JP2009523211A - Turbine blade having a concave tip - Google Patents

Abstract

Turbine blade (1) with side (7) having an aerodynamic profile. The turbine blade (1) further comprises an end face (6) and is mounted in the turbine, whereby the end face (6) is defined by a gap from the turbine casing (8). The end face (6) comprises a recess (2) that is formed to act as an improved aerodynamic seal.
[Selection] Figure 6

Description

Field of Invention

  The present invention relates to the field of turbine blades in high load axial rotor turbines, for example used in aero engines or for power generation.

Background art description

  In axial turbines, the tip clearance flow that occurs in the rotor blade row accounts for about one third of the aerodynamic losses in the blade row and is often a limiting factor in blade life.

  The tip leakage vortex occurs from the pressure side to the suction side when the leaking fluid crosses the gap between the rotor blade tip and the casing and rolls up into a vortex on the blade suction side. The flow through the tip gap has a high velocity and high temperature, and the heat transfer from the hot fluid to the blade is very high in the blade tip region. Blade tip cooling is commonly used to avoid blade tip burnout and machine failure.

  Tip clearance flow has recently been studied with many contributions in the published literature. An important contribution that is always referenced is the work by Rains [1] (see list of references below) that studied tip clearance flow in axial pumps motivated by cavitation concerns. Moore & Tilton [2] investigates tip leakage flow experimentally and analytically. The flow structure inside the gap and the heat transfer to the blade are discussed. A flow model is provided that is responsible for gap losses associated with complete mixing behind a constriction (generally the point in the fluid flow where the flow diameter is the smallest) that results in a uniform flow condition at the gap outlet. Binder [3] measured and studied tip clearance loss formation. He was able to divide the total end wall loss into a loss formed in the tip gap, a mixing loss of the tip leakage vortex, and a secondary / end wall loss. He noted that not only the tip leakage mass flow is important in loss formation (48% of the total loss seen in mixing loss), but the flow structure in the gap also plays an important role (of the total loss formed in the gap). 39%). He also presented a conceptual model for tip clearance loss formation. Bindon & Morphis [4] studied losses in different blade tip shapes. They found that the total loss remains unchanged, but the gap loss varies greatly between the baseline sharp edge flat tip and the contoured blade tip. The sharp-edged case showed a high loss in the gap with a strong peeling bubble at the pressure lip, whereas the contoured case showed a low loss because no peeling bubble was formed at the gap inlet However, increased tip clearance mass flow was found.

  In a study of different squealer tips by Heyes, Hodson & Dailey [5], an exfoliating bubble with an associated constriction effectively seals the gap, reducing tip clearance loss by reducing tip mass flow It was also concluded that it could be This study also includes a flow model for the squealer tip and at the gap exit, expressed as a combination of the isentropic jet between the casing and the gap mid-height and the wake formed behind the exfoliation bubble. A model for the flow is also included. The computational study by Ameri [6] on heat transfer at the blade tip shows that a flat tip with sharp edges is more efficient and total pressure compared to a flat tip with an average camber line squealer tip and a rounded blade tip. It has been shown to work best with respect to loss. Further computational studies by Tallman [7], [8] on the linear cascade experimentally studied by Bindon & Morphis [4] show the effects of tip clearance height and relative casing wall movement on physical flow in the tip gap. Arguing. Further experimental work on the physical process of tip clearance flow due to moving casing walls is given in a two part study by Yaras & Sjorander [9, 10]. It has been found that a moving belt that simulates relative casing motion significantly reduces tip clearance mass flow.

  Further, the tip passing vortex is also drawn to the suction side, thereby providing a throttling effect. In addition, a reduction in the pressure differential driving the flow into the gap was also observed.

  In a recent study, detailed heat transfer to the concave blade tip was first studied by Dunn et al. [12]. The concave blade tip is equipped with a heat flux gauge and has been studied experimentally in a full-stage rotating turbine. Nusselt numbers were shown at different vane / blade spacings. It has been found that the leading edge Nusselt number at the bottom of the cavity exceeds the blade stagnation value.

List of references

  [1] Rains, D .; A. , 1954, “Tip Clearance Flows in Axial Flow Compressors and Pumps”, California Institute of Technology, Hydrodynamics and Engineering Engineers. 5, June 1954.

  [2] Moore, J. et al. & Tilton, J.M. S. , 1988, “Tip Leakage Flow in a Linear Turbine Cascade,” ASME Journal of Turbomachinery, Vol 110, pp. 18-26.

  [3] Bindon, J .; P. , 1989, “The Measurement and Forum of Tip Leakage Loss,” ASME Journal of Turbomachinery, Vol 111, pp. 257-263.

  [4] Morphis, G .; & Bindon, J.M. P. , 1992, “The Development of Axial Turbine Leakage Loss for Two Profiled Tip Geometrics Using Linear Cascade Data,” ASME Journal of Volt. 114. 198-203.

  [5] Heyes, F .; J. et al. G. Hodson, H .; P. , & Dailey, G. M.M. 1992, “The Effect of Blade Tip Geometry on the Tip Leakage Flow in Axial Turbine Cascades,” ASME Journal of Turbomachinery, Vol. 114. 643-651.

  [6] Ameri Ali, A. , 2001, “Heat Transfer and Flow on the Blade Tip of a Gas Turbine Equipped with a Mean-Camberline Strip,” ASME Paper 2001-GT-0156.

  [7] Tallman, J. et al. & Lakshminarayana B. , 2000, “Numerical Simulation of Tip Clearance Flows in Axial Flow Turbines, With Emphasis on Flow Physics, Part I-Effect of Tip Clean Jourmanship. 314-323.

  [8] Tallman, J. et al. & Lakshminarayana B. , 2000, “Numerical Simulation of Tip Clearance Flows in Axial Flow Turbines, With Emphasis on Flow Physics, Part II-Effect of OutTurning ME”. 324-333.

  [9] Yaras, M .; I. & Sjorander, S .; A. , 1992, “Effects of Simulated Rotation on Tip Leakage in a Planar Cascade of Turbine Blades, Part I—Tip Gap Flow,” ASME Journal of Turp. 652-659.

  [10] Yaras, M .; I. & Sjorander, S .; A. , 1992, "Effects of Simulated Rotation on Tip Leakage in a Planar Cascade of Turbine Blades, Part II-Downstream Flow Field and Bure. 660-667.

  [11] Basson, A .; H. & Lakshminarayana, B .; 1995, “Numerical Simulation of Tip Clearance Effects in Turbomachinery,” ASME Journal of Turbomachinery, Vol 109, pp. 545-549.

  [12] Dunn, M .; G. & Haldemann, C.I. W. , 2000, “Time Averaged Heat Flux for a Received Blade Tip, Lip and Platform of a Transonic Turbine Blade”, ASME Paper GT2000-0197.

  [13] Behr, T .; , Kalfas, A .; I. Abhari, R.A. S. , “Unsteady Aerodynamics in Casing Injection for Tip Leakage Treatment in an One-L / 2-Stage Unwrapped Turbine”, ASME Paper No. GT2006-90959.

  [14] Sell M.M. Schliinger J., et al. , Pfau A. , Treiber M .; Abhari R .; S. 2001, “The 2-stage Axial Turbine Test Facility LISA”, ASME Paper No. 2001-GT-049.

Summary of the Invention

  The present invention relates to a concave blade tip for a high load axial flow rotor turbine blade having application in a high pressure axial flow turbine in aero engines or power generation.

  In order to overcome the problems known from the prior art, a blade tip structure with a recess at the blade tip is proposed rather than a simple flat blade tip. A recess (cavity) inside the blade tip serves as an aerodynamic seal that improves turbine performance and / or reduces turbine tip heat load in that it affects flow distribution. Since the material on the upper surface of the blade is moved from the rotation axis to the lower radius, the mechanical stress at the blade root can be reduced. Also in the case of tip friction, i.e. when the rotor blade contacts the casing during rotation, only the thin cavity edge is damaged. Wear on the casing is also limited and the blade tip cooling purge hole is located inside the cavity so that friction does not damage the hole outlet. Therefore, efficient cooling is ensured even when friction occurs. Finally, the recessed cavity may act as a labyrinth seal that may be advantageous in reducing tip clearance mass flow.

  It has been found that proper profiling of the recess shape can significantly reduce the total tip heat transfer Nusselt number, for example, 15% lower than a flat tip and 7% lower than the baseline recess shape. The experimental results also show that the overall efficiency of the entire turbine is improved by 0.3% compared to the flat tip case with the improved recess structure, which is a CFD analysis. Proves the 0.38% prediction from

  Using three-dimensional computational fluid dynamics (CFD), the flow field near the tip of the blade is investigated for different shapes of the recessed cavities. Control of the cavity vertical structure results in an improved structure, highlighting the differences from the blade tips common in the prior art.

  Tip clearance between the rotor blade tips and the casing is necessary for free rotation of the rotor blade rows. However, the gap allows fluid to cross the blade tip from the pressure side of the blade to the suction side due to the pressure difference between the pressure side and the suction side. This flow is associated with two main problems. First, about one third of all aerodynamic losses in the rotor row are related to tip leakage vortices that occur when tip leakage over the blade tip enters the flow through again on the blade suction side. It creates both mixing losses as it mixes with the main flow and disturbs the pressure field on the blade tip wall responsible for the blade lift. Also, the fluid traversing the gap is not rotated by the blade, and therefore no work is extracted from it. It is therefore interpreted as a lost work extraction. Second, the fluid that traverses the tip clearance has a relatively high temperature due to hot streak movement, which results in a high thermal load at the blade tip. In fact, the blade tip will burn out if not properly cooled, and is therefore one of the limiting factors in blade life.

  It is also desirable to minimize the tip clearance gap height to improve performance by reducing the tip leakage mass flow. However, this reduction in gap height sometimes increases the risk of rotor blade friction in the casing during the operating envelope. This can occur, for example, due to transient currents, rotor dynamic trajectories, casing ovalisation, or due to casing thermal deformation or when the rotor extends farther than the casing. If a blade with a flat tip wears hard on the casing, the tip can wear out and catastrophic cooling losses can occur. Even with relatively little wear at the flat tip, any cooling holes located at the tip may be damaged, resulting in poor cooling and eventually burning the blade tip.

  Unlike the case of a flat blade tip, the more complex physical flow at the concave blade tip is more difficult to understand. Also, systematic design procedures for cavity size and shape are not available. Research can overcome this problem, provide a better understanding of the physical processes and aerodynamics of the heat transfer in the recessed cavity, as well as new for standard high-load rotor blades that represent high-pressure turbines. It was possible to give a design boundary. Special three-dimensional CFD instruments have been used extensively for this purpose.

the term

  The invention itself and the computing instrument for improving the results will generally be described with reference to the drawings. Since they are not available on the market, calculators and solutions for pretreatment have been developed by the inventors. These instruments may partially interact with commercial products for post processing.

A pre-designed axial turbine test case was used to perform the intended computational and experimental studies. The 1.5 stage uncovered turbine shape models a high load (DH / U ² = 2.36) low aspect ratio gas turbine environment. The air loop of the test apparatus is semi-closed and includes a radial compressor, a two-stage water-air heat exchanger, and a graduated venturi nozzle for mass flow measurement. Prior to entering the turbine section, the flow passes through a 3m long straight duct containing a flow straightener to ensure a uniformly distributed inlet flow field. On the downstream side of the turbine, the air loop is open to atmospheric conditions. The DC generator absorbs the turbine output and controls the rotational speed of the turbine. An accurate torque meter measures the torque transmitted to the generator by the rotor shaft. The TET (turbine inlet temperature) is controlled with an accuracy of 0.3% and the RPM (rev / min) is kept constant within ± 0.5 min ⁻¹ by the DC generator. More information on the operation of the turbine structure (Behr et al. [13]) and experimental equipment (Sell et al. [14]) can be found in the published literature.

The table below shows the main parameters of the “LISA” 1.5 stage axial turbine research facility at the design operating point (Table 1).

  An optimization of the computational design for a nominally recessed cavity commonly used in axial turbine rotor blades is given. Extensive parameter studies provide an improved recessed cavity structure. An extensive aerodynamic and heat transfer comparison between the new structure and the flat tip blade and the nominal recess tip is given. The calculated data were compared with experimental data from the Swiss Federal Institute of Technology (ETHZ). In this case, the 3D flow structure and performance of a rotor blade with a flat tip and a new recess structure was measured. A qualitative comparison with experimental data from OSU was also used to verify the predicted heat transfer data. The following conclusions can be drawn from this study.

  A further understanding of the three-dimensional flow in the recessed cavity was obtained. Three cavity vortices were found to dominate the leakage flow through the cavity.

  Cavity shape changes affect the formation and interaction of major recess vortices. Specific recirculation at the suction front responsible for high heat transfer can be eliminated and can result in a new structure with improved heat transfer behavior.

  The beneficial effect of forming an aerodynamic seal has been shown in both recess structures. The tip leakage mass flow can be reduced by about 25% in the new recess structure according to the present invention compared to the flat tip. CFD showed increased output even in the new structure. Experimental measurements showed a 0.3% increase in turbine efficiency between the flat tip and the new recess tip at the design point.

  The heat load at the blade tip is found to be balanced between the heat loads at the different blade tip elements: tip edge, cavity edge wall, cavity bottom, and rear flat blade tip. The new recess structure has a total heat load of about 7% lower than the baseline recess structure and 15% lower than the flat tip.

  To the best of the author's knowledge, this is the first time detailed profiling of the blade tip recess cavities is shown to improve performance and reduce thermal load.

  • Three-dimensional geometric profiling of the blade tip recess cavity wall significantly improves overall efficiency while effectively reducing both thermal load and mechanical stress at the blade tip.

  • Three-dimensional geometric profiling of the concave cavity wall can be achieved by using non-uniform edge wall thickness and three-dimensional shaping inside the cavity concave volume.

  3D geometric profiling occurs by changing the recess edge thickness and cavity depth individually or in combination.

  3D geometric profiling of recessed cavities optimizes leakage flow and its interaction with vortices in the cavity, while creating flow in the cavity due to flow at the cavity wall and high thermal loads Suppress.

  ・ Three-dimensional geometric profiling of blade tip recess cavities with the above features reduces and suppresses secondary vortex formation and limits the space available for vortex formation in both circumferential and radial directions To do.

  • Three-dimensional geometric profiling of the concave cavity wall results in a higher working output of the blade due to the additional blade surface of the cavity wall in the case of advantageous cavity pressure gradients.

  3D geometric profiling according to the above features defines the optimal cavity volume to combine.

  -Three-dimensional geometric profiling that exhibits the above features reduces the number of vortices fitted in the circumferential direction.

  3D geometric profiling that exhibits the above characteristics results in a vortex pattern with less energy dissipation in the cavity.

  ・ Three dimensional geometric profiling of blade tip recess cavities with the above characteristics can be achieved for cavities with constant edge thickness that are not contoured, such as surface areas with reduced wear, cooling hole protection, and reduced mechanical stress. Maintain the benefits.

  -Three-dimensional geometric profiling of blade tip recess cavities exhibiting the above characteristics is applicable in uncovered and certain partially covered turbine blades.

  A three-dimensional geometrically contoured tip recess cavity wall exhibiting the above characteristics provides passive flow control of leakage flow by acting as a throttling mechanism.

  A 3D geometrically contoured tip cavity wall based on a 3D flow structure that exhibits the above features to redesign the blade tips of both existing and alternative blades for improved performance Can be used and incorporated into new structural forms.

  The present invention relates to a turbine blade having a side surface having an aerodynamic profile. The turbine blade has an end face located at a mounting position in the turbine defined by a gap from the turbine casing. The end face is provided with a recess formed to act as an aerodynamic seal and / or to reduce blade tip thermal load. The recess is defined by a side wall having a wall thickness that is variable in the circumferential direction. The side wall has a substantially constant wall thickness in the height direction (the length direction of the turbine blade). In certain embodiments, the wall thickness of the sidewall has an overall maximum in the region where the turbine blade has its maximum thickness. In a further embodiment, the wall thickness of the sidewall has a maximum value on the suction side of the turbine blade, with a relative length of + 0% to + 50% relative to the blade leading edge (trailing edge ± 100%). In certain embodiments, this maximum value is an overall maximum value. However, in certain other embodiments, the wall thickness of the sidewall has a minimum value on the pressure side of the turbine blade. Depending on the field of application, this may be located in a relative length range of −0% to −70% with respect to the blade leading edge stagnation point (0%). The wall thickness on the pressure side may be at least partially constant. Good results may be achieved in that the overall maximum wall thickness of the sidewalls is approximately 5.5 to 6.5 times greater than the overall minimum. In one embodiment, the recess has a substantially constant depth.

  The invention described herein will be more fully understood from the accompanying drawings and the following detailed description given herein, which should not be considered as limiting the invention as set forth in the appended claims. The drawings are shown in a simplified and schematic manner.

Description of embodiment

  The following embodiment of the present invention will be described in more detail. Similar features are in different figures with the same reference numerals.

  FIG. 1 shows a computer model of two turbine blades for numerical calculation of the behavior of a recess located at the tip of a turbine blade.

  The applied analysis model is generally based on a numerical grid (mesh) of the turbine blade 1. The numerical grid used was generated using an in-house developed grid generator called MELLIP. The multi-block grid generator uses a two-dimensional NURBS library as input data to mesh the computational domain boundary. Using a set of geometric transformations, inner block boundaries are defined according to the intended grid topology. Using both a non-linear interpolation algorithm with flexible clustering specifications during each block meshing and a two-dimensional Poisson-type elliptic partial differential equation, high grid quality, ie, smooth grid lines, limited aspect ratio, distortion A cell-to-cell ratio is obtained. As several topologies are implemented, these topologies divide the calculation area for a blade tip having a recessed area into 18 blocks in the blade recessed case. In particular, the use of upstream and downstream wake blocks with adjustable size and grid density helps to maintain low grid distortion in the trailing edge region and prevents numerical diffusion of shed wakes. The grid used for this study exhibits high resolution in the blade tip region to obtain a flow gradient in this region.

  This helps to keep the number of grid points at about 900,000. This is because clustering in the vicinity of the wall need not be as aggressive as in the case of two-layer turbulence model calculations. Therefore, a large number of grid points in the blade tip region results in a uniform mesh density with a smooth cell-to-cell ratio distribution. The dense, dense tip region grid block spans approximately 10% above the blade range.

  The preferred numerical flow solution is called MBStage3D, a three-dimensional structured Navier-Stokes solution in multi-stage turbomachine applications. The time marching algorithm preferably used in MBStage3D is a Jameson type algorithm, ie, an explicit method using a residual averaging technique applied to increase stability. Time discretization is preferably achieved by a five-stage Runge-Kutta technique with fourth order accuracy. All calculations discussed here are performed using an algebraic Baldwin-Lomax turbulence model with Sommerfeld logarithmic wall functions to calculate turbulent viscosity at the wall.

  The extensive post-processing necessary to gain an understanding of flow physics is achieved by 3D, 2D, 1D and scalar investigations of the subject flow field. 3D visualization is performed using TECPLOT, and 3D data generation subroutines for integrator collection and TECPLOT are developed in-house.

  FIG. 2 schematically shows a typical turbine blade 51 known from the prior art. The turbine blade 51 has an outer surface 57 that forms an aerodynamic profile. The turbine blade also includes an end face 56 at a mounting position within the turbine that is spaced from the turbine casing 58 (FIG. 5). The end surface 56 includes a recess 52 formed to serve as an aerodynamic seal. The recess 52 at the blade tip 53 has a length L and a uniform depth D. The wall 54 surrounding the recess 2 has a constant thickness t in the vertical and circumferential directions. In a three-dimensional flow field investigation around a recess 52 and blade tip 53 having a length L of 80% of the axial chord (blade depth) and a depth D twice the tip gap height, The flow characteristics could be confirmed in various ways.

  The geometry of the tip recess 52 and its effect on the flow field distribution around the blade tip 53 is similar to the standard (nominal) recess structure 52 (FIGS. 2, 4, 5) to find a more improved recess structure. (See below). The main geometric parameters that varied were the cavity length L, the cavity depth D, and the shape of the cavity edge, which can be almost generally any function of the previous two parameters. However, the shape was most often determined as a function of length.

  The cavity wall of the recess 2 (see FIG. 6) provides an additional surface that generates additional output that needs to be added to the main output generated by the outer blade wall. It was also observed that the tip leakage flow at the blade tip of the recess case was lower than that at the flat tip. The total mass flow over the computational domain remained unchanged. This suggests that the change in leakage mass flow occurs before the suction side throat region, thus providing a constant corrected flow through the rotor. It is also important to note that the overall pressure loss coefficient change follows the tip gap mass flow change. The relationship between aerodynamic losses and verified tip leakage mass required further reduction of tip leakage mass flow as an important design criterion. The change with depth was rather nonlinear. In this case, the minimum tip leakage mass flow was confirmed with the optimum depth. This is also an important difference from the development of tip leakage mass flow in a flat tip case where the tip leakage mass flow varies linearly with gap height. The change in the length of the recess was almost linear.

  Therefore, an understanding of detailed flow physics is particularly important in the design process and will be investigated next for three major test cases. The first test case is a flat tip blade. The second test case is a nominally recessed cavity known from the prior art with 80% length of the axial chord and twice the depth of the tip gap and with a constant edge thickness. . This test case represents the current design in the recessed cavity. The final test case is a newly designed recess shape based on extensive physical modeling and geometric perturbations.

  FIG. 3 illustrates in a typical manner the flow field around a typical turbine blade 100 having a flat tip 101 with two main flow structures. The first major flow structure is the tip-passing vortex 102 that occurs when the generated driving fluid enters the tip gap from the suction side at the leading edge and exits the tip gap again after about 20% axial chord. Furthermore, the tip leakage that occurs on the pressure side between the leading edge and the approximately 15% axial chord crosses the tip gap and mixes with the generated flow in the tip passing vortex. The supply of tip-through vortices is well organized. In this case, the flow passing through the dividing streamline induces the formation of a tip passing vortex. The subsequent pressure side leakage flow supplying the tip passing vortex can also be seen.

  The second major flow characteristic observed is the tip leakage vortex 103 resulting from the tip leakage flow across the gap from the pressure side starting at about 15% axial chord. A split streamline between the pressure side leakage supplying the tip passing vortex and the tip leakage vortex outer fluid layer can be confirmed. The outer fluid layer in the tip leakage vortex 103 is due to the main part of the pressure driven low gap shear loss that generates the leaking jet. A tip leaking vortex core is formed by the blade tip boundary layer fluid.

  A cutting plane perpendicular to the blade average camber line reveals a well-known interstitial flow structure. When the tip leakage flow enters the gap from the pressure side, a separation bubble is formed, thereby providing a constriction. The leaking jet exiting the constriction then forms a wake at the bottom of the gap. This wake forms a mixing loss and is later found in the tip leaking vortex core. The leaky jet above the wake is often modeled as an isentropic jet and forms an outer fluid layer around the tip leaky vortex core depending on the axial position as it exits the gap on the suction side.

  FIG. 4 schematically shows the flow structure around the turbine blade 51 according to FIG. The recess 52 has a substantially constant wall thickness in the circumferential direction and the vertical direction (the length direction of the turbine blade) over substantially the entire length thereof excluding the trailing edge 62. It should be noted that the flow structure is dependent on the aerodynamic structure of the turbine, which varies somewhat in different turbine blade rotor structures and may therefore not exactly match the description below. Despite this, many of the flow characteristics generally remain the same in modern axial high work turbines, and it is believed that those skilled in the art and sensitivity related to geometrical changes remain applicable. ing. A total of six main flow characteristics were found to affect the cavity flow. Starting from the pressure side leading edge 60, fluid above the fluid at the passing blade leading edge stagnation point 61 enters the cavity 52. The fluid traverses the cavity with low loss and at the fluid flow angle at the leading edge and impinges on the corner between the cavity bottom and the cavity suction sidewall immediately after peak suction. After striking the corner wall, this pressure side leading edge jet rolls up into the vortex structure C, moves downstream inside the cavity 52, partially exits the cavity 52, and in the suction passage flow Enter again. The boundary layer on the recess edge that enters the cavity along the entire pressure side 20 due to the tip clearance pressure gradient on the first 20% axial chord immediately on the suction side is between the cavity bottom and the cavity wall. Rolls up into the vortex A at the corners, thereby spreading along the entire pressure side 20 to the end of the cavity and reaching the cavity wall to 20% of the axial chord on the suction side 21. After 10% of the axial chord, this suction side portion of the vortex C is lifted away from the cavity bottom by a generated drive suction side leak jet entering the cavity. The third important flow characteristic is the vortex B that occurs when the casing boundary layer fluid rolls up against the pressure side tip leak jet TL. This vortex remains on the casing wall and diverts the pressure tip leak inside the cavity as shown in FIG. A split streamline DS1 is established between the two vertical flow structures.

  Downstream of the 20% axial chord, the flow behavior of the pressure side leak is similar to the flat tip case with the difference that the leak is deflected by the cavity vortex and interacts with the cavity vortex. After leaving the gap on the suction side, this fluid again forms the outer layer of tip leakage vortices and tip passing vortices. The core of the tip passing vortex is formed by the same generated tip leakage that lifts away from the cavity corner vortex as it enters and exits the cavity with 10% axial chord to 20% axial chord. The core of the tip leakage vortex is the wake behind the separation bubble at the suction side edge that occurs when the pressure side leakage jet exits the cavity.

  In order to further clarify the flow characteristics inside the cavity, a cut plane perpendicular to the camber line located downstream of the vortex former formed by the pressure side leading edge jet is shown in FIG. Reference is made to the three main cavity vortices. Between the casing vortex B and the edge boundary layer winding vortex A, the pressure side leakage traverses the cavity lifting vortex C caused by the pressure side leading edge cavity jet. Two stripping bubbles are formed on the pressure side rim edge against the outer blade wall when the tip leak jet enters the gap, and the suction side rim edge with the cavity wall when the tip leak exits the cavity again The other is formed. The tip leak jet is recognized as a low entropy region between the high entropy regions where the vortices are located. The casing vortex is somewhat squeezed, which explains why more fluid gradually enters the adjacent vortex downstream of the blade.

  The nominal structure showed many vortex structures in the recessed cavity. In particular, the front of the cavity is affected by these structures. As can be seen, boundary layer fluid leaking from the edge into the cavity rolls up in the vortex along the corner between the cavity bottom and the cavity edge wall. The purpose of the new design was to eliminate the recirculation zone at the front of the blade to minimize aerodynamic losses and reduce the head movement coefficient.

  FIG. 6 shows the flow distribution around the tip 3 of the turbine blade 1 according to the present invention. The recess 2 is surrounded by a wall 4 having a variable wall thickness t in the circumferential direction. Thereby, the wall 4 has a maximum thickness t on the suction side 21 in the region where the turbine blade 1 has a maximum value over all thicknesses. In the side surface 5 of the recess 2, the walls 4 are here arranged substantially parallel to the longitudinal axis of the turbine blade 1 (substantially perpendicular to the end face 6 of the turbine blade 1). However, other structures may be suitable depending on the field of application. One advantage is that the structure shown is relatively easy to form by a standard grinding process.

  One advantage of the improved structure is that the streamline separating the recirculation region from the pressure side leading edge jet is moved. FIG. 6 shows the effect of the improved structure on the cavity flow pattern. Recirculation at the leading edge 10 is suppressed. Here, the pressure side leading edge jet spreads in the whole cavity 2. The casing boundary layer fluid rolls up into the vortex B and interacts with the fluid from the pressure side leading edge jet. At this time, the fluid trapped by recirculation in the nominal structure enters the cavity and is pushed out again by the pressure side leading edge jet. After the fluid exits the cavity, it supplies a tip passing vortex.

  FIG. 7 shows the eddy current pattern around the tip 3 of the turbine blade 1 and in the cavity 2 along a cutting plane parallel to the longitudinal direction of the turbine blade 1. As shown, the end face 6 of the turbine blade 1 is defined by a gap 12 from the turbine casing 8. The vortices A and B interact positively. Thus, the vortex B is no longer limited to the casing and can occupy the entire cavity volume. Vortex C is formed by the fluid that resulted in suction side leading edge recirculation in the nominal case. The vortex is formed when the fluid separates at the cavity edge while being pushed out of the cavity by the pressure side leading edge jet.

  Thereafter, results on the aerodynamic performance of the three test cases are introduced. In aerodynamic performance, tip leakage mass flow is investigated. Because it is intensively involved in the overall pressure loss. To assess the effect of the new structure on heat transfer, the cumulative heat flux vector and Nusselt number distribution on the blade tip wall are compared.

FIG. 8 individually shows the change in accumulated tip clearance mass flow from the leading edge to the trailing edge on the pressure and suction sides for the three investigated test cases. In the figure, x means length and c ax means axial chord, mdt ^* indicates the leakage mass flow over the blade tip as a function of axial distance. On the pressure side, the most important feature is the cumulative tip gap mass flow growth from the leading edge to the trailing edge at the flat tip compared to a linear increase as long as the recessed cavity opens behind the edge. . Downstream of the cavity trailing edge, the accumulated tip clearance mass flow also varies non-linearly and resembles a flat tip case in both recessed structures. The linear increase may be explained by the mean static pressure change at the tip gap inlet on the pressure side and at the tip gap outlet on the suction side.

  As shown in FIG. 9, static pressure decreases non-uniformly on both the pressure and suction sides in the case of a flat tip. The figure shows the static pressure change according to the axial length. However, in the case of a concave case, the static pressure always remains at a constant level when the tip leak enters the cavity, which is the case in the entire pressure side gap, but before the suction side where the generated fluid enters the cavity. It is also a case in the department. The recessed cavity acts like a reservoir where the pressure remains constant. The sealing effect and the resulting reduction of the accumulated tip clearance mass flow is clearly observed. The nominally recessed case showed a 23% reduction in leakage mass flow compared to the flat tip, and the new structure had 25% less mass flow across the gap than the flat tip.

  FIG. 10 shows the accumulated tip clearance mass flow at the gap exit on the blade suction side, flat tip blade (prior art), recess with constant wall thickness (nominal recess, prior art), and variable wall. It is shown with respect to an improved recess having a thickness (shown as a new structure according to the invention). The generated drive tip leakage mass flow is much more severe in the recessed case than in the flat tip case. Occurrence leaks reach 22% axial chords from the leading edge in all three cases. However, the amount of mass flow that enters the recessed cavity is approximately twice that at the flat tip. This again maintains the cavity in a state that acts as a reservoir to be filled with fluid. The difference in the amount of mass flow further entering the cavity compared to the flat tip gap is almost exactly consistent with the decrease in total accumulated tip leakage mass flow across the suction side. It can be seen that there is no significant difference in tip clearance mass flow between the nominal structure and the new recessed structure.

  FIG. 11 shows heat transfer at the tip of three different turbine blades. FIG. 11a shows, in a comparative manner, the Nusselt number (Nu number) distribution of a typical turbine blade having a flat end face as shown in FIG. 3, while FIGS. 11b and 11c show prior art (“nominal”). And the distribution in a turbine blade having a recess according to the present invention (shown as “new”). A high Nusselt number occurs at the leading edge in all three cases. This is the case when the hot fluid first contacts the blade tip and has the highest heat transfer. At the pressure side edge, the Nusselt number has the same magnitude in all three cases. The thin edge of the nominal recess shows a similar distribution at the suction side front compared to the new structure. When a tip leakage vortex occurs (about 25% axial chord), the Nusselt number value decreases in both the flat tip case and the nominal recess case. However, this observation cannot be made at the suction side edge of the new structure. Both the flat tip and the nominal recess show a high Nu number at the leading edge. However, in the new structure, the Nu value does not reach a high value at the cavity bottom. From FIG. 11 the difference between an improved structure with a recess according to the invention (shown as “new”) and a flat tip (shown as “flat”) is visualized. Blocking the suction side recirculation zone with a thicker edge did not result in a higher Nusselt number at the edge.

  FIG. 12 visualizes the heat load of different turbine blades. With a blade tip of a general flat blade tip (shown as “flat”) and a recess according to the prior art (shown as “nominal”) and the present invention (shown as “new”) The cumulative heat flux vector at the blade tip provides a thermal load in three test cases.

  From FIG. 12, where the heat load is divided according to the affected wall type, the following observations can be made. The overall predicted tip heat load is highest at the flat tip and lowest in the improved structure ("new" structure) with variable wall thickness. The reduction in thermal load (about 14%) between the flat tip and the new structure is about twice as high as between the nominal and new recess structure (7%). However, it should be noted that the new structure also has a higher heat load area than the nominal structure. This is the case at the cavity wall and tip edge to which the cavity edge wall and the remaining rear flat blade tip belong. Increased edge thickness and deeper and shorter cavities are responsible for this increased heat load. However, it has also been found that changing the flow field at the leading edge in the cavity by removing the suction side recirculation zone is effective in reducing the thermal load at the bottom of the cavity.

  In FIG. 13 and FIG. 14, the calculation result from CFD prediction (refer FIG. 3-12) is compared with the experimental result performed in the top secret mode in Swiss Federal Institute of Technology (ETHZ). The flat tip and the new recessed structure blade were experimentally evaluated in the ETHZ axial flow turbine facility LISA.

  FIG. 13a shows the experimental relative total pressure coefficient distribution at the trailing edge of the 14% axial chord downstream rotor blade in a flat tip blade. Initially, the predicted experimental relative total pressure loss coefficients at the flat tip and the new recess structure are compared at a 2D axial cut plane located 14% axial chord downstream of the rotor trailing edge. The experimental data for the flat tip blade shown in FIG. 13a is a snapshot at a given point during unstable flow. Thus, unstable flow characteristics that do not exist in single row steady state CFD results are resolved. The fact of instability data also reveals changes in the low relative total pressure region that identify secondary flow vortices and tip leakage vortices.

  FIG. 13b shows CFD prediction from single row steady state calculations. The computationally predicted relative total pressure loss factor resolves the experimentally measured secondary flow structure with sufficiently good accuracy. Hubs and tip passing vortices are captured both by their spatial extension and the magnitude of the loss. Also, the trailing edge wake of the rotor blade is captured in the CFD result. However, as described above, the loss region associated with the tip clearance vortex is overestimated compared to the experiment.

  FIG. 13c shows the CFD predicted and measured pitch average radial distribution of the relative flow yaw angle at 14% downstream of the axial chord downstream of the rotor blade trailing edge in a flat tip blade. As shown in the figure, the change in relative yaw angle caused by the tip leakage vortex in the 80% range to 100% range is largely overestimated by the calculation result. The magnitude of the relative yaw angle change due to the hub and tip secondary flow structure is well predicted.

  14a and 14b show that experimental data and CFD prediction data for turbine blades with improved recesses according to the present invention (shown as "new" structures) with variable wall thickness in the circumferential direction are compared. Is shown. The experimental data shown in FIG. 14a is a snapshot of unstable data as before. Snapshots of the flat blade tip and the new recessed structure were both taken at the same time point. The CFD predicted relative total pressure coefficient from the steady state calculation is shown in FIG. 14b. Again, as shown, the predicted CFD results resolve the same features obtained by the measured flow field. The predicted relative total pressure loss coefficients showing the hub and tip passing vortices are in good agreement with the experimental data. The overestimation of the tip leakage vortex noted in the flat tip case is also found in the new concave structure case. It can be noted that the spatial extension of the tip leakage vortex loss core in the new recess structure has been reduced compared to the computationally predicted relative total pressure loss factor in the flat tip case.

  FIG. 14c shows the measured and predicted pitch average relative flow yaw angle distribution in the new structure. The CFD predicted relative flow angle distribution is consistent with experimental data up to about 80% of the range. The portion in the 80% range to 100% range is affected by the tip leakage vortex. While the secondary flow characteristics at the hub and tip are accurately predicted, the flow angle differences due to tip leakage vortices are considerably overpredicted.

  In FIG. 15, the experimentally measured relative yaw angle distribution for the flat tip and the new recess structure is given in the tip region from the 60% range to the 100% range casing. As shown, the new recess structure exhibits less over-rotation than a flat tip blade. This result clearly shows that the recess cavity affects the tip leakage vortex responsible for over-determined rotation.

  Experimentally measured performance data show that the “new” recess structure in the turbine used has an overall efficiency of 0.3% when compared to a flat tip at exactly the same all turbine operating conditions. ing. The expected difference between both efficiencies was 0.38%. This is in good agreement with the experimental data quantitatively.

  The predicted heat transfer is quantitatively compared with data provided by the Ohio State University Gas Turbine Laboratory [12]. Turbine blades with recessed cavities similar to the nominal structure given here were equipped with heat transfer gauges to measure heat transfer at the cavity bottom near the leading, trailing and central edges. The edges were also equipped with several gauges. Nusselt numbers were reported at different vane / blade spacings. The trend of Nu number change at the investigated locations is similar. The highest Nu number is found in the leading edge region. The second heat flux gauge was located 12.5% blade axial chord downstream from the peak suction. This reports almost half of the Nu number at the leading edge. This can also be observed in the nominally recessed case from FIG. Finally, the third heat gauge at the 62% blade axial chord corresponding to the 80% cavity axial chord reports approximately the same Nu number at all vane intervals, which is the previous Nu at the short interval. Reported in small vane intervals with respect to numbers.

  16 and 17 show the wall thickness distribution in the circumferential direction of several recesses of the turbine blade according to the present invention. The graph is shown in a normalized manner in that the thickness t is shown relative to the reference thickness t0 on the y-axis so that it is possible to compare different shapes of the turbine blade recesses. ing. Referring to FIG. 6, the local thickness t of the wall 4 is that of the end face 6 of the turbine blade 1 perpendicular to the outer edge 13 of the turbine blade 1 formed by the side face 7 and the end face 6 of the turbine blade 1. Measured at height. In the graph, the local wall thickness t is shown on the x-axis, starting from the blade leading edge stagnation point 0 (0% circumferential length), both in the circumferential direction. The relative position of the trailing edge 11 is indicated by values 1 (100%) and 1 (-100%), respectively. The suction side is positive (+), and the pressure side of the outer shape is negative (−). In the graph, the trailing edge 11 approximately located in the range of +0.6 (+ 60%) to +1 (+ 100%) and 0.6 (-60%) to -1 (-100%) is the thickness of the turbine blade in this region. Is not taken into account in the figure. As shown, the wall thickness t increases on the suction side and, depending on the field of application, on the suction side of the turbine blade in the range of +0.1 (+ 10%) to +0.4 (+ 40%) relative length. It has a (local) maximum value of 15. In another embodiment, the maximum value 15 is arranged at + 11%, 19%, 30%, 37% (error ± 5%), respectively. The maximum value 15 is, in certain embodiments, located in the region of the maximum thickness 14 of the turbine blade. As can be read from the illustrated graph, certain embodiments are within the range of +0.05 (+ 5%) to +0.15 (+ 15%) with a locally reduced increase in thickness. It has a shoulder 16 (see, for example, FIG. 6). In one embodiment, the increase in wall thickness (indicated by arrow 17) already begins on the pressure side of the turbine blade, for example in the range of 0.1-0. As can be seen from the graph, in certain embodiments, the increase and decrease in wall thickness is generally similar to each other and is in the range of 300% to 400% for every 0.1 normalized circumferential length. As can be further read from the graph, the maximum wall thickness is about 5.5 to 6.5 times thicker than the minimum wall thickness on the pressure side.

  On the pressure side, in certain embodiments, the wall thickness is constant, which is generally indicated in this region by the horizontal progression of the graph. However, certain embodiments (see outline 3) may have a local maximum 18 on the pressure side.

  Although the present invention has been described with reference to specific embodiments thereof, many other variations and modifications and other uses will become apparent to those skilled in the art. Accordingly, the present invention is preferably not limited by the specific disclosure herein, but only by the appended claims.

2 is an example of a three-dimensional computational grid turbine rotor blade having a recessed cavity. It is an example of the recessed part cavity structure which concerns on a prior art. 3 is a three-dimensional CFD flow distribution over a flat tip blade known from the prior art. It is a three-dimensional CFD flow distribution around the front-end | tip of 1st Embodiment which has a recessed part based on a prior art. It is a two-dimensional CFD flow distribution in the recessed part which concerns on FIG. 3 is a three-dimensional flow distribution around a blade tip having a recess according to the present invention. It is sectional drawing of the flow distribution which concerns on FIG. Pressure side CFD normalized tip mass flow. Static CFD predicted tip edge. Suction side CFD predicted normalized tip mass flow. It is a CFD blade tip Nusselt number distribution. CFD predicted normalized heat load at flat tip, nominal recess and new recess. 7 is an experimental relative total pressure coefficient distribution at the trailing edge of a 14% axial chord downstream rotor blade in a flat tip blade. FIG. 4 is a CFD predicted relative total pressure coefficient distribution at the trailing edge of a 14% axial chord downstream rotor blade in a flat tip blade. Stable CFD prediction at 14% axial chord downstream rotor blade trailing edge on flat tip blade vs. pitch average experimental relative yaw angle. 7 is an experimental relative total pressure coefficient distribution at the trailing edge of a 14% axial chord downstream rotor blade in a new recessed tip blade. FIG. 5 is a CFD predicted relative total pressure coefficient distribution at a trailing edge of a 14% downstream rotor blade in a new recessed tip blade. CFD prediction vs. experimental pitch average relative yaw angle at the trailing edge of the 14% axial chord downstream rotor blade in the new recessed tip blade. Experimental pitch average relative yaw angle at the trailing edge of the 14% axial chord downstream rotor blade in a new recessed tip blade. The 1st graph of the wall thickness distribution of the recessed part in the turbine front-end | tip is shown. The 2nd graph of wall thickness distribution of the crevice in the turbine tip is shown.

Claims (15)

A side surface (7) having an aerodynamic profile and an end surface (6) disposed at a mounting position in the turbine defined by a gap (12) from a casing (8) of the turbine;
In the turbine blade (1), the end face (6) comprises a recess (2) formed to act as an aerodynamic seal,
Turbine blade, characterized in that the recess (2) is defined by a side wall (4) having a wall thickness (t) that varies in the circumferential direction of the turbine blade (1).   The turbine blade according to claim 1, characterized in that the side wall (4) has a substantially constant wall thickness (t) in the height direction (D).   In the region where the turbine blade (1) has the maximum thickness (14), the wall thickness (t) of the side wall (4) has an overall maximum value (15) on the suction side of the turbine blade (1). The turbine blade according to claim 1, wherein the turbine blade is characterized in that   The wall thickness (t) of the side wall (4) is between a relative length range of + 0% to + 50% with respect to the blade leading edge (0), and the maximum value (15 The turbine blade according to any one of claims 1 to 3, characterized by comprising:   Turbine blade according to claim 4, characterized in that the maximum value (15) lies between a relative length range of + 10% to + 40% with respect to the blade leading edge (0).   Turbine blade according to claim 4 or 5, characterized in that the maximum value (15) is an overall maximum value.   Turbine blade according to claim 4 or 6, characterized in that the wall thickness (t) decreases continuously on both sides of the maximum value (15) on the suction side of the turbine blade (1).   The wall thickness (t) of the side wall (4) is a minimum value on the pressure side of the turbine blade (1) between a relative length range of −0% to −70% relative to the blade leading edge (0) The turbine blade according to claim 1, comprising:   Turbine blade according to any one of the preceding claims, characterized in that the wall thickness (t) on the suction side of the turbine blade (1) is at least partly constant.   Turbine blade according to any one of the preceding claims, characterized in that the wall thickness (t) on the pressure side of the turbine blade (1) is at least partly constant.   Turbine according to claim 10, characterized in that the wall thickness (t) on the pressure side is constant between a relative length range of -5% to -50% with respect to the blade leading edge (0). blade.   12. The overall maximum value (15) of the wall thickness (t) of the side wall (4) is about 2 to 8 times greater than the overall minimum value according to any one of the preceding claims. Turbine blades.   The turbine blade according to claim 12, characterized in that the overall maximum value (15) of the wall thickness (t) of the side wall (4) is approximately 5.5 to 6.5 times greater than the overall minimum value. .   Turbine blade according to any one of the preceding claims, characterized in that the recess (2) has a substantially constant depth (D).   Turbine having a turbine blade (1) according to any one of the preceding claims.

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