th
15 International Symposium on Transport Phenomena and Dynamics of Rotating Machinery, ISROMAC-15 February 24 - 28, 2014, Honolulu, HI, USA
Numerical Investigation on Under Expanded Wall Jet Separation in a Steam Turbine Valve Diffuser 1
1
1
Domnick, C.B. *, Benra F.-K. , Dohmen H.J. , Musch C.
2
1
Department of Turbomachinery University of Duisburg-Essen Forsthausweg 2, Duisburg , 47058, Germany
[email protected] 2
Siemens AG Rheinstraße 100, Mülheim an der Ruhr, 45478, Germany
[email protected]
Abstract In steam turbine inlet valves sudden changes in the vibrational level are observed. The CFD calculations which are presented in this paper reveal that these changes are related to a change in the flow topology. In order to reduce the computational amount simplifications of the CFD Model are evaluated and applied. It is shown that the change of flow topology is related to the separation of supersonic Coanda wall jets. Based on this finding a prediction tool is developed which enables a fast prediction of flow separations. Nomenclature DSeat h p pmax pmin pout pSeat pt pt,in p∞ p� q r R u uis x ν ξ
Seat diameter of the valve Height of the gap Pressure Maximum of dynamic pressure Minimum of dynamic pressure Outlet pressure of the valve Pressure at the valve seat Total Pressure Total inlet pressure Ambient pressure Non dimensional pressure fluctuation Non dimensional mass flow rate Radius Radius of the valve seat Velocity Isentropic velocity Distance Kinematic viscosity Non dimensional distance
m m bar bar bar bar bar bar bar bar [-] [-] m m m/s m/s m m²/s [-]
π πSeat πloc πValve ρ
Pressure ratio Pressure ratio at valve seat Local pressure ratio Pressure ratio of the entire valve Density
[-] [-] [-] [-] [kg/m³]
Introduction Steam turbine control valves are used in power stations to adjust the power output of the steam turbine to the electric power demand. By throttling the valve, the turbine inlet pressure and the mass flow rate are reduced and hence the power is decreased by a large amount. If a small turbine power output is demanded, the pressure difference of the valve is large and can even exceed 100bar at certain operation points. At such severe operation conditions, high vibrational stresses can arise. Zaryankin [15] investigated the failure rates of turbine inlet valves in Russia. The reliability characteristic Zaryankin developed shows that the frequency of failure in cogeneration plants is two to three times higher than the failure rate in base load power plants. In valves used for turbines in industrial plants, in which the power is throttled frequently, the failure rate is even several times higher. Due to the increasing feed of fluctuating renewable energy into the power grid, fossil power stations will have to throttle the power output more often. So higher vibrational loads are expected and methods for avoiding vibrational problems have to be developed. Therefor a better understanding of the physical effects causing the vibration will be nec-
1
essary to ensure a reliable operation of steam turbine valves designed for future needs. Literature Review Several papers concerning valve vibrations have been published. The following overview shows that several mechanisms exist which can cause valve vibration. In two cases it is reported that acoustic modes are excited. Ziada [17] discovered an acoustic mode in a turbine control valve that excited by flow impingement and led to unwanted high sound levels in the control valve. A similar mechanism has been found by Nakano [11] in a pressure reducing gas valve. In this case an oblique shock was additionally excited to an unsteady movement and led to high dynamic forces. Also normal shocks can cause severe vibrational problems in turbine inlet valves. Stanstny [13] investigated a control valve for a nuclear turbine. In this vale a convergent-divergent gap is formed between the valve plug and the seat. In this gap an oscillating shock is observed under certain conditions. Furthermore Darwish [3] reports on valve vibration excited by vortex shedding in check valves. In this case, a coupling between the torsional mode of the shaft and the frequency of the vortex shedding appeared. Other researchers e.g. Zhang [16], Morita [10], Engeda [4] and Heymann [6] found flow separations being a cause for undesired high vibrations. Morita [10], who focused on mushroom-shaped valves which are used in some steam turbines, found rotating pressure fields beneath the valve which are likely to cause the valve instability. The investigations of Zhang [16] showed that asymmetric flow pattern beneath the valve can occur and may cause vibration. Engeda [4] carried on the research on an optimized plug geometry proposed by Zhang [16]. He found different flow patterns, which are related to certain ranges of valve lift and pressure ratio. Certain flow topologies showed higher pressure oscillations than other flow topologies. In the investigation of Heymann [6] two basic flow topologies were observed, which are called annular flow and core flow. If annular flow occurs, the jet formed by the valve plug and valve seat is attached to the seat and the noise caused by the valve is low. If core flow exists, the jet is detached from the seat
and a single free jet in the core of the diffuser is formed. In this case the sound emission is significantly higher. Also Heymann found out that the pressure ratio at which the transition from core flow to annular flow takes place is depending on the valve lift. A similar change of flow topology occurs in the valve which is discussed in this paper. Overview The previous section shows that a broad range of effects causing valve vibrations exist. This paper focuses on wall jet separation which occurs in the valve investigated here. The research of Heymann [6] shows that the detached flow causes higher pressure fluctuations but the reason for these flow separations could not be revealed at that time. A similar behavior was observed in test rig measurements and data records from power plants. Based on 3D and 2D CFD calculations the physical effect of these separations will be explained in this paper and results of a prediction tool based on these findings will be shown. Design of the investigated valve In this paper a characteristic design of a steam valve as installed in large power stations is investigated. If it is operated in the admissible range of design mass flow, changes in vibration caused by flow separation can be clearly measured but the vibrational level does not exceed the acceptable level. A picture of the flow domain of the valve is given in Fig.1. The flow coming from the live steam line enters the valve at the inlet (A). After that it flows through a steam strainer (B) which is located around the emergency stop valve. Then it passes by the emergency stop plug (C) which can either be completely opened or completely closed. After passing the first diffuser (D) the steam enters the annular chamber (E), in which the flow is distributed uniformly. In the annular chamber a rib (H) is installed in order to avoid the formation of a vortex. Additionally four flow straighteners (G) are placed directly in front of the control valve plug (F). The valve plug, which is used to throttle the flow, is can-shaped in order to damp pressure fluctuations. Downstream the valve seat (I) the second diffuser (J) is installed in order to reduce pressure losses if the valve is completely opened. The outlet (K) of the valve is directly connected to the steam 2
turbine. The Investigation is focused on the control cone as this is used to throttle the flow.
ratio (equ.2) is the quotient of the static valve outlet pressure divided by the total inlet pressure.
B A
p� = C D E F
ΠValve =
K
pout pt,in
(2)
0,15
I p [-]
0,125
Fig. 1: Sketch of the Valve
Measured data Two prior experiments concerning flow separations and valve vibrations were conducted at the Siemens test facilities by Krämer [7]. The results of these experiments will be presented in a short manner because they are used to confirm the accuracy of the prediction tool. pin Flow meter
(1)
A sudden change of dynamic pressure can be seen when the pressure ratio is increased.
G J
H
pmax − pmin pt,in
Model valve
Ambient air
Bypass valve pout
0,025 0 0
0,1
0,2
0,3 0,4 π Valve [-]
0,5
0,6
Fig 3: Dynamic pressure in model valve
A full size experiment is carried out using a valve pulled from the production. The valve is operated with high pressure air supplied form a pressure vessel. Also the effect of flow separations is observed. The vibrational level of the plug is measured by an acceleration sensor. The acceleration is plotted versus the pressure ratio in Fig.4. It drops by 76% abruptly at a lift of 8.3% and a pressure ratio of 0.22. 1
Normalized acceleration [-]
In the first test rig air is sucked through a model valve by a vacuum pump. The model is scaled by a geometric factor of 1:2. A sketch of the test rig is shown in Fig 2. The valve outlet pressure is adjusted by a bypass valve. Dynamic pressure probes are installed at the valve plug and at the valve in order to detect flow separations. In Fig. 3 the peak-to-peak amplitude of the dynamic pressure at the valve cone is plotted versus the pressure ratio of the entire valve. The lift of the valve remains constant while the outlet pressure is increased. A non-dimensional representation of the amplitude is obtained by dividing the amplitude by the total pressure at the inlet (Eq. 1). The pressure
0,075 0,05
Vacuum pump
Fig.2: Model test rig
0,1
0,8 0,6 0,4 0,2 0 0
0,1
0,2
0,3
π Valve [-]
0,4
0,5
0,6
Fig.4: Vibrational acceleration in full size experiment
Also operational records of power stations show a transition in the vibrational level. In Fig.5 a record showing the pressure ratio, the valve lift and the normalized vibrational acceleration is shown. While the lift and the pressure ratio increase 3
smoothly, a sudden drop in the vibrational level is reported. 1
20
Transition point
0,6
15
0,4
10
0,2
Vibrational acceleration Pressure ratio
Valve lift [%]
Pressure ratio [-]
0,8
25
5
Valve lift
0
0 Operation time
Fig 5: Data record of a valve applied in a power station
An overview of the measurements that where conducted is given in table 1. The three measured transition points differ although they were obtained on similar geometries. The reason behind this is that the valve’s operational curves differ in the two experiments and in the power plant. It is shown in the next sections that these changes in vibrational level are due to the attachment of a separated wall jet. Case Suction test High Pressure test Power station
Lift [%] 16,3 8,3 18
Pressure Ratio 0,42 0,22 0,48
Table 1: Measured transition points Numerical investigations In order to find out the reason for this behavior, several numeric investigations were performed using the commercial CFD code Ansys® CFX release12.1. In order to reduce the computational effort, three types of models are used: • Full CFD model • 90°-segment of the diffuser • 2D- models According to the measurements of Heymann [6] a dependency of the flow separations on the valve lift was expected. So the valve was investigated at different valve lifts. The investigation is focused on small valve lifts because in this operational range separations are expected in typical power plant use.
3D investigations were undertaken at the following valve lifts: • 1,8% lift • 6,8% lift • 13,6% lift • 20,5% lift A lift of 100% refers to a completely opened valve; at lift of 0% the valve is closed. Special care has to be taken on the grid generation process due to supersonic flow which occurs in the region of the valve seat. Numerical approach For the spatial discretization a block structured grid with hexahedral elements is chosen. The hexahedral grid enables the generation of high quality cells which are needed in the region where the flow is supersonic. Due to high velocities close to the wall and the high density of steam at typical operation conditions, the spacing of the first cell has to be very small in order to retain good y+ values. In the case of 5% valve stroke, at which the highest Mach numbers are reached, the height of the first cell has to be set to 0.001mm in areas of high speed flow. Using this cell height the y+ values can be kept below 45. The grid is refined in areas where supersonic flow prevails. A previous grid study shows that the structure of the under expanded wall jet is strongly depending on the grid. In this study the pressure distribution obtained for different grids is compared. The grid spacing at the valve seat, where expansion waves and shocks occur, is varied from 5mm to 0.25mm. The non-dimensional pressure distributions are shown in Fig 6. The normalized length ξ is defined by Eq. 3 in which the stream wise coordinate is divided by the diameter of the valve seat. The nondimensional local pressure ratio πLoc is given by Eq. 4. ξ=
x
DSeat
πLoc =
p
pt,in
(3) (4)
It can be seen that sufficing grid independence is reached at a cell length of 0.5 mm. In the subsequent diffuser, where the flow is subsonic again, 4
the cell length is about 20 mm. Larger cells are used inside the valve plug, where flow velocities are small. 0,14
Grid spacing
5mm 2mm 1mm 0.5mm 0.25mm
0,12
π Loc[-]
0,1 0,08 0,06 0,04 0,02 0 0
0,2
0,4
0,6
0,8
1
1,2
ξ [-]
Fig. 6: Normalized pressure distribution
Real gas properties of steam are taken into account. A previous calculation of the mass flow shows that an ideal gas model deviates from the IAPWSModel by 30%. So the material properties of steam were modeled using the IAPWS-IF-97 formulation, which is the standard reference for thermodynamic calculations in industrial steam processes. The effects of turbulence are taken into account using the k-ω-SST turbulence model. It enables a short calculation time and is established as a standard model for steady state CFD simulations in turbo machinery components.
Cutting surface
In order to reduce the computational effort, a comparison between the full model (Fig. 7) and a 90° model (Fig. 8) was carried out. The 90°-model only comprises a part of the annular chamber, the control cone and the subsequent diffuser. The cutting surface at which the rest of the valve geometry is cut off is shown in Fig. 7. The two side planes of the 90°-model are treated as a rotationally periodic interface. The first calculation with the full CFD model, which is given in Fig. 7 was carried out for the highest valve lift investigated in this simulation series. It shows that the total pressure losses between the inlet and the cutting plane are very small. The pressure losses in this part amount to 0.23% of the total inlet pressure whereas the losses in the subsequent part of the valve amount to 37.2% of the total inlet pressure. So this upper part is negligible in terms of total pressure. In the following step, the influence on the flow distribution in the subsequent diffusor is investigated.
Fig. 8: Reduced model at 20.5% lift
Mach number contours in the plane of the valve seat are shown in Fig. 9 for the full model (left side) and the reduced model (right side). A high level of rotational symmetry can be seen in the full model. The convexities in the high speed wall jet shown in Fig. 9 are caused by vortices which are created at the flow straighteners. Mach number High
Fig. 7 : Full CFD model at 20,5% lift
Simplification of the CFD model During the process of grid generation it became clear, that high grid resolutions in the area of the valve seat are necessary to resolve the supersonic wall jet. In addition, many nodes have to be used to resolve the boundaries of the flow straighteners. So the mesh size is between 17.7 million nodes in case of the 20,5% valve lift and 22.5 million nodes in the 1,8% lift case.
Low
Fig. 9: Mach number distributions of full model and reduced model
The distribution of the Mach number is in the reduced model nearly the same. Confirmed by this 5
Fig. 10: 2D model
Investigation of wall jet separation For small valve lifts, the valve plug forms a small gap through which the steam has to pass. This flow constriction acts as a nozzle, and the steam is accelerated. After passing this constriction, the high speed flow enters the volume below the valve plug, where mainly low and medium speed flow exists. So a jet is formed in the Valve diffuser. It can be either detached or attached In Fig. 11 vector plots for both cases are shown. The valve lift is in both cases 20.5%. For the detached jet (left side) the pressure ratio is 0,467 and for the attached jet (right side) the pressure ratio is 0.480. The pressure ratio is defined according to Eq.2. Because of the low valve pressure ratio and additional pressure recovery in the diffusor the local pressure ratio at the valve gap is in both cases below the critical pressure ratio of 0.549. So the valve is chocked. As the flow is chocked and the inlet pressure is fixed, the mass flow rate is equal in both simulations. Additionally the jet has at the exit of the gap the speed of sound and the critical pressure due to the chocked flow. As the pressure downstream the gap is lower than the critical pressure the jet is accelerated to supersonic speed. Due to the hyperbolic nature of supersonic flow, expansion waves are created during this expansion process. The expansion waves are reflected at the boundary of jet and
Valve plug
form compression waves or oblique shocks, depending on the conditions. Valve plug
result, the further 3D calculations are performed using reduced models. A further reduction in model complexity and computational effort offer the opportunity for 2D simulations. In Fig.10 the 2D flow domain is shown. In this simplified geometry the influence of the flow straighteners is not represented. Due to that an effect on the flow separation at the wall is likely. Therefore in the next section the flow separations will be calculated with the 2D and 3D model and a comparison of the results is presented.
Mach number High
Low Detached flow Attached flow
Fig. 11: Flow topology in case of detached and attached flow
In case of the attached flow the jet follows the curved valve seat because of the so called Coandaeffect. According to Lubert [8] this effect is caused by entrainment of ambient fluid to the jet. If one side of the jet is proximate to a wall, fluid can’t be entrained from this side and a low-pressure region is created. Thus the jet is deflected to the curved wall. The centrifugal forces of the jet are compensated by the pressure gradient caused by the suction effect. This balance of forces can mathematically be described by Eq.5 given by Cornelius [2] ∂p ρ ∙ u2 = ∂r r
(5)
As the jet has passed the curved seat, it forms an annular wall jet in the diffusor, where its high kinetic energy is dissipated gradually by friction. In case of detached flow the flow pattern changes completely. A short distance downstream the valve gap the jet separates from the valve seat. This separation is induced by an oblique compression shock caused by the strong expansion waves. Due to the adverse pressure gradient and the flow deflection caused by the shock, the jet does not follow the 6
0,6 0,5
π Valve [-]
0,4 0,3 0,2
90° 3D model
0,1
2D model
0 0
10
20
30
40
50
Stroke [%]
Fig. 12: Pressure ratio of reattachment v.s. valve lift Relation to the Coanda wall jet detachment
A dependency of geometric parameters on the pressure ratios of attachment and detachment of under expanded Coanda wall jets is reported by evaluating experimental investigations. Many independent experiments were conducted by Gregory-Smith [5], Matsuo [9], Lowry (Published by Taylor[14]) and Bevilaqua [1]. It can be summarized, that the pressure ratio depends on the ratio of the initial jet height to the radius of the curved surface. The definitions of the radius of curvature R and the jet height h of a basic Coanda-jet configuration are given in Fig 13. In the experiments of Matsuo and Gregory-Smith compressed air is used to create a supersonic jet that flows over curved surfaces. The jet height is varied in the experiments by changing the height of the orifice. Additionally, Matsuo [9] changes also the radius of the curved wall. The experiments show that the pressure ratios of detachment and reattachment depend on the radius to height ratio. In Fig. 14 the pressure ratio of reattachment measured in three experiments is depicted versus the geometric ratio R/h. It can be seen that the results for the reattachment pressure ratio of the experiments match well. pt
p∞
h
wall contour anymore. It flows into the center of the diffusor where the separated annular wall jet changes to a single jet. This process is likely to cause the pressure fluctuations. After that, the single high speed free jet flows along the core of the diffusor. Like in Heymans measurements the appearing flow topology depends on the pressure ratio. For low pressure ratios the jet is separated and for high pressure ratios the jet attaches to the wall. Due to a hysteresis effect a pressure ratio of detachment and a pressure ratio of reattachment can be defined. The pressure ratio of detachment is defined as the pressure ratio at which an attached jet separates from the wall. Conversely the pressure ratio of reattachment is the pressure ratio at which a detached jet reattaches to the wall. For example in case of 20.5% valve lift, the pressure ratio of detachment is 0.400 and the pressure ratio of reattachment is 0.480. The pressure ratio of reattachment is more important than the pressure ratio of detachment because above pressure ratio of reattachment the wall jet is definitely attached to the wall independent of the history of the flow field. Hence this investigation is focused on the pressure ratio of reattachment. In Fig. 12 the reattachment pressure ratio is plotted versus the valve lift. It is shown that the pressure ratio of reattachment increases for increasing valve lift The pressure ratios were obtained by series of simulations. In order to consider hysteresis effects care has to be taken during the simulation runs. So the simulation series are started with a flow field containing the separated jet. When the results are obtained for a certain pressure ratio, they are used as initial values for the simulation of the next higher pressure ratio. During a simulation run the total inlet pressure is kept constant and the outlet pressure is increased in steps of 1.3% of the total inlet pressure. The axisymmetric simulations (2D) and the 3D simulations are in good agreement. So the wake generated by the flow straighteners doesn’t affect the separational behavior of the jet much and 2D simulations can be used for a fast prediction of the wall jet separation.
π Coanda =
p∞ pt
Fig. 13: Basic Coanda wall jet configuration
As the viscosity of high pressure steam differs from the viscosity of ambient air, which was used in the experiments, Reynolds-Number effects have to be taken into account. The Reynolds-Number of 7
the wall jet is defined in Eq.6 using the kinematic viscosity, the height of the orifice and the isentropic velocity of the jet. Re =
h ∙ uis ν
π [-]
0,5 0,4 0,3 0,2 0,1 0 0
20
40 R/h [-]
pout
rSeat
Bevilaqua [1] Gregory-Smith [5] Matsuo [8] Taylor [14] 3D CFD 2D CFD
0,6
pseat
(6)
The range of the Reynolds-Number in the performed simulations is from 2.9 ∙ 105 to 8.0 ∙ 106 . In the experiments of Matsuo the range is from 1.8 ∙ 105 to 1.0 ∙ 106 and in the experiment of Gregory-Smith it is from 1.0 ∙ 105 to 5.1 ∙ 105 . Newman [12], who investigated Re-Number effects in Coanda-wall jets, found that above a ReynoldsNumber of 4 ∙ 105 the separation of the jet is independent of the Reynolds-Number. So the difference in Reynolds-Number between the experiments and the simulation can be neglected. These experimental values are compared to the local pressure ratio at the center of the valve seat when the jet attaches. The definition of the pressure ratio is given in Eq. 7. The local pressure at the seat is taken according to Fig 15. Due to the pressure recovery in the diffuser the local pressure ratio is smaller than the overall pressure ratio. The R/h-ratio is taken from the radius of the valve seat and the gap exit height between the plug and the seat according to Fig. 15. 0,7
Gap height
60
80
100
Fig. 14 : Comparison to experimental results
The agreement between the experiments and the valve simulation in Fig. 14 is quite well except the point at an R/h-ratio of 96. For this ratio the 2D as well as the 3D simulation deviate from the measurements.
Fig 15: Definition of seat and outlet pressure
Πseat =
pseat pt,in
( 7)
The deviation to lower values at this R/h-ratio can be explained by the shape of the gap between the cone and the seat. At very low lift positions the cross section profile of the gap is not pure convergent as in the other cases but it is convergentdivergent like a De-Laval-nozzle. Due to the supersonic expansion in the divergent part of the gap the strength of the sudden expansion after the gap is slightly reduced. So the jet attaches at smaller pressure ratios. Similar effects are reported by Cornelius [2] and Bevilaqua [1]. The observations of both researchers show that the pressure ratio of detachment is lowered when using a convergent divergent channel. Prediction tool Based on the above described findings a new prediction tool is developed. This tool is able to compute the pressure at the valve seat by a pressure recovery correlation in dependence of mass flow rate, the pressure ratio of the entire valve and the geometry. Using these parameters the lift of the valve and hence the height of the gap between the seat and the cone is computed. The R/h-ratio and the local pressure ratio at the seat are compared to a best fit curve of the experiments presented in Fig. 14. By varying the mass flow at several pressure ratios, the reattachment curve can be obtained. It shows the mass flow rate at which the jet separates versus the vale’s pressure ratio. In Fig. 16 a comparison between the model and the CFD calculations is shown. q is the non-dimensional mass flow rate which is defined by Eq. 8. q=
mass flow rate critical massflux ∙ Area of valve seat
(8)
8
turbine, is plotted in Fig. 18. The point, where the transition of vibration occurred in the power station is marked on the operational line. It can be seen that the point of the pressure ratio where the flow transition takes place is close to the pressure ratio at which the operating line meets the line of jet attachment.
0,35 Wall jet tool 0,3 90° 3D model 0,25
q [-]
2D model 0,2 0,15 0,1
0,7
0,05
Lines of constant valve lift
0,6
50% Lift
0 0,1
0,2
0,3
π Valve [-]
0,4
0,5
Fig. 16: Comparison of wall jet tool and CFD simulation
The agreement between the CFD simulations and the prediction tool is quite well. Also a comparison to experimental results is presented. The detachment curve shown in Fig. 17 differs slightly from the curve in Fig. 16 due to different diffuser geometries in the CFD model and the test rig which was built for a previous design generation. Wall jet tool 0,3 Measurements
q [-]
0,25 0,2 0,15 0,1 0,05 0 0,1
0,2
0,3
40% 0,4
Curve of attachment
Operational curve
30% 0,3 0,2 0,1
20% 10%
Tansition point
0 0,1
0,2
0,3
0,4
0,5
π Valve[-]
0,6
0,7
0,8
0,9
1
Fig. 18: Calculated results of the prediction tool
So this single measurement confirms that the separated wall jet creates more vibration than the attached wall jet.
0,35
0
0,5
0,6
q [-]
0
0,4
0,5
0,6
π Valve [-] Fig. 17: Comparison of wall jet tool and transition
points measured in the experiments of Krämer [6]
A quite good agreement between the measured transition points at which a sudden change in pressure fluctuations or vibration is observed and the predicted reattachment curve can be seen in Fig. 17 Also the line of attachment is computed for the valve used in the power station whose operational records are shown in Fig. 5. In Fig. 18 the results of the calculation tool are shown. Additionally the operational line of the valve, which is determined by the swallowing capacity of the connected steam
Conclusion In the investigated valve wall jet separations occur at certain operation conditions. The separation is induced by oblique shock waves, which are caused by the large pressure gradients in the under expanded jet. A dependency of the reattachment pressure ratio on the valve lift is found: The larger the lift is the higher is the pressure ratio of wall jet attachment. As the results of 2D and 3D simulations are in coincidence, the numerical effort can be significantly reduced. The behavior of the wall jet separation is in coincidence with the basic behavior of the under expanded Coanda-wall jet. Based on this finding a new calculation tool for the prediction of the pressure ratio of wall jet attachment was developed. It shows a good agreement to the performed CFD simulations. In addition operation points of model valves which show a sudden decrease in pressure fluctuations and vibration lie close to the calculated line of reattachment. The measurement of a power plant shows that the valve vibrations suddenly decrease at the point where the operating line crosses the line of reat9
tachment. So the higher vibrational level can be related to the detached flow topology. By applying the calculation tool, the flow attachment can be predicted for valves with different geometric and thermodynamic parameters within a few minutes. In contrast to that, 3D CFD studies would consume some months and 2D CFD studies would consume some days on a single computer. Using this tool an attachment line with a higher slope can be obtained by changing the valve geometry in a reasonable time. By this optimization the point where the operating line meets the line of attachment can be shifted to lower mass flow rates and pressure ratios where less vibration is expected. Hence valves with less vibration can be designed in a comparative short time. Outlook A further step will be the investigation of the unsteady behavior of the flow and the effects of flow induced forces on the valve structure. References [1] Bevilaqua, Paul M., Lee, John D., 1980, “Development of a Nozzle to improve the turning of Supersonic Coanda Jets”, Technical Report Air Force Wright Aeronautical Laboratories [2] Cornelius, K. C., and Lucius, G. A., 1994, “Physics of Coanda Jet Detachment at High-Pressure Ratio”, Journal of Aircraft, 31(3), 519-596 [3] Darwish, M. and Bates C. L., 1977, “Flow Vortex Shedding Forces in Check Valves”, Advances in instrumentation, 4. ISA Conference and Exhibition 32, 79-90 [4] Engeda, A., 2012, “Performance Study and Instability Analysis of Steam Turbine Control Valve”, Proceedings of ASME Turbo Expo 2012, June 11-15; Copenhagen, Denmark [5] Gregory-Smith, D. G. and Gilchrist A. R., 1987, “The compressible Coanda wall jet-an experimental study of jet structure and breakaway”, International Journal of Heat and Fluid Flow, 8, 156-164 [6] Heymann, F. J. and Statiano, M. A., 1973, “Steam Turbine Control Valves Noise”, 85th Meeting of the Acoustical Society of America Boston, Massachusetts, April 10-13 [6] Krämer, and Judith, 1976, “Strömungstechnische Modellversuche an SS-Abfangklappen und HDVentilen” (Fluid Dynamic Experiments on Models of Emergency Stop Check Vales and High Pressure Control Valves), Technical Report Siemens AG
[7] Lubert, C. 2011, “On Some Recent Applications of the Coanda Effect”, International Journal of Acoustics and Vibration, 16(3), 144-151. [8] Matsuo, S. et al, 1998, “Study on the Characteristics of Supersonic Coanda Jet”, Journal of Thermal Science 7(3), 165-175 [9] Morita, R., Inada, F., Mori, M., Tezuka, K., Tsujimoto, Y., 2007, “CFD Simulations and Experiments of Flow Fluctuations around a Steam Control Valve”, Journal of. Fluid Engineering, 129(1), 4854. [10] Nakano, M. Outa, K. and Tajima, K., 1988, “Noise and Vibration Related to the Patterns of Supersonic Annular Flow in a Pressure Reducing Gas Valve, Journal of Fluid Engineering, 110 (1), 55-61 [11] Newman, B. G., 1961, “The deflection of plane jets by adjacent boundaries: Coanda effect” In Boundary Layer Control Principles and Application, Lachmann G. V 232-254 Pergamon, New York. [12] Stastny, M., Bednar, L., Tajc, L., Kolar, P., Martinu, P. and Matas R. 2003, “Pulsating Flows in the Inlet of a Nuclear Steam Turbine”, 5th European Conference on Turbomachinery, Prague, Czech Republic [13] Taylor D. W. 1975 “ Experimental Investigation of the High Velocity Coanda Wall Jet Applied to Bluff Trailing Edge Circulation Control Airfoils”, Naval Ship Research and Development Center [14] Zaryankin A. E., Simonov B.P. 1996, “New Control valves for Steam Turbines, Their Characteristics, and Experience with their Operation” Thermal Engineering, 43(1), 19-24 [15] Zhang, D. and Engeda, A. 2003, “Venturi valves for steam turbines and improved design considerations” Journal of Power and energy, 216, 66-74 [16] Ziada, S. and Bühlmann, E. T., 1989, “Flow Impingement as an Excitation Source in Control valves”, Journal of Fluids and Structures, 3, 529549
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