New design of butterfly valves for modern pump-storage power plants Gerhart Penninger Voith Siemens Hydro Power Generation GmbH&Co. KG Engineering & Design – TUK Linzer Strasse 55 A- 3100 St Pölten mailto:
[email protected]
Helmut Benigni Institute for Hydraulic Fluidmachinery, Technical University of Graz Kopernikusgasse 24/IV A-8010 Graz mailto:
[email protected]
Introduction The present paper discusses the design of emergency-closing butterfly valves in tandem-arrangement for modern pump-storage power plants. For newer power plant projects it is characteristic that the gallery rises in flow-direction of the turbine, whereby the valve chamber, which is usually situated at the passage from the gallery to the penstock, for draining the penstock, is no longer necessary. To determine the forces in this tandem-arrangement, a model test on a scale of 1:21.5 was chosen on the 4-quadrant-test-rig of the TU Graz. At the same time, 3-dimensional numeric CFD-simulations were realized, which allowed detailed insights into the geometry of the valves and modern interpretations of the forces for the structure-mechanical calculation. In order to guarantee a high safety standard, two similar shut-off valves are placed in the valve chamber, which have to be suitable for emergency closing independently from each other. As it may happen that the shut-off valve in the emergency closing fails, the second valve has to be operated independently and should be able to close. In order to reduce construction costs of the valve chamber, it is necessary to place the valves as close in series as possible (in this case 1.05D) without having an impact on each other. Single butterfly valves were tested very accurately 30 years ago, but there exist only a few publications on valves in tandem-arrangement (Keller, Morris) and these publications only discuss lens-type and no lattice-type. The purchaser of the valves also demanded, that the valves can close safely in pump mode in case of failure. The rig-program was designed for a matrix of 121 positions, which included both pump mode and turbine mode and thus the interaction of the tandem was verified as well as all coefficients of torque, force and flow rate were determined. The measurements were made by means of pressure gauges and DMS full bridges, whereas the primary pressure was varied for every working pressure in order to verify the influence of cavitation as well. The CFD-models were built automatically and, starting from a built CAD-model, gridded, automatically calculated and analized, whereas the calculations were made on the UNIX parallel cluster of the TU-Graz with Ansys CFX 5.7. Most of the simulations were made free of cavitation, for cavitation-afflicted simulations the Rayleigh-Plesset-model was used. The free of cavitation situation as philosophy of design was approved by the measurements. By means of measurements and accompanying numeric simulation a loss-optimized geometry was built and executed as well. In order to get values comparable to the values from literature, the nomenclature of the dissertation of Grein was used and thus a comparison was made with various butterfly valves measured at the TU Graz and Voith-Siemens Hydro. At the end of the paper the constructive highlights are presented.
1. Background Plug-off valves are absolutely necessary in power plant configurations for maintenance and emergency-closing of the waterway. Often, spherical valves, needle-valves or even butterfly-valves – flat plates rotating through a horizontal axis – are used. Butterfly-valves are space-saving and cheap apparatus for shut-off, in our configuration also an emergency shut-off and a revision possibility are needed and not only a valve for the operating procedure. The rotating axis of the butterfly valve is not only normal to the pipe axis, an eccentric offset is a feature as well. The main functions are the ‘opening’ and the ‘shut-off’ in case the pipe bursts (see Grein). When this happens, a multiple of the normal flow rate is possible. In reality the largest diameter was over 7.3 m and pressures of 21 bar were realized (Flury). State-of-the-art in design process is the structure mechanical simulation of this sensible part. In part, CFDcalculations were carried out.
In our case the two butterfly-valves mounted in series (distance between 4500mm) and the interaction between the two valves should be estimated and an optimized configuration should be realized together with the manufacturer of the butterfly-valve. The plant owner is interested in getting information on the fundament (forces) and the diameter needed for the piping and thus the amount of the break-out material. 1.1 The Prototype The prototype for which these studies and research activities were done has following main data: D = 4300mm p = 160mWc e = 65mm
Qturbine direction (emergency-closing) = 200m3/sec Qturbine direction (normal operating) = 146m3/sec Qpump direction = 118m3/sec
The drive of the valve must be able to close under all conditions, independent from the direction and the amount of discharge and the position of the other valve. The counterweight is situated on one side of the valve and must be able to keep the valve closed up to a maximum water level at the downstream side. 1.2 The test rig The following experimental setup was carried out on the four-quadrant cavition test rig of the institute, which is a closed circuit. The task was the estimation of torque, force and discharge coefficient at different cavitation situations for eleven positions for each valve and also to get the necessary information on forces for the intermediate pipe and the dismantling pipe. So 121 combinations for the turbine mode and an equal amount of combinations for the pump mode were tested and interactions between the two valves were investigated. The pressure was measured 5*D before and 10*D after the valve housing, torque and moment with strain gauges full bridges (see Fig. 2). These strain gauges were gluing on the bending and on the torsion rod inside the shaft sealing, because no reaction force and torque of this should be measured. In order to prevent cavitation for the first measuring points, the system runs with over-pressure. The calibration was carried out with weights on a 0.5m lever arm, stepwise to a maximum torque of 200 Nm to a linear correlation between moment and electrical signal. Head losses in front of and after the valve were estimated by means of an extra test and therefore allow for the results.
Plane 2 Pressure measurement transformer (Rosemount) Valve housing (tandemconfiguration)
Plane 1
Inductive flow rate measuring device CopaXM 200 Carrier frequency measuring amplifier
Flowdirection in pump-mode and turbine-mode
Fig. 1: Test rig
For cavitation free operation (σ > 2) the Reynolds number varied from Re ≈ 2.04*106 at α = 0° to Re ≈ 0.25*106 at α = 80°, as the flow through the system decreases rapidly. For strong cavitation the Reynolds number varied from Re ≈ 1.85*106 at α = 0° to Re ≈ 0.21*106 at α = 80° very linear. The reproducibility of the measurements gets a distribution of kQ = 0.732%, kF1 = 1.017%, kM1 = 1.782%, kF2 = 1.98%, kM2 = 1.623%.
Fig. 2: Torque und force measurement with strain gauges full bridges
1.3 The numerical simulation model As the test rig model, the numerical simulation model was scaled with the factor 21.5. The model was split in parts with its block boundaries. The flow direction is from the positive to the negative x-direction and follows through the four subdivisions: - Inflow-pipe, length = 2 m, diameter = 0.2 m (structured) - Valve 1, length = 0.209 m, diameter = 0.2 (unstructured) - Valve 2, length = 0.209 m, diameter = 0.2 (unstructured) - Outflow-pipe, length = 2 m, diameter = 0.2 m (structured) Grid generation works scripted and automized pre-phase, solve- and post-processing phase. From paper published geometry a 3D-Pro-engineer®-Modell was generated, continued with the grid generation in CFXBuilt 5.6. The pipegrid was generated with TASCgrid. In order to solve the Navier-Stokes-equations, Ansys CFX5.7 SP1 was used to calculate the approx. 4 mio. node-models on the UNIX-Cluster of the TU Graz in parallel. In the beginning, a short evaluation was carried out with different turbulent models to use the same model and boundary conditions for the whole calculation matrix. Finally the SST-model (Menter) was used, also the k-epsilon-model, the k-omega-model and the Reynolds-stress-model. With the local time step factor 10 the best convergence at different test runs was found. All simulations were carried out with the same flow rate and isotherm for all positions. After the measurements and the calculations with a common design the disadvantages were pointed out and a optimization run with CFD was done to get a better geometry with lower head losses and a better torque characteristic during closing in pump mode (change from closing tendency to opening tendency at greater angles).
2. Results of the simulated test rig model In Fig. 4, left side, the swirl in the lower part, which flows through lesser, in the second valve is shown when the first valve is already significantly closed (50°). In this simulation a 360° full-model is used (no symmetry boundary in the middle), so a weak asymmetric solution is the result, which corresponds to reality (see Grein), however one power lower than the main-force- direction. Numerically it is possible that such effects reinforce a numerical tilting – such effects are very distinctive at draft tubes of bulb-turbines (Gehrer). The differences between numerical simulation and test rig (see Fig. 4) are primaryly based on turbulence modelling and the too early or wrong detachment from the flow and also on the numerical viscosity. In addition to the central method of equations, the commercial numerical solver uses friction terms of the second and fourth power. Those act like a physical friction tensor and are complicated location functions, which are related not only to grid quality but also to flow distribution. These apply to the gridface under 90 degrees – not very positive for very oblique-angled grids. In Fig. 3, right side, planes are shown with cp, quite similar to figures from Grein and the values published by him. Some of the dimensional coefficients for the operation and revision (emergency) valve for pump and turbine mode are shown in Fig. 4 and Fig. 5. With the same colour each numerical and test plate data result lined. There is a very good correlation between CFD and experimental results for each position of the first valve and variation of the second valve. The influence of the fixed valves raises to higher angles up to approx. 70 degrees. After this, the first valve is dominating. In pump mode there is the same situation. For the fully open situation (blue line for zero degree) there is a small offset to be seen as result of several weeks operation of the test plate – a friction coefficient displacement gets more and more important, but only with effects in this operation point. In the literature (cp. Idelchik or Miller) zeta instead of kQ is shown, but for near fully closed situations the tendency to infinite is a disadvantage for this illustration method.
Fig. 3: Left view into flow distribution – right cp on different planes (turbine mode)
In Fig. 5 km2 of the test-plate and the CFD-configuration are shown – in addition also a single valve CFD-calculation is outlined in grey as an envelope of the tandem valve configuration. All dimensionless torques are lower than this envelope. At fully open position (both valves in 0°) nearly no torque is acting on the valves. When the revision valve is getting closed, there is an opening torque acting on the operation valve until α1 = 25° is reached by the revision valve. This also happens for operation valve positions less than 7°. For the variation of the operation valve a closing tendency for the revision valve up to 50° is outlined (over 50°: opening tendency). 0.9 Valve1=00°-KQ-Num.Sim.Test-plate Valve1=10°-KQ-Num.Sim.Test-plate Valve1=20°-KQ-Num.Sim.Test-plate Valve1=30°-KQ-Num.Sim.Test-plate Valve1=40°-KQ-Num.Sim.Test-plate Valve1=60°-KQ-Num.Sim.Test-plate Valve1=80°-KQ-Num.Sim.Test-plate Valve1=00°-KQ-Test-plate Valve1=15°-KQ-Test-plate Valve1=25°-KQ-Test-plate Valve1=40°-KQ-Test-plate Valve1=60°-KQ-Test-plate Valve1=80°-KQ-Test-plate
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Fig. 5: kM2-turbine mode, variation of the revision valve
When 70° of the revision valve are reached, an opening tendency of the operation valve is in force – this means that even if you shut-off the valve an opening tendency is the result of the flow distribution. But, because of the nearly closed valve, a low flow distribution is the result and thus also very small opening moments.
3. Optimized geometry results 0.13
turb-Kl.1=00°-KM2-Num.Sim.Real.Conf. turb-Kl.2=00°-KM2-Num.Sim.Real.Conf. turb-Kl.1=00°-KM2-Num.Sim.Test-plate turb-Kl.2=00°-KM2-Num.Sim.Test-plate turb-Kl.1=00°-KM2-Test-plate turb-Kl.2=00°-KM2-Test-plate pump-Kl.1=00°-KM2-Num.Sim.Real.Conf. pump-Kl.2=00°-KM2-Num.Sim.Real.Conf. pump-Kl.1=00°-KM2-Num.Sim.Test-plate pump-Kl.2=00°-KM2-Num.Sim.Test-plate pump-Kl.1=00°-KM2-Test-plate pump-Kl.2=00°-KM2-Test-plate
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Fig. 6: kQ-comparison – test plate simulation with real design simulation
This configuration reaches better values in view of kQ. For a conversion into zeta an equal segmentation has to be done for an estimated value 0.93 – this with no backward-mounted sealing. This procedure gets its confirmation with the calculation of a single valve configuration. More benefits are reached with this design with a different eccentricity and valve form for the ‘closing-law’ and its torque distribution reaching the significant shut-off phase after 50° in pump mode (see Fig. 6). The numerical simulation was carried out with constant mass flow rate.
10000 9000 8000 7000 6000 5000 4000 3000 2000 1000 0 -1000 -2000 -3000 -4000 -5000
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Demounting pipe –- Force in y-direction Model [N], Prototype [kN]
Pipe between the valves - Force in y-direction Model [N], Prototype [kN]
For a comparison the results of the post processing force have to be converted with real flow rates. This can be done with the Bernoulli and the use of a position-free constant. For a detailed force analysis you can find a very good comparison also with the complementary valve positions. The dismantling pipe (after the second valve in turbine mode) has a length of 1.98*D. At the test plate only a rough distribution of the measurement pressure holes is possible – in the numerical simulation a sensible distribution of this holes is a normal feature. These forces of the numerical simulation are very similar to the results of the test plate and to Grein (see Fig. 7). In this diagram the results of the dismantling pipe and the pipe in between the valves are quite similar in a wide range, only between 025° they are different. After 50° the flow rate decreases very strongly – also the forces go down in reality, but in the numerical simulation the flow rate was constant and so this method of conversion is needed.
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Pipe between Prototype recalc. Valve2=0° Valve1=variable Forces Grein Pipe between
Pipe between Num.Sim. Model Valve2=0° Val ve1=variabel Forces Grein Demounting pipe
Demounting pipe Prototype recalc. Valve1=0° valve2=variable Demounting pipe Prototype recalc. Valve2=0° valve1=variable
Demounting pipe Num.Sim. Model Valve 1=0° Valve2=variabel Demounting pipe Num.Sim. Model Valve 2=0° Valve2=variabel
Fig. 7: Force estimation on the prototype
4. Cavitation The questions of different length and time scales is the most import to solve in order to catch the problem of right cavitation (eg. collaps of bubble approx. 10 E-5 sec., flow through radial pump approx. 1 sec.) and follows two different proceedings: Interface-tracking methods and continuum methods which are used in CFX (also ‘two-phase flow schemes’ or ‘interface capturing’). Cavitation is characterized by the cavitation coefficient by Thoman and implemented in CFX with the help of the Rayleigh-Plesset-model. Different positions were calculated with the model and the fixed bubble diameter was 1 µm. The results seem to be very plausible because of the areas of beginning cavitation (see Fig. 8). Very significant results at the seal in the housing and around both sides of the trunnions are shown. In this project there was no effort to reduce cavitation but it was very amazing to see the comparison between the model test and the numerical simulation, not only the areas (surfaces) of beginning cavitation, also the area of cavitation in the flow.
5. Details of the prototype’s construction
Fig. 8: Cavitation in the valve
The counterweight was constructed in a way, that if there is water on the downstream side with a maximum pressure of 142mWc the valve must keep closing with a safety factor of minimum 1.5. This loadcase demands a counterweight of about 36.5 tons at a radius of 2m, the client and the supplier only wanted one counterweight per valve because of synchronism problems and the high stresses at the trunnions in that case. Another stipulation of the customer was, that the valve has to be able to close in each position and in both direction of the flow. The highest torque with opening tendency resulting of the flow in pump direction is 712kNm. The friction torque in this point is 805kNm, so 774kNm out of the servomotor are necessary. The system pressure of the hydraulic unit is 150bar.
During an emergency-closing of 200m3/sec, 6700kNm have been braked in the servomotor and the client limited the brake pressure with 200bar. All these boundary conditions defined the servomotor with a rod diameter of 480mm and a piston diameter of 580mm. The supplier decided to construct the valve with a one-piece housing and therefore with screwed on trunnions. The diameter of the trunnions is 900mm, the whole piece manufactured out of stainless steel. To reach the stiffness at the connection trunnion-rotor 12 bolts M90 are necessary.
6. Conclusion The results of the measurements and the calculations were in excellent correlation. Modern simulations get into power-plant shut-off valve design very slowly, old conservative design methods are dominant, such as the test rig. With the support of CFD an optimized design was estimated and shows modern design-methods for this part of mechanical engineering as well. As a special feature of the construction the sealing was analized more in detail and loss-analysis was realized. The optimized geometry has its advantages in the ‘closing law’ and better mechanical calculations in the piping-system around the valve could be carried out with the additional information from the CFD-calculations.
Short nomenclature The valve-head Hkl is no real total energy head, it is a calculation value which is equivalent to the difference of total energy head from the valve added with the velocity energy head. Also for coefficients the Hkl over both valves is used (only the pipe losses were correcting), in difference to other publications (Ziegler). w 2 ( p1 − p2 )Rohr ⎛p −p ⎞ w 2 , where hVR = ζR = HKl = ⎜ 1 2 − hVR ⎟ + 2⋅g ρ⋅g ⎝ ρ⋅g ⎠ 2⋅g Q kQ = A ⋅ 2 ⋅ g ⋅ HKL pII −px } 2 ⋅ ∆p Pressure coefficient c p(StelleX) = ρ ⋅ w2 M k 8⋅ρ Torque coefficient: k = and M = M2 ⋅ ⋅ Q2 M D3 ⋅ ρ ⋅ g ⋅ HKL k Q D ⋅ π2
Force coefficient: k = F
F and F = kF ⋅ 8 ⋅ ρ ⋅ Q2 D ⋅ ρ ⋅ g ⋅ HKL k Q 2 D2 ⋅ π 2 2
turbine mode
F n δ1
D
F e
α2 Valve 2 Operation valve
Ft
w
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Valve 1 Revision valve
2
1 Fig. 9: Nomenclature
References 1. 2. 3. 4. 5. 6. 7. 8. 9.
Illwerke, „Kopswerk II – Die Energiezukunft unserer Wasserkraft“, Powerpoint Präsentation, http://www.illwerke.at/Downloads/KopsII/KOW-IIFolder.pdf, 15. Okt. 2004. Grein, H., „Experimentelle Untersuchungen der Kräfte und Momente auf das System Drosselklappe – Rohrleitung in Wasserkraftanlagen“, Dissertation Technische Hochschule Darmstadt, 1977. Keller C., Vuskovic, I., „Strömungsversuche an Sicherheitsorganen von Wasserkraftanlagen“, 100 Jahre Turbinenbau, Escher Wyss Mitteilungen, S. 91- 96, 1942/1943. Morris, J.M., „The performance of two butterfly valves mounted in series”, ASME Journal of Fluids Engineering, Vol. 107, S. 101-105, 1985. Strohmer, F., „Investigating the characteristics of shut-off valves by model tests“, Water Power & Dam Construction, July 77, S. 41-46, 1977. Flury, H., „Wasserkraft vom Himalaya“, aus: Technische Rundschau Sulzer, 4/1990, S. 28-35, 1990. Menter, F.R., „Two-equation eddy-viscosity turbulence models for engineering applications”, AIAA-Journal, 1994. Gehrer, A., Benigni, H., Köstenberger, M., „Unsteady Simulation of the Flow Through a Horizontal-Shaft Bulb Turbine”, 22nd IAHR Symposium on Hydraulic Machinery and Systems, Stockholm, 2004. Ziegler, G., „Drosselklappen bei kleinen Sigma-Werten“, Interne Veröffentl. TU Graz, 1990.
The Authors Helmut Benigni, Dr. Dipl.-Ing., Studies of mechanical engineering at the Technical University of Graz, specialization in numerical simulation, dissertation in optimization of hydraulic machines, nowadays Institute for Hydraulic Fluidmachinery, CFD-group Gerhart Penninger, Dr. Dipl.-Ing., Studies of mechanical engineering at the Technical University of Graz, specialization in measurements at hydraulic machines, dissertation in blade vibrations of axial flow pumps, nowadays Voith Siemens, department of Hydraulic Technology Engineering and Design, St. Pölten.
Benigni, Penninger 2005
New design of butterfly valves for modern pump-storage power plants HYDRO 2005 International Conference and Exhibition Policy into Practice
Villach, Austria, 17-20 October 2005
Gerhart Penninger Voith Siemens Hydro Power Generation GmbH&Co. KG, St. Poelten
[email protected]
Helmut Benigni Institute for Hydraulic Fluidmachinery, Technical University of Graz
[email protected]
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The power plant project
Main data: D = 4300mm
Pturbine: approx 450MW
p = 160mWc
Ppump: approx 450MW
e = 65mm Qturbine direction (emergency-closing) = 200m3/sec Qturbine direction (normal operating) = 146m3/sec Soure: http://www.illwerke.at
Qpump direction = 118m3/sec
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Test rig procedere 4-quadrant cavitation test rig, scale factor 21.5 121 combinations for pump and turbine mode Torque and force measurement with strain gauges full bridges
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The numerical simulation model
Scale factor 21.5 Scripted grid generation Automized pre-phase, solveand post-processing phase SST-model (Menter) UNIX-Cluster in parallel
Software: pro-engineer, Ansys CFX 5, CFXbuilt tascgrid Structure (total approx 4 Mio. nodes): - Inflow-pipe, length = 2 m, diam. = 0.2 m (structured) - Valve 1, length = 0.209 m, diam. = 0.2 (unstructured) - Valve 2, length = 0.209 m, diam. = 0.2 (unstructured) - Outflow-pipe, length = 2 m, diam. = 0.2 m (structured)
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A view into the flow distribution
1st valve: significant closed pos. Full 360° - model, weak asymetric solution (compare to Grein)
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A view into the flow distribution 2
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Test plate results vs. numerical simulation: KQ (zeta) 0.9 Valve1=00°-KQ-Num.Sim.Test-plate Valve1=10°-KQ-Num.Sim.Test-plate Valve1=20°-KQ-Num.Sim.Test-plate Valve1=30°-KQ-Num.Sim.Test-plate Valve1=40°-KQ-Num.Sim.Test-plate Valve1=60°-KQ-Num.Sim.Test-plate Valve1=80°-KQ-Num.Sim.Test-plate Valve1=00°-KQ-Test-plate Valve1=15°-KQ-Test-plate Valve1=25°-KQ-Test-plate Valve1=40°-KQ-Test-plate Valve1=60°-KQ-Test-plate Valve1=80°-KQ-Test-plate
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KQ figured out with zeta kQ =
⎛p −p ⎞ w2 HKl = ⎜ 1 2 − hVR ⎟ + ⎝ ρ⋅g ⎠ 2⋅g
Q A ⋅ 2 ⋅ g ⋅ HKL
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Test plate results vs. numerical simulation (KM) 0.16
Valve1=00°-KM2-Num.Sim.Test-plate Valve1=10°-KM2-Num.Sim.Test-plate Valve1=20°-KM2-Num.Sim.Test-plate Valve1=30°-KM2-Num.Sim.Test-plate Valve1=40°-KM2-Num.Sim.Test-plate Valve1=60°-KM2-Num.Sim.Test-plate Valve1=80°-KM2-Num.Sim.Test-plate Valve1=00°-KM2-Test-plate Valve1=15°-KM2-Test-plate Valve1=25°-KM2-Test-plate Valve1=40°-KM2-Test-plate Valve1=60°-KM2-Test-plate Valve1=80°-KM2-Test-plate
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0.09 0.08 0.07 0.06
Valve1=05°-KM2-Num.Sim.Test-plate Valve1=15°-KM2-Num.Sim.Test-plate Valve1=25°-KM2-Num.Sim.Test-plate Valve1=35°-KM2-Num.Sim.Test-plate Valve1=50°-KM2-Num.Sim.Test-plate Valve1=70°-KM2-Num.Sim.Test-plate Valve1=85°-KM2-Num.Sim.Test-plate Valve1=10°-KM2-Test-plate Valve1=20°-KM2-Test-plate Valve1=30°-KM2-Test-plate Valve1=50°-KM2-Test-plate Valve1=70°-KM2-Test-plate SingleValveConfiguration
0.05 0.04 0.03 0.02 0.01 0 -0.01 -0.02 -0.03 0
10
20
30
Turbine-mode: tandem-configuration
40
50
Valve-angle
60 70 80 90 Valve1=Revision-valve | Valve2=Operation-valve
BUTTERFLY VALVES
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Benigni, Penninger 2005
Modification and KM 0.13
turb-Kl.1=00°-KM2-Num.Sim.Real.Conf. turb-Kl.2=00°-KM2-Num.Sim.Real.Conf. turb-Kl.1=00°-KM2-Num.Sim.Test-plate turb-Kl.2=00°-KM2-Num.Sim.Test-plate turb-Kl.1=00°-KM2-Test-plate turb-Kl.2=00°-KM2-Test-plate pump-Kl.1=00°-KM2-Num.Sim.Real.Conf. pump-Kl.2=00°-KM2-Num.Sim.Real.Conf. pump-Kl.1=00°-KM2-Num.Sim.Test-plate pump-Kl.2=00°-KM2-Num.Sim.Test-plate pump-Kl.1=00°-KM2-Test-plate pump-Kl.2=00°-KM2-Test-plate
0.12 0.11 0.1 0.09 0.08 0.07
KM2
0.06 0.05 0.04 0.03 0.02 0.01 0 -0.01 -0.02 -0.03 -0.04 0
10
20
KM2: pump- and turbine-mode
30
40
50
Valve-angle
BUTTERFLY VALVES
60 Valve No.1: 70 in each mode 80 the first valve 90 pump-mode: valve 1 = Operation-valve turbine-mode: valve 1= Revisions-valve
10
Benigni, Penninger 2005
Cp on dismouting pipe and pipe between the valves
Force estimation
%] X[
BUTTERFLY VALVES
11
Benigni, Penninger 2005
Force [N]
Force post processing for all different turbine modes
Length [%]
BUTTERFLY VALVES
Dismounting pipe, turbine mode 12
Benigni, Penninger 2005
Comparison with Grein
BUTTERFLY VALVES
13
Benigni, Penninger 2005
Cavitation
Rayleigh-Plesset-model. Fixed bubble diameter 1 µm. The results seem to be very plausible because of the areas of beginning cavitation Very significant results at the seal in the housing and around both sides of the trunnions
BUTTERFLY VALVES
14
Benigni, Penninger 2005
Prototype (D = 4,3 m / p = 16 bar) housing (33 to)
lock counterweights (36,5 to)
lock in case of disass embled counter weight
servomotor with bases
rotor (49 to) base for housing
BUTTERFLY VALVES
15
Benigni, Penninger 2005
fla in ma
on nni tru
el) ste s s inle (sta
e ev sle nd
ea ar r e bb u r
a pin
ro to rs to p
After priming (Sept. 2005)
es ng
All for corrosion protection problematic areas are covered with stainless steel
BUTTERFLY VALVES
16
Benigni, Penninger 2005
Assembly for pressure test (Sept. 2005)
BUTTERFLY VALVES
17
Benigni, Penninger 2005
Thank you for your advance ...
http://www.hfm.tugraz.at | http://www.vs-hydro.com
BUTTERFLY VALVES
18
