of 15° even 20 times when refering to the local best efficiency point. The pressure pulsation at a GVO of 15° amounts to 13% of the specific energy. Reaching the zero-discharge the pulsations decrease again. It can be observed that the characteristics are similar for every GVO although the values of the measured standard deviation increase with the GVO. The pressure fluctuations of the sensors installed in the surrounding piping are hardly measurable and therefore presented in a different scale of the drawn circles (Fig. 4 (b)). Only


in the range of the runaway curve pressure fluctuations become noticeable. Especially sensor PS 3, located on the pressure side of the model, detects a higher standard deviation. The pulsations measured with sensor SS 2 – positioned in the suction tube cone, are neglectable. Again, the standard deviation increases with the GVO. An increase in pressure pulsations goes along with enhanced vibrations and amplified noise. The pressure fluctuations in the guide vane channels are far more distinctive than in the surrounding piping of the pump turbine model. Therefore, it can be suggested that the source of the instability can be found in the runner or in the vaneless space between the runner and the guide vanes. This area was chosen for further laser-optical analysis. By closing the plunger valve the usually unstable branch of the characteristic curve could be investigated. Unlike the behaviour with an opened plunger valve, no distinctive interaction with the hydraulic system appeared and the pump turbine model behaved in a stable way. Despite stable operation, high values of the standard deviation occurred. This suggests that the mechanism which is responsible for the instability also occurs in the stiffer system. Further analyses are carried out for GVO 6° and 15° to get information on the frequency components occurring in the pump turbine model. The measured signal amplitudes were scaled by calculating the pressure fluctuation factor pE(t). The non-dimensional factor is calculated by means of Eq. (2).

(2)

A FFT analysis of the presented pressure measurement operation points was performed. The pressure fluctuation factor is plotted as a function of the normalized frequency and the number of the measurement points. Fig. 5 (a) shows a global view of the occurring frequency components for an internal sensor (GVO 2) and the GVO of 6°. The number of the measurement points is defined in Fig. 4 (a). At every measurement point the pressure fluctuation in the guide vane channels is dominated by the blade passing frequency (f = 9·fn).

Fig. 5 Pressure fluctuations amplitude spectra of pressure sensor GVO 2 installed between runner

and guide vanes at 6° (a) and 15° GVO (b)

The amplitudes decrease in the area of the local optimum and increase again to a higher level when operating in turbine brake mode. The amplitudes of the pulsations recorded in the range of f = 0·fn -1·fn in the region of the instability reach less than 0.5% of the amplitude of the pressure fluctuation factor. However, it is not possible to identify a single and clear peak in frequency at this GVO. At a GVO of 15° the amplitudes of the blade passing frequency reach up to 2% of the amplitude of the pressure fluctuation factor. The high standard deviations in turbine brake mode at a GVO of 15° (see Fig. 4 (a)) correspond to a single frequency peak in the waterfall diagram (see Fig. 5 (b)). These findings correspond to the measurement results of Hasmatuchi [8].


The amplitudes at the operation points 16, 17 and 18 raise up to 6% of the amplitude of the pressure fluctuation factor at a frequency of f = 0.65·fn. Compared to the highest amplitudes of the blade passing frequency they are about 3.5 times higher. The spectral analysis of the pressure pulsations shows a low frequency component at a GVO of 15°. At the small GVOs (6° and 8°), as reported by Gentner et. al. [2], this kind of component could not be detected. Fig. 6 (a) presents the time history (see denotation “TH6°” in Fig. 4 (a)) of the pressure fluctuations at a GVO of 6°. A stochastic behaviour of the recorded pressure fluctuations can be detected without any dominant frequency below the local blade passing frequency. Fig. 6 (b) presents the pressure fluctuations at a GVO of 15° (see denotation “TH15°” in Fig. 4 (a)) with a clear phase shift between pressure sensor GVO 1 and GVO 2. The rotor-stator interaction is carried by the low frequency component (~65% of the impeller rotational frequency) at a GVO of 15°. This frequency component is within the range in which rotating stall is expected [6]. Analysing the phase shift of the pressure-time behaviour of the sensors mounted circumferentially at the guide vane apparatus, it can be concluded that one rotating stall cell is detected.

Fig. 6 Guide vane region pressure fluctuation at GVO of 6° OP17 (a) and 15° OP16 (b)

b. PIV-measurement results

The velocity components captured in the course of the PIV-measurements are basically related to the absolute system valid for the pump-turbine model. Fig. 6 shows how the measured velocity components were connected to the velocity triangles. It shows a sketch of the pump-turbine model including the laser light section for the radial view direction. Additionally, a velocity triangle is shown for point P marked in the sketch. Referring to the velocity triangle it is obvious that cu(r) is the velocity component recorded in the course of the PIV-measurement for the radial direction of view.

Fig. 7 Pump turbine model with velocity triangle and measured velocity components


In order to enable a comparison of PIV-results for different operation points, a normalization was carried out by referring the measured velocity to the circumferential velocity u1 at the inlet of the runner. The illustrated velocity is the normalized velocity , see Eq. (3).

(3)

The velocity contours presented in Fig. 8 present the globally averaged PIV-results for eight different operation points marked in Fig. 3. The recorded velocity fields were averaged in order to obtain more valuable results and to separate the stochastical component from the mean value of the flow [5]. The measurement results for the three different guide vane positions 6° and 15° are represented by the two lines. The columns refer to the operation range. While the results in the left column represent the local best efficiency points, the right column contains the results for the turbine brake mode. As expected, the velocity contour plots referring to the local best efficiency points seem widely homogenous – especially as far as it concerns the velocity distribution with respect to the height of the blade channel.

Fig. 8 Globally averaged PIV-results at different guide positions and operation points

A completely different distribution of velocities is shown in the approach to the runaway points represented by the pictures in the second column from the left. The ratio between the measured projection of absolute velocity and circumferential velocity becomes smaller and the velocity distribution appears more inhomogeneous with maximum values that are in a first approximation located in the middle of the pictures. Furthermore, strongly pronounced inhomogeneities are visible in the pictures at the right side of the table which refer to the turbine brake mode of the pump-turbine-model. Compared to the runaway points the mean value of the normalized velocity measured shown in the contour plots increased although the total flow rate significantly decreased. The evaluation of globally averaged velocity contours allow to conclude, that for all investigated guide vane positions an inhomogeneous velocity distribution develops in the blade channel when the runner is operated in strong part load. Additionally, it has to be pointed out that the flow patterns detected at GVO 6° and 15° look slightly different. While the inhomogeneous flow regions extend along the whole blade channel for a GVO of 6°, they are located in the middle of the blade channel for GVO 15°. This finding correlates quite well with the differences between almost closed and widely opened guide vanes detected in the course of the dynamic pressure measurements. Stochastic amplitudes in a frequency range below the local blade passing frequency are detected in part load at GVO 6°. On the contrary to this, a clear peak rising at a frequency component of 0.65·fn can be found for a GVO of 15°.


4. CONCLUSION

The aim of the present study was to improve the understanding of fluid mechanical mechanisms responsible for the occurence of unstable behaviour by investigating a reduced scale model of a radial pump turbine. Dynamic wall pressure sensors installed in the pump turbine model provided the possibility to perform measurements of rotor-stator interactions as well as to investigate the development and the spread of flow instabilities. The main conclusions of the results of the dynamic pressure analysis and the laser optical measurements can be summarized as follows:

• It can be observed, that for every single measurement position the pressure fluctuations increase in the detected unstable area. According to the difference of the pressure fluctuations at small GVOs measured inside the pump turbine model and in the environment it can be supposed that the source of the instability is located in the runner or in the vaneless gap between the runner and the guide vanes.

• Starting at the best efficiency point, the spectral analysis of the pressure pulsations shows a dominant low frequency component at a GVO of 15°, while it is not possible to identify a single and clear peak in frequency at a GVO of 6° or 8°. A stochastic behaviour of the recorded pressure fluctuations can be detected.

• In turbine brake mode of a GVO of 15° a low frequency component (~65% of the impeller rotational frequency) dominates the pressure fluctuations amplitude spectra. It is within the frequency-range in which rotating stall is expected, and by analysing the phase shift of the pressure-time behaviour it can be concluded that one rotating stall cell is detected.

• The results of the PIV experimentally confirm a complex non-uniform flow field in turbine brake mode between the rotor and the guide vanes with high velocities in the centre of the channel between two single blades.

• For the investigated GVO an inhomogeneous velocity distribution develops in the blade channel when the runner is operated in strong part load. The flow patterns detected at GVO 6° extend along the whole blade channel while the inhomogeneous flow regions for GVO 15° are located in the middle of the blade channel.

5. ACKNOWLEDGEMENTS

The present study was carried out within the frame of the EcoVar project (FFG N° 829081), in a partnership with ANDRITZ Hydro. The authors would like to thank ANDRITZ Hydro for their involvement and support as well as Prof. Woisetschläger and the Institute for Thermal Turbomachinery and Machine Dynamics for their support with the laser optical measurements. 6. REFERENCES

[1] C. Gentner, M. Sallaberger, C. Widmer: Progress in pump turbine stability, HYDRO 2013, 2013.

[2] C. Gentner, M. Sallaberger, C. Widmer: Comprehensive experimental and numerical analysis of instability phenomena in pump turbines. 27th IAHR Symposium Hydraulic Machinery and Systems, Montréal, 032046, 2014.


[3] F. Senn, M. Guggenberger, H. Jaberg: PIV als experimentelle Methode für die Erklärung der S-Schlag Instabilität von Pumpturbinen. GALA e.V., München, 2013.

[4] T. Staubli, F. Senn, M. Sallaberger: Instability of Pump-Turbines during Start-up in Turbine Mode, Proceedings of the Hydro 2008, Oct. 6 – 8, 2010, Ljubljana, pp. 6–8, 2008.

[5] M. Guggenberger, F. Senn, J. Schiffer, H. Jaberg, C. Gentner, M. Sallaberger, C. Widmer: Experimental investigation of the turbine instability of a pump-turbine during synchronization, 27th IAHR Symposium Hydraulic Machinery and Systems, 032015, Montréal, 2014.

[6] C. Widmer, T. Staubli, N. Ledergerber: Unstable Characteristics and Rotating Stall in Turbine Brake Operation of Pump-Turbines, J. Fluids Eng., vol. 133, no. 4, p. 041101, 2011.

[7] P. Dörfler, R. N. Pendse, P. Huvet, M. V. Brahme: Stable Operation Achieved on a Single-Stage Reversible Pump-Turbine Showing Instability at No-Load, Proceedings of the 19th IAHR Symposium on Hydraulic Machinery and Cavitation, Singapore, pp. 431–440, 1998.

[8] V. Hasmatuchi, S. Roth, F. Botero, F. Avellan, M. Farhat: High-speed flow visualization in a pump-turbine under off-design operating conditions, IOP Conf. Ser. Earth Environ. Sci., vol. 12, p. 012059, 2010.

[9] IEC 60193: Hydraulic turbines, storage pumps and pump-turbines - Model acceptance tests, 1999.

7. NOMENCLATURE

Kcm (-) discharge factor Ku (-) rotational speed factor n (rpm) runner speed D (m) runner outlet diameter E (m2.s-2) specific energy ρ (kg.m-3) Water density f (s-1) frequency fn (s-1) runner speed GVO (°) guide vane opening u1 (m.s-1) circumferential velocity

p(t) (Pa) time-based pressure (Pa) average time-based pressure

(-) pressure fluctuation factor (m.s-1) mean velocity

(-) normalized mean velocity c (m.s-1) absolute velocity

(m.s-1) circumferential component (m.s-1) meridional component

u (m.s-1) circumferential velocity w (m.s-1) relative velocity

investigating the dynamic aspects of the...

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