The Pennsylvania State University. The Graduate School. Department of Aerospace Engineering AERODYNAMIC AND HEAT TRANSFER ASPECTS OF TIP AND CASING

Transcription

1 The Pennsylvania State University The Graduate School Department of Aerospace Engineering AERODYNAMIC AND HEAT TRANSFER ASPECTS OF TIP AND CASING TREATMENTS USED FOR TURBINE TIP LEAKAGE CONTROL A Dissertation in Aerospace Engineering by Baris Gumusel 2008 Baris Gumusel Submitted in Partial Fulfillment of the Requirements for the Degree of Doctor of Philosophy December 2008

2 The dissertation of Baris Gumusel was reviewed and approved* by the following: Cengiz Camci Professor of Aerospace Engineering Dissertation Advisor Chair of Committee Dennis K. McLaughlin Professor of Aerospace Engineering Savas Yavuzkurt Professor of Mechanical Engineering Timothy F. Miller Senior Research Associate George S. Lesieture Professor of Aerospace Engineering Head of the Department of Aerospace Engineering *Signatures are on file in the Graduate School

3 iii ABSTRACT Axial flow turbine stages are usually designed with a gap between the tips of the rotating blades and a stationary outer casing. The presence of a strong pressure gradient across this gap drives flow from the pressure side of the blade to the suction side. This leakage flow creates a significant amount of energy loss of working fluid in the turbine stage. In a modern gas turbine engine the outer casing of the high-pressure turbine is also exposed to a combination of high flow temperatures and heat transfer coefficients. The casing is consequently subjected to high levels of convective heat transfer, a situation that is aggravated by flow unsteadiness caused by periodic blade-passing events. An experimental investigation of the aerodynamic and heat transfer effect of tip and casing treatments used in turbine tip leakage control was conducted in a large scale, low speed, rotating research turbine facility. The effects of casing treatments were investigated by measuring the total pressure field at the exit of the rotor using a high frequency response total pressure probe. A smooth wall as a baseline case was also investigated. The test cases presented include results of casing treatments with varying dimensions for tip gap height of t/h=2.5%. The results of the rotor exit total pressure indicate that the casing treatment significantly reduced the leakage mass flow rate and the momentum deficit in the core of the tip vortex. The reductions obtained in the tip vortex size and strength influenced the tip-side passage vortex and other typical core flow characteristics in the passage. Casing treatments with the highest ridge height was the most effective in reducing the total pressure loss in the leakage flow of the test blades. This was observed at a radius near the

4 iv core of the tip vortex. It appears that casing treatments with the highest ridge height is also the most effective from a global point of view, as shown by the passage averaged pressure coefficient obtained in the last 20% of the blade height. The effect of the new blade tip concept, inclined squealer tip, on tip leakage flow with and without casing treatments is also investigated. The results of the rotor exit total pressure indicate that the inclined squealer tip arrangement has significant effects on both passage core flow and the interaction between the leakage vortex and the tip side passage vortex. A steady-state method of measuring convective heat transfer coefficient on the casing of an axial flow turbine is also developed for the comparison of various casing surface and tip designs used for turbine performance improvements. The free-stream reference temperature, especially in the tip gap region of the casing varies monotonically from the rotor inlet to rotor exit due to work extraction in the stage. In a heat transfer problem of this nature, the definition of the free-stream temperature is not as straight forward as constant free-stream temperature type problems. The accurate determination of the convective heat transfer coefficient depends on the magnitude of the local freestream reference temperature varying in axial direction, from the rotor inlet to exit. The current investigation explains a strategy for the simultaneous determination of the steadystate heat transfer coefficient and free-stream reference temperature on the smooth casing of a single stage rotating turbine facility. The heat transfer approach is also applicable to casing surfaces that have surface treatments for tip leakage control. The overall

5 v uncertainty of the method developed is between 5% and 8% of the convective heat transfer coefficient. The test cases presented show that the casing heat transfer is affected by the tip gap height. The heat transfer coefficient increases as the tip gap increases for both with and without casing treatments. It is also shown that the effect of ridge height on heat transfer coefficient is negligible for tip gap height of t/h=0.9%.

6 vi TABLE OF CONTENTS LIST OF FIGURES...ix LIST OF TABLES...xiii NOMENCLATURE...xiv ACKNOWLEDGEMENTS...xvii Chapter 1 Introduction Axial Turbine Blade Terminology Turbine Rotor Passage Flow Tip Leakage Flow in Turbine Stage Tip Leakage Desensitization Turbine Casing (Shroud) Heat Transfer Objectives of Current Research...19 Chapter 2 Facility Description The Axial Flow Turbine Research Facility (AFTRF) The Turbine Stage Turbine Blade Tip Modification Instrumentation Monitoring Instrumentation Performance Measurement High-Frequency Total Pressure Probe Heat Transfer Instrumentation Data Acquisition...34 Chapter 3 Convective Heat Transfer Measurements on the Casing Inner Surface Heat Transfer Coefficient Measurement Locations Steady-state Heat Transfer Method Lateral Conduction Losses in Aluminum Casing Plate Boundary Conditions for Conduction Loss Analysis Lateral Conduction Analysis Results Heat Transfer Coefficient from Different Power Settings The Correct Free Stream Reference Temperature Direct Measurement of T aw as the Correct Free Stream Reference Temperature A Validation of T aw measurement Axial Distribution of h on the Casing Plate...53

7 vii Chapter 4 Casing Treatment Definitions AFTRF Casing Modifications Casing Surface Treatments Casing Surface Roughness...63 Chapter 5 Tip Clearance Aerodynamics Baseline Rotor Total Pressure Measurements Baseline, No Casing Treatment Effect of Blade Tip Shapes Effect of the Tip Gap Height...74 Chapter 6 Aerodynamic Influence of Casing Treatments Baseline at Tip Clearance of t/h = 2.5% Casing Treatments at Tip Clearance of t/h = 2.5% Straight Casing Treatment Panel No.S Straight Casing Treatment Panel No.S Straight Casing Treatment Panel No.S Curved Casing Treatment Panel No.C Curved Casing Treatment Panel No.C Curved Casing Treatment Panel No.C Passage to Passage C pt Variation near the Casing Comparison of the Averaged Total Pressure Coefficient Effect of Casing Surface Roughness...92 Chapter 7 Inclined Squealer Tip Squealer Tip vs. Inclined Squealer Tip Smooth Casing Treatment Curved Casing Treatment...99 Chapter 8 Convective Heat Transfer Influence of Casing Treatments Definition of Casing Surfaces for HT Measurements Blade Modifications for HT Measurements Effect of Casing Treatments Smooth Casing Surface Curved Casing Treatments Effect of Casing Surface Roughness Chapter 9 Conclusions Bibliography...117

8 Appendix A Total Pressure Probe Characteristics A.1 Angular Sensitivity A.2 Uncertainty Analysis for Dynamic Pressure Measurements Appendix B Uncertainty Analysis for Heat Transfer Measurements viii

9 ix LIST OF FIGURES Figure 1-1: Blading Terminology....3 Figure 1-2: C p Distribution on Blade Tip, From [1]....4 Figure 1-3: Turbine Rotor Passage Flow, From Yamamoto [3]...6 Figure 1-4: Tip Gap Flow Conceptual Model, From Bindon [8]....9 Figure 2-1: AFTRF Schematic...22 Figure 2-2: Rectangular Window for the Removable Casing Segment...23 Figure 2-3: Removable Turbine Casing in AFTRF Figure 2-4: Aluminum Blades Before and After the Modification...27 Figure 2-5: Modified Rotor Blades with SLA Plastic Tip...27 Figure 2-6: Location of Test Blades Figure 2-7: Schematic of Monitoring Instrumentation, From [39]...30 Figure 2-8: Schematic of Performance Instrumentation, From [39]...31 Figure 2-9: Transducer and Probe Arrangement Figure 2-10: Heat Transfer Measurement Area...33 Figure 2-11: Removable Turbine Casing Cross Section...35 Figure 3-1: Heat Transfer Coefficient Measurement Locations on the Casing Surface (Five Axial Locations)...38 Figure 3-2: Heat Transfer Model for Convective Heat Transfer Coefficient Measurements on the Turbine Casing Surface Figure 3-3: 3D Solid Model and Conduction Analysis Results on Removable Turbine Casing Surfaces...43 Figure 3-4: Temperature Distributions on the Plastic Spacer and Aluminum Casing Plate (Flow Side) Figure 3-5: Lateral Conduction from the Four Sides of the Area Facing the Heater and the Final Energy Balance (I 2 R=6.53 W)...47

10 x Figure 3-6: Energy Balance in the Heat Transfer Surface in Function of Power Setting...48 Figure 3-7: Simultaneous Determination of Convective Heat Transfer Coefficient h and Free Stream Reference Temperature from Multiple Heater Power Settings...50 Figure 3-8: Influence of Proper Free Stream Reference Temperature on Convective Heat Transfer Coefficient...51 Figure 3-9: Measured Heat Transfer Coefficient at Five Axial Locations on the Casing Surface...54 Figure 3-10: Distribution of the Heat Transfer Coefficient with respect to Axial Position on the Casing Surface...55 Figure 4-1: Casing Treatment on the Inner Surface of the AFTRF Casing...56 Figure 4-2: Smooth Casing Treatment...57 Figure 4-3: Straight Casing Treatments...58 Figure 4-4: Curved Casing Treatments...58 Figure 4-5: Solid Model of Casing Treatment with Patterns...60 Figure 4-6: The Rotor Blade Position with respect to the Casing Treatment...61 Figure 4-7: SLA Insert Flush Mounted on the Inner Side of the Removable AFTRF Window Figure 4-8: Application of Sandpaper to the Inner Surface of the Casing Window...64 Figure 5-1: Total Pressure Coefficient Contours with Flat-Tip Blades (t/h = 2.5%)..68 Figure 5-2: Total Pressure Coefficient Contours for All the 29 Rotor Passages (Baseline)...71 Figure 5-3: Total Pressure Coefficient Contours with Squealer-Tip Blades (t/h = 2.5%)...72 Figure 5-4: Effect of Blade Tip Shape on the Passage Averaged Coefficient...73 Figure 5-5: Effect of Blade Tip Shape on the Area Averaged Total Pressure Coefficient....74

11 xi Figure 5-6: Total Pressure Coefficient Contours with Squealer-Tip Blades (t/h = 1.5%)...75 Figure 5-7: Total Pressure Coefficient Contours with Squealer-Tip Blades (t/h = 0.75%)...76 Figure 5-8: Effect of Reducing the Tip Clearance of Test Blades on the Wake Profile at r = 0.96h...77 Figure 5-9: Effect of the Tip Clearance on the Passage Averaged Coefficient of Blade Figure 6-1: C pt Distribution at 30 % Downstream of AFTRF Rotor (Squealer Tips)...80 Figure 6-2: C pt Distribution at 30 % Downstream of AFTRF Rotor (S1) Figure 6-3: C pt Distribution at 30 % Downstream of AFTRF Rotor (S2) Figure 6-4: C pt Distribution at 30 % Downstream of AFTRF Rotor (S3) Figure 6-5: C pt Distribution at 30 % Downstream of AFTRF Rotor (C1)...86 Figure 6-6: C pt Distribution at 30 % Downstream of AFTRF Rotor (C2)...88 Figure 6-7: C pt Distribution at 30 % Downstream of AFTRF Rotor (C3)...89 Figure 6-8: C pt Distribution at a Fixed Radius r = 0.96h with Significant Casing Treatment Influence...90 Figure 6-9: Effect of Casing Treatments on the Passage Averaged Coefficient of Blade Figure 6-10: Rotor Averaged Coefficient with Different Casing Roughness Treatments Figure 7-1: Inclined Squealer Tip Manufactured by SLA Technique Figure 7-2: Total Pressure Coefficient Contours (Smooth Casing)...96 Figure 7-3: Total Pressure Coefficient Contours with Inclined Squealer Tips (Smooth Casing) Figure 7-4: Flow over Inclined Shelf, From [56] Figure 7-5: Total Pressure Coefficient Contours (Curved Casing Treatment)...99

12 Figure 7-6: Total Pressure Coefficient Contours with Squealer Tips (Curved Casing Treatment) Figure 7-7: Total Pressure Coefficient Contours with Inclined Squealer Tips (Curved Casing Treatment) Figure 7-8: Area Averaged Total Pressure Coefficient Figure 8-1: Removable Turbine Casing in AFTRF (with Smooth Al Casing Plate ) Figure 8-2: Tip Clearance Distribution for HT Measurements Figure 8-3: Distribution of the Heat Transfer Coefficient with respect to Axial Position on the Smooth Casing Surface Figure 8-4: Distribution of the Heat Transfer Coefficient with respect to Axial Position on the Casing Surface with Curved Patterns Figure 8-5: Effect of Pattern Heights at Small Tip Clearances Figure 8-6: Effect of Pattern Heights at Large Tip Clearances Figure 8-7: Effect of Casing Surface Roughness at Tip Gap Height of t/h = 0.9% and 2.5% Figure A-1: Kulite Probe Response to Incidence Angle Figure B-1: Influence of Reference Temperature Measurement Error δt f on δh/h δqconv/qconv = ±0.01, δk = ±0.221 W/mK, δl = ±25 μm Figure B-2: Influence of Heater Power Setting on δh/h δt H = ±0.15 K, δk = ±0.221 W/mK, δl = ±25 μm xii

13 xiii LIST OF TABLES Table 2-1: AFTRF Facility Design Performance Data...25 Table 2-2: AFTRF Stage Blade & Vane Data Table 3-1: Thermal Conductivity and Thickness Values for the Heat Transfer Surface Components...39 Table 4-1: Dimensions of Casing Surfaces...59 Table 8-1: Dimensions of Casing Surfaces for HT Measurements Table A-1: Uncertainty and Nominal Values in Measured Parameters Table A-2: Uncertainty in Derived Parameters Table B-1: Uncertainty Estimates...128

14 xiv NOMENCLATURE C ax Blade tip axial chord length, m. C p C pt C pt, R C pt, A C pt, P Specific heat at constant pressure, kj/kg.k Total pressure coefficient. Rotor averaged total pressure coefficient. Area averaged total pressure coefficient. Passage averaged total pressure coefficient. h Blade height, m. h Heat transfer coefficient, W/m 2 K. i j k N Circumferential index. Radial index. Thermal conductivity, W/mK. Rotor speed, rpm. P o, p o Total or stagnation pressure, Pa. q Heat flux, W/m 2. r Radius, m. r,θ,x Cylindrical coordinates. T Temperature, K. t Tip gap height, m. T o Total or stagnation temperature, K. U x,y,z Blade speed, m/s. Cartesian coordinates.

15 xv GREEK SYMBOLS Diameter, m. α Absolute flow angle. ρ Density, kg/m 3. SUBSCRIPTS 1 Stage inlet. 2 Nozzle exit / rotor inlet. 3 Rotor exit. amb aw f H hub RH tip w Ambient (pressure or temperature). Adiabatic wall. Fluid. Heater. Rotor hub. Rohacell. Rotor tip. Wall. ACRONYMS B(#) Blade number. EGV Exit guide vane. HP High pressure.

16 xvi N Nozzle. NGV Nozzle guide vane. OTL PS R SLA SS Over tip leakage. Blade pressure surface or pressure-side. Rotor. Stereo Lithography. Blade suction surface or suction-side.

17 xvii ACKNOWLEDGEMENTS I would like to thank my advisor, Dr. Cengiz Camci, for all of his help and guidance throughout my research work. I appreciate all of his insights into the research that I have been involved in and for all of his help in completing this work. I would also like to thank Dr. McLaughlin, Dr. Yavuzkurt, and Dr. Miller for serving on my doctoral committee and for providing me with valuable suggestions. I would like to acknowledge the valuable comments and advice provided by Dr. Raymond E. Chupp of GE Energy throughout this research. I would also like to acknowledge Dr. Nikhil Rao for all of his help. I would particularly like to thank Mark Catalano, Rick Auhl and Kirk Heller for their technical support in overcoming the technical difficulties. Additional thanks go to Harry Houtz for providing technical support in maintenance and modifications of the research facility. Very special thanks go to my wife, Elif for her love, support and encouragement. I could not have accomplished so much without her. Finally, I would like to thank my parents and my brother for all of their support through the years.

18 Chapter 1 Introduction The effects of tip clearance flows have been studied by many researchers in the past. It is known that a flow is initiated through the gap from the pressure to suction side due to the pressure difference between the pressure and suction sides of the blade. The mixing of this high momentum flow with the passage flow causes aerodynamic losses. It also exposes the tip surface to high temperatures. The situation is not helped by the fact that the tip is the hardest region to cool in a turbine blade. The tip clearance problem is particularly severe at the high pressure end of a multi-stage turbine. High pressure stages operate at higher temperatures, have a higher hub-tip ratio and shorter blades, which means that the tip clearance is a larger percent of the blade height, and do not easily permit the use of a shroud to alleviate the tip clearance problem. Many different techniques have been studied to reduce the tip leakage flows and their adverse effects. The research reported in this thesis is specific to unshrouded, high pressure turbine rotors. This chapter summarizes the current understanding of the tip leakage process, previous studies aimed at minimizing the effects of tip leakage, blade tip and casing heat transfer, and the objective and organization of this thesis.

19 1.1 Axial Turbine Blade Terminology 2 The geometric terms and parameters used to define a blade passage are shown in Figure 1.1. The blades shown in this figure is the turbine blades in the rotating turbine research rig at The Pennsylvania State University. The concave side of the blade is the pressure surface (PS), and the convex side is the suction surface (SS). A pressure distribution is generated on both PS and SS as the flow is turned in the blade passage. Detailed pressure distributions obtained on the rotating tip surfaces of turbine blades are helpful in explaining the physical aspects of tip leakage flows. A good understanding of the static pressure field on blade tips is essential in implementing new aerodynamic desensitization schemes. Tip static pressure measurements performed by Xiao [1] in AFTRF are shown in Figure 1.2. The measurements are given for an average tip gap of 0.97 mm and a relative casing speed of 64 m/s. The tip platform static holes that are distributed around the perimeter of the tip airfoil are flush with the tip platform. The pressure measured at the corner between the pressure side and tip platform is extremely low. The pressure in this zone is even lower than the suction side corner. The ultra low pressure at the corner between the pressure surface and the tip is due to the sharp radius of the pressure side corner. A small separation bubble just downstream of the pressure side corner forms a vena-contracta effect at the entrance of the tip gap area. The width of the bubble depends on tip gap size and the corner radius. The separation bubble causes significant total pressure loss. The low-pressure zone is also because of the strong acceleration of the leakage flow into the vena-contracta dominated zone of the tip flow path.

20 3 Tip Leading Edge Suction Surface Pressure Surface Hub (Root) Trailing Edge Figure 1-1: Blading Terminology

21 4 C p data at blade tip on pressure side data at blade tip on suction side x/c Figure 1-2: C p Distribution on Blade Tip, From [1] 1.2 Turbine Rotor Passage Flow The three dimensional features of the flow in a turbine rotor is given with reference to Figure 1.3. The inlet boundary layer (1) is exposed to a strong adverse pressure gradient set up by the blades as it approaches the leading edge and rolls up into a horseshoe vortex (2) which wraps around the two sides of the blade. The pressure-side leg (3) of the horseshoe vortex merges with pressure-driven cross-passage flow in the rotor endwall boundary layer (8) and becomes the passage vortex (9). The suction-side

22 5 leg (4) is drawn into an adjacent passage and wraps around the passage vortex as shown by Sharma and Butler [2]. With the presence of tip gap, the flow pattern at the casing endwall becomes more complicated. The strength of the adverse pressure gradient enforced by the blade varies with the gap height and no stagnation point exists on the casing endwall as the inlet boundary layer approaches the leading edge of the blade. As a result, the region of maximum pressure within the passage occurs slightly away from the pressure surface. On one side of this region, due to the presence of a strong pressure gradient across the tip gap drives flow from the pressure side of the blade to the suction side. Hence, the fluid is accelerated into the tip gap to form the tip leakage flow. On the other side of the region, the fluid is accelerated towards the suction side of the passage and rolls up into a tip passage vortex. The tip leakage vortex (12) and the tip passage vortex rotate in opposite directions. The interaction between these vortices generates limiting streamlines on the blade suction surface. In general, the vortices transport fluid with high momentum and temperature into the boundary layers and deposit the low momentum boundary layer fluid into the mainstream. The rotational kinetic energy is dissipated without the extraction of useful work. The interaction between vortices also leads to increased turbulence and mixing.

23 6 Figure 1-3: Turbine Rotor Passage Flow, From Yamamoto [3] 1.3 Tip Leakage Flow in Turbine Stage The tip gap, also known as the tip clearance, required between the rotating blade tip and the stationary casing is a significant source of aerodynamic inefficiency in turbine stage. The tip leakage flow driven by the pressure difference between the pressure side and the suction side of the blade tip rolls into a streamwise vortical structure along the corner of suction side and interacts with the rotor passage flow. The leakage flow mixing with the passage flow causes total pressure loss and reduces turbine stage efficiency. It is stated by Booth [4] that more than 30% of the losses in a turbine stage can be associated with the tip clearance. It is also suggested by Schaub et al. [5] that 45% of the losses in the rotor and 30% of the losses in the stage can be attributed to the tip leakage flow. Hence, it is very important to understand the physics of the tip leakage flow to reduce the

24 7 losses, and to improve the turbine efficiency. Linear cascades, composed of turbine rotor blades, have been used primarily to understand the physics and effects of the tip leakage flow. The flow characteristics of leakage flow in cascades have been described comprehensively in the open literature. Results from rotating rig experiments are limited and as such no data exists for flow within the tip gap or near tip surface flow. The first investigation of the tip leakage flow was performed to understand and identify the flow physics by Rains [6]. Tip leakage flow visualization experiments conducted by Booth et al [7] in different water flow rigs validated Rains' [6] hypothesis that viscous effects are fairly small within the tip gap. Rains [6] suggested a model for the formation of tip leakage vortex that the leakage flow in the tip gap region rolls up into a vortical structure as a result of the interaction with the passage flow. The first detailed measurement of OTL flow within the tip gap, reported by Bindon [8], conducted in a linear cascade of turbine blades. Bindon [8] studied the formation and development of tip clearance flows for a flat blade tip with sharp cornered edges. He compared the growth of total integrated loss coefficient from the leading edge to the trailing edge of a blade passage. This coefficient includes losses from secondary flow, tip leakage flow, and endwall and viscous losses. While the loss increased linearly with tip clearance gap size, the rise from zero clearance to the smallest gap size was not linear and rather dramatic. He concluded that the tip gap introduced a new loss generation mechanism. Based on measurements of velocity, total pressure, and endwall static pressures in a linear cascade Bindon [8] presented a conceptual model for tip clearance flow, as shown in Figure 1.4. This figure shows that the gap flow occurs in three channels. Channel 1 represents the inlet boundary layer and core flow which pass through

25 8 the tip gap in the front part of the blade. This flow mixes with the mainstream flow of the neighboring passage at an angle of α as shown in the figure. Channel 2 is the flow over the separation bubble f and subsequent reattachment on the tip surface. Thus, the leakage flow mixing with the mainstream is composed of a reattachment wake d and a core flow c. Finally, channel 3 that is located at approximately midchord, where the tip leakage velocity reaches a maximum. The advancing separation bubble is entrained by this high velocity fluid and ejected out to the suction corner flow as an entrainment wake e. There is intense mixing of the leakage jet and the fluid in the separation bubble. This is the primary source of internal gap losses, which amounts to about 39% of total tip leakage losses. Another 48% of the total loss is attributed to the mixing of the leakage jets and wakes with the mainstream flow. The remaining loss is caused by the endwall/secondary loss. Mixing loss appears only over the last 20% of the axial chord. Graham [9] performed experiments in the tip gap region of a linear water flow cascade with a moving belt in order to clarify the effects of tip gap heights and relative motion speeds on the tip leakage flow. He observed the typical separation bubble near the entrance section of the gap and found that the size of the separation bubble decreases as the clearance is reduced. Graham [9] also found that the leakage flow rate reduced by both increasing the relative speed and decreasing the gap height. Yaras, et al., [10] studied the flow in the tip gap of a linear turbine cascade by conducting detailed surveys of the velocity magnitude, flow direction, and total pressure within the gap, and static pressure measurements on the blade surface and the endwall. They found that the velocity is small in the chordwise direction, except close to the tip wall and is almost constant in the spanwise direction. Yaras, et al., [10] concluded that

26 9 this validated the assumption that tangential velocity is conserved across the gap, and also that the flow is completely accelerated before it enters the tip gap. The flow near the tip on the other hand shows significant chordwise variation, caused probably by the separation and reattachment of flow over the tip. Based on the endwall static pressures and uniformity of leakage flow closer to the endwall they concluded that at a given chordwise location the tip leakage flow is essentially governed by the blade to blade pressure distribution. Figure 1-4: Tip Gap Flow Conceptual Model, From Bindon [8] Linear cascades have been used to provide detailed understanding of the physics of the flow in the tip gap. However, linear cascades do not capture the effect of relative motion between turbine rotor blades and outer casing. Casing relative motion is in a direction opposite to the leakage flow and its effect would depend on the gap height.

27 10 Sjolander [11] refers to an investigation by Yaras & Sjolander [12] on the effect of relative wall motion on tip leakage flows. The relative endwall motion was simulated in a linear cascade by using a moving belt. They found that the endwall shear layer was very thin, thus suggesting that the tip leakage flow was still mostly inviscid. A reduction in the gap velocity was evident and this increased with wall velocity. They estimated that at design speed the leakage mass flow was reduced by about 50%. They concluded that the casing motion reduces the driving pressure field, thereby causing a reduction in the gap leakage. This was confirmed by Morphis & Bindon [13] through their study, which simulated the relative casing motion in an annular cascade. They measured the surface pressure in the tip gap and found it had decreased. Another effect of the wall motion was that the passage vortex was enhanced and the relative motion caused the fluid to move towards the suction side. This caused a blockage to the flow exiting the tip gap thereby, reducing the leakage vortex. Kaiser & Bindon [14], and Heyes & Hodson [15] found a region of high axial velocity near the endwall followed by a region of lower velocity. Kaiser & Bindon [14] also found that the tip clearance affects about 35% of the span from the tip. However, their main conclusion was that the gap mass flow is little influenced by relative motion. Graham [9] studied the effect of relative tip endwall motion by flow visualization. He found that at any given clearance increasing the relative speed between the tip and endwall reduces the leakage flow rate by changing the leakage flow direction and reducing the leakage layer thickness, and that the leakage vortex moves towards the suction surface. While the effect of the relative casing motion may be simulated in cascades, the development of loss mechanisms inside the rotating rig appears to be affected by both

28 11 centrifugal and coriolis forces. Yamamoto, et al [16] studied the development of loss mechanisms in turbine rotors and compared the leakage vortex location with that in a stationary cascade. They found that the rotor hub passage vortex confined the over tip leakage vortex to near the casing, possibly due to centrifugal effects in rotors. In stationary cascades the vortex responded to the spanwise pressure gradient and moved radially inwards towards the hub. Morphis and Bindon [13] compared results of the losses due to tip leakage flows downstream of a cascade and rotor. They found the secondary flows behave differently in a rotor than in a cascade. Kaiser and Bindon [14] noted that the over tip leakage flow has relatively higher energy than that found in cascades, since energy extraction from leakage flow is incomplete. Velocity measurements taken inside the AFTRF rotor passages are given in McCarter [17]. It is shown that the path of the leakage vortex along the blade passage was not as steep as that measured in cascades. The leakage losses also increased further downstream of the rotor. These results show that it is important for tip desensitization investigations to be conducted in a rotating environment. 1.4 Tip Leakage Desensitization The tip clearance required between the tip of a turbine rotor blades and the stationary casing of an axial flow turbine has been shown, in the previous sections, to be an important source of efficiency loss. The leakage flow through the tip gap caused by the pressure difference between the pressure side and suction side of the rotor blade usually rolls into a tip leakage vortex along the blade suction side. Total pressure losses

29 12 of the leakage flow at the exit of a turbine rotor stage are directly proportional with the tip gap height. The gap height is also subject to growth due to oxidation, erosion and rubbing against stationary outer casing. The effect of tip clearance on turbine efficiency, as shown in Bunker [18], is also confirmed by Ameri, et al [19]. Their numerical study reconfirmed that the variation of efficiency with tip gap height is linear. Tip clearance flows also have a severe thermal effect on the tip platform, Bindon [20]. The high velocity associated with the leakage flow leads to higher heat transfer coefficients on the tip platform, particularly in the pressure side corner. The amount of heat transferred to the tip platform is also higher since the leakage flow originates from fluid that does not generate significant work. Methods of reducing the negative effects of the tip leakage flow are generally referred to as tip desensitization. Most of the tip desensitization efforts involve geometric modifications of the blade tip in an effort to reduce the tip gap mass flow rate, thereby reducing pressure losses and thermal loads on the blade tip. One of the most common methods to reduce the flow on the tip is to use a squealer tip. The flow passing over the pressure side of the blade encounters a sudden expansion as it enters the squealer cavity, thereby generating significant total pressure loss in this re-circulatory flow zone [21]. Bindon [22] visualized flow over squealer tips using smoke in a linear cascade and found that at large gap heights the leakage flow essentially passed over the tip surface as if the surface were flat. Yang et al. [23] came to similar conclusions in using squealer tips and showed that the leakage flow rate over a squealer tip is less than that of a flat tip with the same clearance gap. However, this difference is minimal at higher gap heights. Linear cascade results of tip leakage flow on flat and squealer tips, obtained by Heyes et al [24],

30 13 show that the use of a squealer tip, especially suction-side squealers, is more effective than a flat tip in reducing leakage flows. The effects of squealer tips on turbine blade tip heat transfer have also been studied both experimentally and numerically. Metzger et al. [25] and Chyu et al. [26] performed experiments to investigate the effects of squealer tip geometry on heat transfer and losses. Metzger et al [25] showed that the heat transfer on the cavity surface was reduced compared to the flat tip; however, the heat transfer level was higher at the downstream of the cavity due to the high heat transfer in the redeveloping flow inside the cavity and the additional heat transfer area created on the side walls. They recommended the use of shallow cavities if overall heat transfer reduction on the cavity surface is the goal. Metzger et al [25] also showed that leakage flow was reduced until the depth to width ratio has reached an optimum value for a given pressure difference across the gap. Numerical simulations were conducted by Yang et al [27] to investigate the flow and heat transfer around the squealer tip blade. They found that squealer tip performed better than flat tip in reducing heat transfer and leakage flow. Yang et al [27] also studied the effect of tip gap height and showed that the leakage rate increases with increasing gap height, and is lower for the squealer tip. The effective reduction in leakage mass flow rate and tip heat transfer coefficient due to the squealer tip is getting lower with increase in gap height. Other concepts that have been reported in available literature include the use of partial squealers, addition of tip winglets that are also known as tip platform extensions, modification of the tip blade profile, tip chamfering and tip injection.

31 14 Dey et al. [28, 29, 30, 31, 32] experimentally evaluated the effects of partial squealers and tip platform extensions in a large scale rotating rig at the Pennsylvania State University. Pressure-side extensions were found to be more effective than the suction-side extensions in weakening the tip vortex forming near the suction-side corner of the blade tip. Booth et al. [7] came to similar conclusions in using blade tip extensions or winglets. According to Dey et al. [31, 32], partial squealer rims applied near the suction-side of turbine blade tips were more effective than pressure-side squealers in recovering the total pressure in the core of the tip vortex. An important conclusion in [28] similar to that of Yang et al. [23], was that squealer performance is best at small tip clearances. Computational investigation of tip leakage flow around partial squealer rims by Kavurmacioglu et al. [33] indicate that a partial squealer rim arrangement can be extremely effective in weakening the tip leakage vortex. Bindon and Morphis [34] experimented with pressure-side corner rounding in an effort to reduce separation related losses in the tip gap. In the 1% - 2% gap height range the total-total stage efficiency of a single stage was found to improve with rounding-off. The efficiency of the second stator row was found to improve with rounding-off of the pressure-side corner of the first stage blades. The effects of radiused pressure-side edge and camber-line strip were numerically investigated by Ameri [35]. It was found that the sharp edge tip was more successful, than the radiused edge, in reducing the tip leakage. The mean camber-line strip reduced the leakage mass flow when implemented with a radiused edge. However, the authors note that the reduction in total pressure loss was not proportionate. An implementation of chord-wise sealing strips on a turbine tip was also

32 15 presented by Bunker et al. [36]. The tip surface heat transfer was found to increase due to rounding-off of the pressure-side edge. Tallman et al. [37] computationally investigated a number of turbine tip geometry modifications. The modifications consisted of chamfering the tip gap region from pressure-side to suction-side, thus creating a diffusing section in the leakage flow path. Chamfering near the leading edge led to increased leakage vortex activity and losses, while chamfering near the trailing edge reduced the vortex size. This was attributed to turning of gap flow towards the camber-line. Chamfering the entire length of the gap caused higher losses due to larger vortex. Dey et al. [38] reported a preliminary investigation of tip desensitization by coolant injection from a tip trench in a large scale rotating rig. No measurable effect on the tip leakage flow was obtained and this was attributed to the coolant injection holes being too small. Rao [39] experimentally studied the effect of tip coolant injection on tip leakage flow in a rotating environment. The configuration used in Dey et al. [38] was modified to allow for greater coolant mass flow rates. The measurements with and without coolant injection are analyzed for performance benefits. An influence of the injection rate on the radial position of the tip leakage vortex and its strength were detected. The results showed that tip coolant injection can be effective in creating beneficial effects previously observed only from small clearances.

33 1.5 Turbine Casing (Shroud) Heat Transfer 16 Convective heat transfer to the static casing of a shroudless HP turbine rotor is a complex aero-thermal problem. The unsteady flow with a relatively high Reynolds number in the tip gap region has strong dependency on the tip clearance gap, blade tip profile; tip loading conditions, tip geometry and casing surface character. Thermal transport by flow near the casing inner surface is influenced by the unsteadiness, the surface roughness character and the turbulent flow characteristics of the fluid entering into the region between the tip and casing. Since the turbine inlet temperatures are continuously elevated to higher levels, casing and tip related heat transfer issues are becoming more critical in design studies. In gas turbines, the gas stream leaving the combustor is not at a uniform temperature in radial and circumferential directions. According to Butler et al [40] the combustor exit maximum temperature can be twice as high as the minimum temperature. The maximum temperature in general is around the mid-span and the lowest gas temperatures are near the walls. The mechanisms related to the distortion of the radial temperature profile as the combustor exit fluid passes through a turbine rotor are complex, as explained by Sharma and Stetson [41] and Harvey [42]. The hottest part of the fluid leaving the upstream nozzle guide vane tends to migrate to the rotor tip corner near the mid pressure surface of the blade. Unfortunately, mostly the hottest fluid originating from the mid span region of the combustor or NGV finds its way to the pressure side corner of the blade tip in the rotating frame. Details of hot streak migration

34 17 in gas turbines can be found in Roback&Dring [43, 44], Takanashi&Ni [45], Dorney et al. [46] and Dorney and Schwab [47]. Due to significant energy extraction in a HP turbine stage, rotor absolute total temperature monotonically decreases in axial direction at a significant rate. This is especially true at the core of the blade passage where most of the energy extraction takes place. However, the fluid finding its way to the area between the casing and blade tips do not participate in the work generation as much as the mid-span fluid. Therefore, it is reasonable to accept that the near-casing fluid does not cool as much as the mid-span fluid when it progresses from rotor inlet to exit. Yoshino [48] and Thorpe et al. [49] have shown that a rotor blade can also perform work on the fluid near the casing surface by means of rotor compressive heating. They obtained time-accurate and phase-locked casing heat flux measurements in Oxford Rotor Facility to show the casing heat loads as the rotor blades move relative to the static casing. They observed a very high heat flux zone on the casing inner surface for each blade in the rotor. Phase-locked measurements clearly indicated that the hot spot moved along the casing with the rotor. The results of this study showed two distinct levels of casing heating, one level for the casing interaction with the tip leakage fluid and a relatively low level for the casing interaction with the passage fluid located between blade tips. Thorpe et al. [49] explained the high heat flux zone on the casing surface by a rotor compressive heating model. They showed that the static pressure field near the tip can do work on the leakage fluid trapped between the blade tip and the casing. The rotor compressive heating model predicts that the absolute total temperature of the leakage fluid may exceed that of the rotor inlet flow. The flow near the casing turns and

35 18 accelerates the leakage fluid to a tangential velocity level that is measurably above the rotor inlet level. Thorpe et al. [50] were successful in predicting the total temperature penalty due to compressive heating using the Euler work equation. Any design effort that will reduce the tip leakage mass flow rate in an axial turbine will also result in the reduction of the total temperature penalty and a corresponding reduction in casing heat load. Past studies show three significant contributors to casing heat loads in shroudless HP turbines. 1. Radially outward and axial migration of a hot streak in each passage results in the accumulation of relatively high temperature fluid near the pressure side corner of the blade tip before it enters the tip gap. 2. A relatively higher total temperature in the near-casing fluid is observed because the near casing fluid does not participate in stage work generation. The leakage fluid does not expand as much as the core-flow in the rotor passage. 3. Rotor compressive heating performed by blade tips when they move against the static casing is significant. The near-casing gas temperature drops at a significant rate in axial direction. There is also a strong circumferential mixing near the casing because of the relative motion of blade tips. The time accurate wall heat flux measured on the casing vary between a passage gas induced low value and tip leakage fluid induced high value. Since near-casing gas temperatures vary at a significant rate in axial direction, any heat transfer measurement approach requires the simultaneous measurement of this local gas

36 19 temperature in the vicinity of the casing, in addition to an accurate determination of convective heat transfer coefficient. A steady state casing heat/mass transfer coefficient measurement method based on a Naphthalene sublimation technique is explained in Rhee and Cho [51]. They report similar casing heat transfer distributions with and without blade rotation. Their sublimation based heat transfer method is inherently intrusive because of the variations of the Naphthalene layer thickness imposed by local mass transfer rate variations in the tip gap region. 1.6 Objectives of Current Research Axial flow turbine stages are usually designed with a gap between the tips of the rotating blades and a stationary outer casing. The presence of a strong pressure gradient across this gap drives flow from the pressure side of the blade to the suction side. This leakage flow creates a significant amount of energy loss of working fluid in the turbine stage. In a modern gas turbine engine the outer casing of the high-pressure turbine is also exposed to a combination of high flow temperatures and heat transfer coefficients. The casing is consequently subjected to high levels of convective heat transfer, a situation that is aggravated by flow unsteadiness caused by periodic blade-passing events. The objectives of this experimental investigation are to study the aerodynamic and heat transfer effect of casing treatments used in turbine tip leakage control and to develop an accurate steady-state heat transfer method for the comparison of various casing surface and tip designs used for turbine performance improvements. The

37 20 aerodynamic influence of casing treatments were studied by measuring the total pressure field at the exit of the rotor using a high frequency response total pressure probe. The heat transfer method is developed for the simultaneous determination of the steady-state heat transfer coefficient and free-stream reference temperature on the smooth casing of a single stage rotating turbine facility. This research study is unique in documenting both aerodynamic and heat transfer effects of casing treatments used for tip leakage control. Time accurate measurements provided valuable aerodynamic information quantifying the near tip flow modifications imposed by casing treatments. The heat transfer method developed in this research presents a non-intrusive measurement approach which is highly effective in reducing the heat transfer measurement uncertainty. This research study is also unique in documenting the experimental results of a new squealer blade tip concept. The blade tip investigated is referred to as an inclined squealer tip. The organization of this thesis is as follows. Chapter 2 describes the facility, aerodynamic design of the turbine stage, instrumentation, and operation of the facility. Chapter 3 describes the convective heat transfer measurements on the casing inner surface. Casing Treatment Definitions are reported in Chapter 4. The results of the tip clearance aerodynamics for the baseline are presented in Chapter 5. Aerodynamic Influence of Casing Treatments is the subject of Chapter 6, where the results of total pressure measurements are presented. Chapter 7 discusses the inclined squealer tip concept. Chapter 8 discusses the convective heat transfer influence of casing treatments. A summary of the results, conclusions and future work recommendations are presented in Chapter 9.

38 Chapter 2 Facility Description The facility used in this investigation is the Axial Flow Turbine Research Facility (AFTRF) at the Turbomachinery Aero-Heat Transfer Laboratory of the Pennsylvania State University. This chapter describes the facility, turbine blades and instruments used for the current research effort. 2.1 The Axial Flow Turbine Research Facility (AFTRF) The AFTRF is a single-stage, low speed, large-scale, cold flow turbine facility with many characteristics of modern high-pressure turbines. A schematic of the facility is shown in Figure 2.1. The facility consists of a large bell-mouth inlet, followed by a test section with a nozzle vane row (N) and a rotor (R). There are 23 nozzle guide vanes and 29 rotor blades followed by a set of exit guide vanes (EGV) which direct the flow out of the rotor to a more axial direction. Measurements in AFTRF are taken either in the stationary or in the rotational frame. Stationary frame measurements utilize a rectangular window located by the side of the turbine outer casing, as shown in Figure 2.2. The rectangular window is used to house the removable casing segment. This segment is a precision machined area designed for many different aero-thermal measurement techniques to be applied around the turbine

39 22 stage. Figure 2.3 shows the removable segment with casing plate which is described in Chapter 5. Rotational frame measurements utilize a rotating instrumentation drum which is mounted on the turbine disk, as shown in Figure 2.1. Instruments, such as pressure transducers, hot wire anemometers, etc, are installed inside the instrumentation drum. The data collected inside the instrumentation drum is carried to the stationary frame through a 150 channel slip ring assembly. Airflow through the facility is provided by two auxiliary, adjustable pitch, axial flow fans. The two fans produce a combined pressure rise of 74.7 mm Hg (approximately 40 water column) with a volumetric flow of 10 m 3 s under nominal operating conditions. The power generated by the rotor assembly is absorbed by an eddy current brake which is also used to control the rotational speed of the rotor. The eddy current brake is cooled by a closed loop chilled water cooling system. Air energized by the fans is then exhausted to atmosphere through the diffuser, which has an external moving endplate that provides a means to control mass flow rate. Figure 2-1: AFTRF Schematic

40 23 Rotating hub NGV Total Pressure Probe Figure 2-2: Rectangular Window for the Removable Casing Segment

41 24 Total Pressure Probe Casing Plate Facing the Rotor Tip Radial traverse driven by a stepper motor (not shown) Probe stem Figure 2-3: Removable Turbine Casing in AFTRF

42 2.2 The Turbine Stage 25 Some of the important design performance data is listed in Table 2.1, while Table 2.2 lists important blade design parameters, including reaction at blade hub and tip sections, Reynolds number at rotor exit and a few blade parameters. Measured/design values of rotor inlet flow conditions including radial, axial, tangential components and data acquisition details of the turbine rig are explained in detail by Camci [52]. The turbine generates about 60 kw of power while operating at a nominal speed of 1300 rpm and a nominal through flow rate of kg/s. The design total-total isentropic efficiency of the turbine stage is Low inlet to exit pressure and temperature ratios, given in Table 2.1, implies that compressibility effects are insignificant. Table 2-1: AFTRF Facility Design Performance Data Inlet Total Temperature;T o1 (K) 289 Inlet Total Pressure;P o1 (kpa) Mass Flow Rate;Q (kg/sec) Rotational Speed; N (rpm) 1300 Total Pressure Ratio; P o1 /P o Total Temperature Ratio; T o3 /T o Pressure Drop; P o1 -P o3 (mm Hg) Power; P (kw) 60.6

43 26 Table 2-2: AFTRF Stage Blade & Vane Data Rotor hub-tip ratio Blade Tip Radius; R tip (m) Blade Height; h (m) Relative Mach Number 0.24 Number of Blades 29 Axial Tip Chord; (m) Spacing; (m) Turning Angle; Tip / Hub o / o Nominal Tip Clearance; (mm) 0.9 Reaction, Hub / Tip / Reynolds Number ( 10 5 ) inlet / exit (2.5~4.5) / (5~7) 2.3 Turbine Blade Tip Modification The rotor blades were designed to represent current blade designs in high pressure turbines, by General Electric. All blades were designed to have a nominal tip clearance of 0.9 mm (t/h=0.72%). Six of the existing blades, referred to as test blades, were modified to install new tip designs in the rotor. Stereo Lithography (SLA) was the method of manufacturing for the new tip inserts to be tested in AFTRF. The near tip section of these six Aluminum blades had been cut off by EDM (Electrical Discharge Machining) wire cutting technique, shown in Figure 2.4. Figure 2.5 shows the modified rotor blade with SLA plastic tip. SLA tips were made to test larger clearances of 2.5% (B24, B25 and B26) and 1.5% (B9, B10 and B11) blade height, shown in Figure 2.6.

44 27 Figure 2-4: Aluminum Blades Before and After the Modification Figure 2-5: Modified Rotor Blades with SLA Plastic Tip

45 Figure 2-6: Location of Test Blades

46 2.4 Instrumentation 29 Instruments used for monitoring the performance parameters consist of thermocouples, pitot-static probes and total pressure probes. Figure 2.7 and 2.8 show the schematics of instrumentation and measurement system used in current research effort. The monitoring system shown in Figure 2.7 is used to control test conditions of the facility. The performance measurement system shown in Figure 2.8 is used to assess the effect of the implemented desensitization methods on tip leakage flow by measuring the total pressure downstream of the rotor exit plane. AFTRF operational characteristics, data acquisition/processing details and other instrumentation characteristics can be found in Dey, Rao and Camci [28, 21] Monitoring Instrumentation The measurements for monitoring test conditions consist of pressures and temperatures at stage inlet and exit, bearing temperature, ambient conditions, and shaft output, as shown in Figure 2.7. A Pitot-static probe and a total temperature probe with K-type thermocouple installed in the bell-mouth inlet and located about 1.5 vane tip axial chord lengths upstream of the nozzle are used to measure turbine inlet flow conditions. The Pitot-static probe is connected to pressure transducers to measure inlet total pressure and inlet dynamic pressure, as shown schematically in Figure 2.7. Total temperature probe with K- type thermocouple, connected to Omega 650 temperature reader, is used to measure inlet

47 30 total temperature which is used to control turbine operating speed. Similarly, the rotor exit total pressure, exit dynamic pressure, and total temperature are measured using a Pitot-static probe and K-type thermocouple installed 0.6 rotor tip axial chord length downstream of the rotor. The pressure transducers used in conjunction with the Pitotstatic probes are a Validyne transducer and Honeywell transducers. Bearing temperatures are monitored using K-type thermocouples on the bearing housings. Shaft power and torque are measured by an inline torque meter connected to a display. The power and torque readings are not connected to the data acquisition system and are noted down. [39] Inlet Pitot- Static Probe Exit Pitot- Static Probe Inlet Thermocouple Room Thermocouple Exit Thermocouple Bearing Thermocouple 6-Channel Pressure Transducer OMEGA Thermocouple Display Units Thermocouple Signal Conditioning Unit P 01 T 03 P 03 T room T brg ρv 1 2 / 2 T 01 ρv 3 2 / 2 V 2ax / U m To High Speed DAS Himmelstein Torque Meter + Display Unit ANALOG INPUTS TIMING I/O CONNECTIONS LOW SPEED DAS ANALOG OUTPUT DIGITAL OUTPUT Power + Torque 1 / rev. Trigger Pulse Encoder Signal Unit 6000 / rev. Clock Pulse Figure 2-7: Schematic of Monitoring Instrumentation, From [39]

48 Performance Measurement The aerodynamic influence of casing treatments on over tip leakage flow is assessed by phase-locked measurements of the absolute total pressure at the turbine stage exit using a high-frequency-response total pressure probe and inlet pressure probe. The signals from these instruments are connected to a high speed data acquisition system, as shown in Figure 2.8. Total Pressure Probe with Kulite Sensor Inlet Pitot-Static Probe Encoder Signal Unit Kulite Power & Signal Unit VALIDYNE Pressure Transducer HONEYWELL Pressure Transducer Analog Signal From Low-Speed DAQ KRONHITE Filter P 03 P 01 ρv 1 2 / 2 ANALOG INPUTS V2ax / U m 1 / rev. Trigger Pulse 6000 / rev. Clock Pulse TIMING I/O CONNECTIONS HIGH SPEED DAS DIGITAL OUTPUT Stepper Motor Controller Figure 2-8: Schematic of Performance Instrumentation, From [39]

49 High-Frequency Total Pressure Probe Phase-locked stage exit total pressures are measured using a Kulite dynamic pressure transducer XCS-062-5D. The transducer element has a range of 34.5 kpa (5 psi) and a frequency response of 150 khz. The sensor is a sealed cylindrical tube of diameter 1.6 mm and operates in a differential mode. Reference side of the differential pressure transducer is connected to ambient pressure. The sensor is mounted with a screen and is known to have a flat magnitude and phase response up to 20 khz. The Kulite transducer is flush mounted into a square cut cylindrical probe of diameter 3.5 mm, as shown in Figure 2.9, for eliminating the time response canceling detrimental effects of a cavity with finite volume. The square face gives the probe an acceptance angle of ±10 degrees. The details of the influence of flow incidence on the total pressure measurements from the Kulite probe is presented in Appendix A. The probe is located 30% chord downstream of the rotor exit plane and is oriented to the absolute velocity vector at the tip (25.4 degrees CCW from axial). The probe is mounted in a computer/stepper motor controlled radial traverse system, shown in Figure 2.3. This allows the probe to be traversed with increments of 1/16th inch and greater. Input/output wire Reference pressure tube Flush mounted Kulite Transducer Figure 2-9: Transducer and Probe Arrangement

50 Heat Transfer Instrumentation Instruments used for heat transfer measurements consist of total pressure probes, Kiel probes, pitot-static probes, thermocouples, and constant heat flux heater. The removable turbine casing is machined so that the back side of the casing plate is instrumented with a heat transfer surface. Figure 2.10 shows the rectangular cut used for the heat transfer surface. The position of the casing plate with respect to the AFTRF casing is accurately determined. Heat Transfer Measurement Area Casing Plate Facing the Rotor Tip Figure 2-10: Heat Transfer Measurement Area The casing plate that is facing the rotor tip and interacting with near-casing fluid is shown in Figures 2.10 and The casing plate could easily be replaced with custom made plates having special casing treatments for tip vortex aerodynamic de-sensitization and supporting heat transfer studies. The removable turbine casing and the casing plate are carefully designed and precision machined so that many subsequent installations of

51 34 the same casing plate and the removable window/casing result in a repeatable tip clearance. Tip clearance repeatability within ±25 µm (± inch) for a blade height of 125 mm (4.85 inches) is possible. This uncertainty corresponds to a change in nondimensional tip clearance of ±0.02 % of the blade height. Under normal circumstances, the inserted casing plate is supposed to be flush with the static casing of the facility. Slight clearance adjustments are possible for the removable segment by altering the thickness of the plastic insulator as shown in Figure Figure 2.11 also shows the five heat transfer coefficient measurement locations which is discussed in Chapter Data Acquisition Data acquisition system is controlled by a virtual instrument (VI) setup using LabView. The speed is measured using the 6000 pulses per revolution signal generated by the shaft encoder. Data is acquired and averaged over a minute. The traverse mechanism and high speed data acquisition is controlled by a personal computer and a National Instruments PCI-6110E interface. This board is capable of simultaneously sampling its four channels at a maximum rate of 1.25 Ms/sec. The data acquisition is initiated by a trigger pulse (once per revolution) and is subsequently controlled by a clock pulse that has 6000 pulses per revolution. A BEI optical shaft encoder, Model H25 is mounted on the turbine shaft. The data acquisition is thus phase-locked and is conducted at a frequency of 132 khz (6000 points per revolution). Frequency spectra at various locations show three peaks, at the blade passing frequency and its harmonics.

52 Figure 2-11: Removable Turbine Casing Cross Section 35

53 36 Additionally, the energy contained in the signal at frequencies above 12 khz is 20 db less than that at the 2 BPF and 30 db less than that at the BPF. Hence, the signal was low-pass filtered at 20 khz. At each radial location 200 ensembles of the rotor exit absolute total pressure are acquired and averaged. The probe is radially moved in steps of 1/16 inch.

54 Chapter 3 Convective Heat Transfer Measurements on the Casing Inner Surface This chapter explains a steady-state method of measuring convective heat transfer coefficient on the casing of an axial flow turbine. The goal is to develop an accurate steady-state heat transfer method for the comparison of various casing surface and tip designs used for turbine performance improvements. The free-stream reference temperature, especially in the tip gap region of the casing varies monotonically from the rotor inlet to rotor exit due to work extraction in the stage. In a heat transfer problem of this nature, the definition of the free-stream temperature is not as straight forward as constant free-stream temperature type problems. The accurate determination of the convective heat transfer coefficient depends on the magnitude of the local free-stream reference temperature varying in axial direction, from the rotor inlet to exit. The current study explains a strategy for the simultaneous determination of the steady-state heat transfer coefficient and free-stream reference temperature on the smooth casing of a single stage rotating turbine facility. A sample calculation of the uncertainty analysis is shown in Appendix B. 3.1 Heat Transfer Coefficient Measurement Locations The five convective heat transfer coefficient measurement locations are shown in Figure 3.1. Location 1 is closest to the leading edge of the blade in axial direction. The

55 38 five selected measurement locations cover the axial distance between the blade leading edge and slightly downstream of the trailing edge. Due to the rotation of the blade, the steady-state heat transfer coefficient distribution in circumferential direction is reasonably uniform. Since work is extracted in the rotor, the free-stream total temperature between the rotor inlet and rotor exit are different. Free-stream total temperature measurement locations at the turbine inlet, rotor inlet, rotor exit and turbine exit are also shown in Figure 3.1. Turbine Exit Rotor Exit Rotor Inlet Turbine Inlet THERMOCOUPLE Figure 3-1: Heat Transfer Coefficient Measurement Locations on the Casing Surface (Five Axial Locations) The free-stream total temperatures at turbine inlet and exit are measured using calibrated K type thermocouples in a Kiel probe arrangement. Rotor inlet and exit

56 39 thermocouples are inserted into the flow at about 25 mm away from the casing surface. The free-stream total temperatures at turbine inlet and exit are measured using calibrated K type thermocouples in a Kiel probe arrangement. Rotor inlet and exit thermocouples are inserted into the flow at about 25 mm away from the casing surface. 3.2 Steady-state Heat Transfer Method Casing convective heat transfer coefficients and corresponding free-stream reference temperatures are measured simultaneously with the help of a constant heat flux heater as shown in Figure 3.2. A constant heat flux heater (MINCO Corp. HK5175R176L12B) with an effective area of (A = 76x127 mm 2 ) is sandwiched between two thin Mylar sheets. The heater can produce a maximum of 75 Watts with an overall resistance of 176 ohms. The overall resistance of the 0.5 mm thick heater has extremely small temperature dependency in the range of the current experiments. This resistance value is continuously measured and recorded during each measurement. The Joule heating value in the heater is I 2 R/A [Watts/m 2 ]. The heat transfer surface has many flat ribbon thermocouples of type K imbedded at many locations (symbol in Figure 3.2). Table 3-1: Thermal Conductivity and Thickness Values for the Heat Transfer Surface Components MATERIAL Thermal Conductivity Thickness [W/m-K] [inch] ROHACELL ALUMINUM PLATE PLASTIC LAYER HEATER (Minco)

57 40 lateral conduction losses in Al plate ROHACELL INSULATOR LAYER ROHACELL INSULATOR LAYER Constant heat flux heater T H I R A qloss TH T = LRH k RH RH 2 I R V. I = A A L Ak Al Al T H T H T w near-casing flow T H T f T q loss = = L Al 1 + k h 2 w q conv = 1 T f q conv Aluminum casing plate Al T f turbine tip region flow T RH h T f lateral conduction losses in Al plate T R, inlet Blade tip rotating at 1330 rpm T R, exit Figure 3-2: Heat Transfer Model for Convective Heat Transfer Coefficient Measurements on the Turbine Casing Surface The flat thermocouple junctions are 12 μm thick. There are two thick layers of Rohacell insulating material flush mounted on top of the heater surface. Table 3.1 includes the material thicknesses in the heat transfer composite surface and thermal conductivity values. Current multi-dimensional heat conduction analysis shows that the lateral conduction of thermal energy at the edges of the heater surface and Rohacell

58 insulator are extremely small and negligible. However the heat conduction loss q loss in Rohacell is not negligible in a direction normal to the heater. 41 q loss ( T T ) ( L k ) = (3-1) H RH RH RH Due to extremely thin structure and the uniform internal heat generation by Joule heating in the volume of the heater, measured top and bottom surface temperatures T H are very close to each other. The amount of heat flux conducted through the aluminum casing plate in a direction normal to the plate is q conv ( T T ) ( L k + 1 h) = ( T T ) ( 1 h) 2 = I R A q = (3-2) loss H f Al Al w f Equation 3-2 shows convective heat flux crossing the fluid-solid interface on the casing surface when the lateral conduction losses in Al plate are ignored. Equation 3-2 also presents the convective heat transfer rate written between the heater T H and the nearcasing turbine fluid T f. In this approach, a heat transfer coefficient h can be measured without measuring the wall temperature T w directly. h is first calculated from Equation 3-2 between T H and T f. The corresponding wall temperature T w is then obtained from the last part of Equation 3-2 written between the wall and free stream. T w H conv ( L k ) = T q (3-3) Al Al The specific experimental approach in finding T w is useful since a non-intrusive wall temperature measurement at the fluid-casing interface is essential for this problem. A more accurate form of h can be obtained by quantifying lateral conduction losses in the

59 42 aluminum plate. A three dimensional conduction heat transfer analysis including all complex geometrical features of the removable turbine casing is presented in the next few paragraphs. This computational effort reduces the measurement uncertainties in the measured convective heat transfer coefficient h. The correction is based on calculating the lateral conduction losses in the casing plate. 3.3 Lateral Conduction Losses in Aluminum Casing Plate Figure 3.3 presents the results from a 3D heat conduction analysis for the removable turbine casing. The removable turbine casing is an extremely thick precision machined aluminum with an average thickness of 50.8 mm (about 2 inches). The lateral heat conduction in the aluminum casing plate in this experiment was deduced from a 3D heat conduction analysis performed under realistic thermal boundary conditions. The steady-state thermal conduction equation 2 T( x, y,z) = 0 was solved in the removable turbine casing with proper thermal boundary conditions. The constant heat flux heater shown in Figure 3.2 is operated at a prescribed power I 2 R [W] value. Joule heating in the heater was simulated by distributing this I 2 R value uniformly over the volume of the thin heater as an internal heat generation term. This is achieved by adding a source term to the energy equation in the numerical solution procedure.

60 43 (a) (b) Figure 3-3: 3D Solid Model and Conduction Analysis Results on Removable Turbine Casing Surfaces

61 3.4 Boundary Conditions for Conduction Loss Analysis 44 On the turbine flow side, the surface temperature upstream of the rotor leading edge is taken as the measured turbine rotor inlet temperature (or NGV exit temperature). The flow side surface temperature downstream of the rotor trailing edge is the same as the measured rotor exit temperature. The flush mounted aluminum casing plate has a convective type boundary condition on the flow side where a typical heat transfer coefficient h and a free-stream reference temperature is specified at five axial positions. Measured ambient temperature outside the rig is specified on the external flat face of the removable turbine casing. All other boundaries on the side walls were taken as adiabatic. The heater surface area is about the same area as that of the small rectangular cut shown in Figure 3.3.a. 3.5 Lateral Conduction Analysis Results The temperature distribution on the Al casing plate as shown in Figure 3.3 facing the rotating blade tips is characterized by the red zone on top of the heater area. Along the dashed line in the measurement area (in axial direction) the temperature distribution is reasonably uniform within 0.5 K. The minimum and maximum temperatures in Figures 3.3 and 3.4 are 310 K (blue) and 317 K (red). All red hues are approximately corresponding to an area of 1 K temperature band.

62 45 (a) (b) Figure 3-4: Temperature Distributions on the Plastic Spacer and Aluminum Casing Plate (Flow Side)

63 46 Figure 3.4.a shows the plastic spacer/insulator inserted between the removable turbine casing shown in Figure 3.3.a and the inch thick aluminum plate. The heater is flush mounted on the convex side of the aluminum plate. Two layers of Rohacell insulator were located on top of the heater surface in order to reduce the heat losses to the ambient from the heater as shown in Figure 3.2. The low conductivity plastic spacer is essential in this measurement approach in reducing the heat losses in the inch thick casing plate. Figure 3.5 presents the lateral conduction heat losses from the aluminum casing plate. Q ABCD = (Q A +Q B +Q C +Q D ) is the sum of all thermal energy (in Watts) laterally conducted from the rectangular area where the heater is flush mounted to the casing plate. Q RH is the heat loss through the Rohacell insulator and Q HS is the heat loss from the extremely narrow side faces of the heater volume with a thickness of 0.5 mm. A new heat loss calculation was performed for each power setting using the 3D conduction analysis with proper boundary conditions. Figure 3.5 shows that the heat losses through the Rohacell layer and from the sides of the thin heater are extremely small when compared to the lateral conduction in the aluminum plate and the convective heat flux to the near-casing fluid in the turbine passage. A proper correction of convection heat flux term q conv in Equation 3-2 using lateral conduction losses is an important part in obtaining heat transfer coefficient on the turbine casing surface. Figure 3.6 presents the variation of convective heat flow over area A [q conv.a], lateral conduction loss Q ABCD in aluminum casing plate, Rohacell layer losses and heater side losses as a function of heater power setting I 2 R. Rohacell layer heat losses and heater side losses are negligible when compared to the convective heat flux and lateral conduction loss.

64 47 Figure 3-5: Lateral Conduction from the Four Sides of the Area Facing the Heater and the Final Energy Balance (I 2 R=6.53 W)

65 Figure 3-6: Energy Balance in the Heat Transfer Surface in Function of Power Setting 48

66 3.6 Heat Transfer Coefficient from Different Power Settings 49 The convective heat transfer coefficient is measured by using an arranged form of Equation 3-2, h = conv Al (3-4) k Al q k ( TH Tf ) L Alq conv where q conv is obtained by subtracting the predicted heat loss q loss from I 2 R/A as shown in Equation 3-2. T H is measured from a flush mounted thermocouple imbedded between the heater and the Al casing plate as shown in Figure 3.2. A non-intrusive measurement of T w is also available from Equation The Correct Free Stream Reference Temperature Finding the most accurate value of the reference fluid temperature T f in this problem is crucial. T f is the reference free-stream fluid temperature in the immediate vicinity of the casing surface facing the blade tips. Since this temperature in a turbine rotor monotonically decreases from rotor inlet to exit, a linear curve fit is obtained from the measured rotor inlet T R,inlet and rotor exit T R,exit as shown in Figure 3.2. The two measurement thermocouples for T R,inlet and T R,exit are inserted into the free-stream before and after the rotor. The junctions are in the turbine flow at a location about 25 mm away from the casing surface. The open circular symbols shown in Figure 3.7 form an h measurement at axial location 1. A line fit passing from all circular symbols obtained from many different power settings is represented by q conv = h (T w -T f ) for the same

67 50 turbine operating point. Maximum attention is paid to keep the corrected speed of the turbine facility and flow coefficient constant during the acquisition of all points at different heater power settings. The slope of this straight line is the convective heat transfer coefficient h. Figure 3-7: Simultaneous Determination of Convective Heat Transfer Coefficient h and Free Stream Reference Temperature from Multiple Heater Power Settings The T f value measured in the turbine free stream flow at this stage is not a proper reference temperature for this convective heat transfer problem. Since the thermocouples providing T f are inserted well into the free stream (25 mm away from the casing) the measured local T f are considerably different from the temperature of the fluid in the

68 51 immediate vicinity of the casing. This observation is consistent with the data shown with circular symbols in Figure 3.7. The solid line connecting the circular symbols does not pass through the origin as q conv = h (T w -T f ) suggests. This is a clear indication of the fact that the initially measured T f (defined by T R,inlet and T R,exit ) is not proper for the casing heat transfer problem. What happens when the solid line does not pass through is directly related to the observation made in Figure 3.8. The computation of h at many different power settings at the same turbine flow condition does not yield to an invariant h. The open triangular symbols suggest that h varies strongly with increased heater power setting. Figure 3-8: Influence of Proper Free Stream Reference Temperature on Convective Heat Transfer Coefficient

69 3.8 Direct Measurement of T aw as the Correct Free Stream Reference Temperature 52 Using multiple heater power settings allows a simultaneous measurement of h and T aw. h and T aw = T f as two unknowns of q conv =h (T w -T f ) could be obtained from two independent measurement points obtained at two different power settings. Having many more points than two and using a first order line fit only reduces the experimental uncertainties in this process. The true reference temperature in this problem (adiabatic wall temperature T aw ) is obtained by shifting the original solid line to the left until it passes through the origin. The amount of this horizontal shift is the correction to be applied to initially suggested T f. The corrected value of T f is actually the same as the actual adiabatic wall temperature T aw. 3.9 A Validation of T aw measurement A useful check on the value of h obtained from the measured reference temperature T aw is shown in Figure 3.8. When h is calculated by using measured T aw at many different power settings a constant level of h is obtained as shown by solid circular symbols. This constant level of h within experimental uncertainty indicates that the measured T aw is the proper reference temperature for this convective heat transfer problem. The procedure described in this section has the ability to measure the heat transfer coefficient h and free-stream reference temperature T f = T aw simultaneously in a non-intrusive way. A direct measurement of T aw in the turbine in the tip gap region by an inserted probe is extremely difficult. Another complexity is that the reference

70 53 temperature continually drops in the turbine due to the work extraction process gradually building up in the in the axial direction. Figure 3.8 is a good display of the fact that the measured heat transfer coefficient h (slope of the solid line in Figure 10) is independently defined from the power setting and thermal boundary conditions ΔT=T w -T f Axial Distribution of h on the Casing Plate Figure 3.9 presents heat transfer coefficient data at all five axial locations defined in Figure 3.1. At this stage, q conv is not corrected for lateral conduction losses yet. The same measurement methodology described in the previous sections is applied at all five locations. Figure 3.10 shows the axial distribution of h at five axial locations on the casing plate facing the blade tips. The solid circles represent the data before a lateral conduction correction in the aluminum casing plate is applied. The magnitude of lateral conduction losses are carefully determined from a 3D heat conduction analysis described in Figures 3.3 to 3.6. A proper heat loss analysis was performed for each power setting level carefully. A significant change in the overall magnitude of the heat transfer coefficients is observed after taking into account all energy losses from the Al casing plate, especially the lateral conduction losses. The casing plate measurement locations see the subsequent passage of tip leakage related fluid and passage fluid at blade passing frequency. The circumferential mixing in this near casing area is inevitable. A proper lateral conduction calculation is essential to reduce the experimental uncertainties in this heat transfer measurement approach.

71 Figure 3-9: Measured Heat Transfer Coefficient at Five Axial Locations on the Casing Surface 54

72 55 Figure 3-10: Distribution of the Heat Transfer Coefficient with respect to Axial Position on the Casing Surface

73 Chapter 4 Casing Treatment Definitions The effect of casing treatment on over tip leakage flow was investigated by using two types of casing surfaces. The first types of casing surfaces with surface treatments in the form of parallel patterns were manufactured using a stereolithography (SLA) technique. The second types of casing surfaces were obtained by applying commercial grade sandpaper to the inner surface of the casing with double-sided adhesive tape. 4.1 AFTRF Casing Modifications The casing window was machined out locally to allow casing surfaces covering 10% of the circumference to be tested with various treatments, shown in Figure 4.1. The machined pocket allows operating at about 0.8% tip clearance when there is no treatment. Machined pocket for casing surface mounting Figure 4-1: Casing Treatment on the Inner Surface of the AFTRF Casing

74 4.2 Casing Surface Treatments 57 Casing treatments in the form of smooth and parallel patterns with different angles with respect to tangential direction were manufactured using a SLA technique. Following figures show three different types of casing surface treatments used in the current investigation along with their definitions. Axial Direction Smooth Casing Treatment Figure 4-2: Smooth Casing Treatment Smooth casing treatment was used to obtain a reference surface. This was done to isolate the effect of pattern treatments from the effect of tip gap height differences due to machined pocket for SLA pattern mounting.

75 58 Axial Direction +45 degrees Tangential Direction + 45 Figure 4-3: Straight Casing Treatments Axial Direction +45 degrees Tangential Direction + 45 Figure 4-4: Curved Casing Treatments

76 59 Figure 4.5 shows the details of the solid models. The dimensions of each casing surface used in this investigation are shown in Table 4-1. All casing surfaces have the same total thickness. Figure 4.6 shows the rotor blade position with respect to the casing treatment. Table 4-1: Dimensions of Casing Surfaces Panel No. Shape Angle [Deg] Ridge Height Ridge Spacing B Smooth S1 Straight +45 Low Wide S2 Straight +45 High Wide S3 Straight +45 Low Narrow C1 Curved +45 Low Wide C2 Curved +45 High Wide C3 Curved +45 Low Narrow

77 Figure 4-5: Solid Model of Casing Treatment with Patterns 60

78 61 Axial Direction Figure 4-6: The Rotor Blade Position with respect to the Casing Treatment

79 A typical SLA insert flush mounted on the inner surface of the AFTRF casing is shown in Figure Figure 4-7: SLA Insert Flush Mounted on the Inner Side of the Removable AFTRF Window

80 4.3 Casing Surface Roughness 63 The influence of surface roughness character on the flow near a rough wall has been studied by many researchers in the past. First fundamental studies in this area were conducted by Nikuradse [53] and Schlichting [54]. Most of the past studies use an equivalent sandgrain roughness size k representing a measure of the size of sand grains used on the surface. Relative to smooth surfaces, they found significant augmentations of mixing and turbulent transport in the boundary layers developing over roughened surfaces. The main goal of the study is to investigate if the artificial roughening of the casing surface facing the rotor blade tips could be used to reduce the tip leakage mass flow rate in an effort to improve turbine stage performance. The original casing surface was artificially roughened by applying two grades of sandpaper to the inner surface of the casing with double-sided adhesive tape, shown in Figure 4.8. Two grades of sandpaper used were a coarse 60 grit paper (mean roughness height k = 166 μm) and a fine 150 grit paper (mean roughness height k = 110 μm). A smooth, thermoplastic layer was also separately applied in a few experiments to obtain a smooth reference surface. This was done to isolate the effect of artificially roughening the casing surface from the effect of tip gap height reduction due to sand paper thickness. The nominal thickness of the applied fine sand paper was mm or 0.33% h. The coarse sand paper was mm thick (0.66 %h).

81 Figure 4-8: Application of Sandpaper to the Inner Surface of the Casing Window 64

82 Chapter 5 Tip Clearance Aerodynamics Baseline Rotor As introduced previously, the pressure difference between the pressure side and the suction side across the tip clearance drives the flow to leak from the pressure side into the suction side of a neighboring passage. The tip leakage flow interacts with the main flow and forms a tip leakage vortex in the blade passage along the corner of suction side thereby affecting the passage flow. The tip leakage flow and the tip leakage vortex have significant effects on turbine in blocking the main flow, heat protection and aerodynamic loss. Almost 1/3 of the losses in a turbine rotor stage can be associated with the tip leakage flow [4]. The results presented in this chapter discuss the effects of blade tip shapes and tip gap height on tip leakage flow and the observed total pressure downstream of the rotor. The effects of blade tip shapes and tip gap height on leakage vortex characteristics are studied in the stationary frame by measuring total pressure at 30% chord downstream of the rotor exit plane. A high resolution total pressure map at stage exit in the cold research turbine could be used as a measurement of aerodynamic efficiency [39]. One of the objectives of this chapter is to set-up baseline data to compare to the effect of casing treatments.

83 5.1 Total Pressure Measurements 66 Total pressure measurements downstream of the rotor are obtained for all 29 blade passages using a phase-locked measurement technique. A Kulite probe with a frequency response of 150 khz was used in the stationary frame. The probe head was located at 30% chord downstream of the rotor exit plane (Figure 2.2). A total of 6000 samples, at an acquisition rate of 20 khz, is collected per revolution. Hence, there are 206 data points in each blade passage. The results are presented in the form of a total pressure coefficient, as defined by Equation 5-1, radial distributions of passage averaged and rotor averaged total pressure coefficients, as defined by Equation 5-2 and Equation 5-3 respectively, circumferential distribution of total pressure coefficient at individual radial locations, and area averaged total pressure coefficient for individual passages, as defined by Equation 5-4. C pt (i, j) P (i, j) P = (5-1) 1 2 ρu m 2 C C pt,p pt,r i C pt (i, j) i (j) = 206 (5-2) 6000 C pt (i, j) i= 1 (j) = 6000 (5-3) C pt,a = C pt (i, j) (5-4) passage 25% blade height

84 67 The passage averaged coefficient is obtained by circumferentially averaging the measured total pressure coefficient over each passage as defined in Equation 5-2, at each radial position. The passage averaged coefficient is computed for the passage that contains the leakage vortex of the blade number referenced. Thus, the passage bounded by the suction surface of B25 and pressure surface of B24 contains the leakage vortex of B25. The radial distribution of the passage averaged coefficient allows for comparison of tip gap behavior with and without casing treatments by isolating the effect of tip leakage in each passage. An area averaged coefficient, defined in Equation 5-4, is also calculated for individual passages to get a single value which captures the effect over the entire passage. The average is computed over an area of one passage (206 points) and a height of 25% span starting at the tip. 5.2 Baseline, No Casing Treatment As described previously, six blades, in two groups of three (B9, B10, B11) and (B24, B25, B26), have been modified for the current research. Baseline test was performed with test blades (B24, B25, and B26). The measurements were conducted using the high response total pressure probe described in Chapter 2. The probe is used in the stationary frame to record the total pressures downstream of the rotor. Thus a complete map of the rotor exit total pressure field is obtained for all the 29 passages per revolution of the rotor. Although a complete map of total pressure coefficient is obtained

85 68 for all the 29 rotor passages, the contour plots presented show only five passages containing the test blades. Figure 5.1 is a contour plot of the total pressure coefficient, as defined in Equation 5-1. These figure documents the original operational characteristics of the AFTRF without any pattern treatment. Boundaries at the hub, the casing, and 50%, 75%, and 85% blade span are marked in the contour plot. Each blade number and the corresponding tip gap height are noted above the casing boundary. The direction of rotation is from right to left, as represented by the blade speed vector below the hub boundary. The contour map shown in Figure 5.1 contains three flat-tip blades (B24, B25, and B26) with large tip clearance (t/h=2.5 %) and two subsequent blades (B27 and B28) with nominal tip gap (t/h=0.85 % and 0.81%, respectively). Figure 5-1: Total Pressure Coefficient Contours with Flat-Tip Blades (t/h = 2.5%)

86 69 The tip leakage vortex described by low total pressure is one of the flow features dominates the last 15% blade height. The total pressure in the core of the tip leakage vortex is the lowest total pressure measured in the whole passage. The leakage vortices of test blades occupy about 15% of the blade span and extend well into the blade passage. These vortices have a well defined core and strong gradients in both radial and circumferential directions. The tip leakage structures from neighboring blades show the effect of the tip clearance on the leakage vortex. The minimum total pressure of the test blades, measured in the core of the tip vortex, is also lower than that of the neighboring blades shown because of the much larger tip clearances of these blades. The size of the tip leakage vortex and the total pressure defect increase with increasing tip clearance. This behavior is also observed in Figures 5.3, 5.6 and 5.7. The second distinct flow feature in the region above 85% blade height is the zone bounded by the outer casing, the passage core and the two adjacent tip vortices. Typical underturning overturning behavior is observed in this zone. This feature is caused by the tip leakage vortex. The deficit in flow turning, due to the tip leakage flow, increases with increasing tip clearance. Therefore, the underturning is the greatest in the passages containing the test blades leakage vortices, and hence the measured total pressures are lower than that of neighboring passages. The main flow regime observed in the region 75%-85% blade height is the tipside passage vortex present just below the tip leakage vortex, as shown in Figure 5.1. It has a smaller total pressure defect than that of the leakage vortex. The results in McCarter et al [17] also confirm the identity of this structure. The tip-side passage vortex region of test blades is not distinct, i.e. the total pressure of the tip-side passage vortex does not

87 70 show up as a strong low total pressure zone due to the interaction with the tip leakage vortex. This behavior is also observed in cascade results presented by Yamamoto [55]. The tip-side passage vortex is weakened by its interaction with the tip leakage vortex, as these two have opposite sense of rotation [28]. The interaction of the passage vortex with the leakage vortex is greatest for the test blades (B24, B25, and B26), as seen by the poor definition of the passage vortex in the wake. It is clear that the passage vortex is dominated by the tip leakage vortex for the case of large tip clearance. The increased tip vortex due to the increased tip gap distorts the tip-side passage vortex. However, it is also clear that an independent passage vortex forms for the case of small tip clearance. The core flow in the passage is clearly visible as characterized by the dark blue high kinetic energy region. It has a well defined core of high total pressure at around 50% blade height. In the passages containing the test blades, the passage cores are shifted radially downwards, probably by the radially inward migration of fluid due to the blockage of the tip leakage vortex. Since variations below mid span were found to be unaffected by tip clearance, all of the data sets cover a span from 0.438h to 0.981h. 5.3 Effect of Blade Tip Shapes Tip leakage flow driven by the pressure difference between the pressure side and suction side of a rotor tip rolls into a streamwise vortical structure, mixes with the rotor passage flow and causes total pressure loss. Because of its nature, tip leakage flow is a source of aerodynamic inefficiency as well as of high thermal loading near the tip. The

88 71 tips of rotating gas turbine blades are subjected to extremely high heat transfer rates due to high temperatures associated with the leakage flow. In order to reduce the tip leakage flow and the tip heat transfer, a common technique is to use a squealer tip. The squealer tip essentially acts as a labyrinth seal, and increases the resistance to the flow, resulting in to lower leakage flow rates and lower heat transfer rates. A squealer tip also allows for a smaller tip clearance than the flat tip without incidental rubs between the blade tip and the shroud. The results presented in this section discuss the effects of two blade tip shapes on tip leakage flow. Figure 5.2 is a contour plot of the total pressure coefficient for all the 29 rotor passages, which contains three squealer-tip blades (B9, B10, and B11) and three flat-tip blades (B24, B25, and B26) with tip clearance of t/h=2.5 %. Figure 5-2: Total Pressure Coefficient Contours for All the 29 Rotor Passages (Baseline)

89 72 The contour map shown in Figure 5.3 contains three squealer-tip blades (B9, B10, and B11) with large tip clearance (t/h=2.5 %) and two subsequent blades (B27 and B28) with nominal tip gap (t/h=0.76 % and 0.75%, respectively). Boundaries are drawn as solid lines around the tip vortex zone to enable easy comparison with results presented later. The dashed lines are used to distinguish the limits of the core flow in the passage. These details are consistently reproduced in all subsequent contour plots. Figure 5-3: Total Pressure Coefficient Contours with Squealer-Tip Blades (t/h = 2.5%) Although the contour plot in Figure 5.3 shows that the squealer-tip blades have almost the same characteristics as the flat-tip blades, the radial distribution of the passage

90 73 averaged pressure coefficient in Figure 5.4 shows that there is an increase in total pressure. This confirms the results obtained by Heyes et al. [24]. They studied squealer tip geometry in a cascade configuration and compared it to flat tip geometry. They showed that the squealer tip geometry has the ability of reducing tip clearance loss. 1 Fraction of Blade Height from Hub B10 - Squealer Tip B25 - Flat Tip Passage Averaged Total Pressure Coefficient Figure 5-4: Effect of Blade Tip Shape on the Passage Averaged Coefficient Figure 5.5 shows the area averaged total pressure coefficient in the passage containing the tip leakage vortices of all 29 blades. The effect of blade tip shapes over the entire passage can be seen clearly. Test blades (B9, B10 and B11) and (B24, B25 and B26) represent the group of three squealer-tip blades and flat-tip blades, respectively. The extent of the area average covers one passage and 25% blade height.

91 74-4 Area Averaged Total Pressure Coefficient Tip Vortex Number Figure 5-5: Effect of Blade Tip Shape on the Area Averaged Total Pressure Coefficient 5.4 Effect of the Tip Gap Height The results presented in this section are to establish the effect of the tip gap height on the tip leakage flow, without any casing treatment. The three cases studied are shown n Figure 5.3, 5.6 and 5.7. These figures show the contours of total pressure coefficient at 30% downstream of the rotor. Figure 5.3, 5.6 and 5.7 show contour plots of total pressure coefficient obtained with the test blade tip clearance of t/h=2.5%, t/h=1.5% and t/h=0.75%, respectively. The clearance was reduced by using different sets of SLA blades. The solid lines marked around the tip vortex zones are from the baseline case,

92 75 shown in Figure 5.3, and used for comparison purposes. The tip vortex is the site of the minimum total pressure recorded at the turbine exit. The figures show that the tip vortex controlled zone from the test blades tends to occupy a smaller space as the gap height decreasing, i.e. the effect of reducing the tip clearance is evident in the distribution and level of measured total pressure at stage exit. The tip leakage vortices of test blades (B9, B10, and B11) are much smaller in size and closer to the suction-side of the passage. The vortex cores documented for the baseline case are clearly eliminated for both t/h=1.5% and 0.75%. The tip leakage vortices in the neighboring passages are unaffected. Figure 5-6: Total Pressure Coefficient Contours with Squealer-Tip Blades (t/h = 1.5%)

93 76 Figure 5-7: Total Pressure Coefficient Contours with Squealer-Tip Blades (t/h = 0.75%) The tip-side passage vortex tends to merge with the tip vortex flow structure when the tip vortex is a dominant flow feature as shown in the baseline case (Figure 5.3). The tip-side passage vortices of test blades (B9, B10, and B11) are now better defined due to reduced interaction with the leakage vortices for both t/h=1.5% and 0.75%. The core flow in the passage is clearly visible as characterized by the dark blue high kinetic energy region starting to extend from 0.75 h. The passage cores have rotated in a counterclockwise direction due to the reduced blockage of the leakage flow.

94 77 The wake plot at r/h = 0.96, presented in Figure 5.8 clearly shows the movement of the wakes of the test blades towards the suction surface of the blade. The red squares indicate a significant total pressure defect in the tip leakage vortex when the clearance is large, at t/h=2.5%. The total pressure defect is reduced (blue triangles) Total Pressure Coefficient, Cpt t/h = 1.5% Baseline: t/h = 2.5% Blade Number Figure 5-8: Effect of Reducing the Tip Clearance of Test Blades on the Wake Profile at r = 0.96h The radial distribution of the passage averaged pressure coefficient in Figure 5.9 shows that there is a definite shift in the flow field towards the blade tip, in addition to the increase in total pressure. This confirms that the lower total pressure seen in the wake plots is due to a change in the passage flow structure. The defect due to the tip-side passage vortex is also better defined than in the baseline sets because the tip vortex/passage vortex interaction is reduced when the tip gap is reduced to t/h=0.75%.

95 78 1 Fraction of Blade Height from Hub Baseline: t/h=2.5% 0.9 t/h=1.5% t/h=0.75% Passage Averaged Total Pressure Coefficient Figure 5-9: Effect of the Tip Clearance on the Passage Averaged Coefficient of Blade 10

96 Chapter 6 Aerodynamic Influence of Casing Treatments The aerodynamic influence of casing treatments on over-tip-leakage flow was investigated in a large scale, rotating, axial turbine rig. Phase-locked measurements of the absolute total pressure in a cold flow turbine research facility were conducted at the turbine stage exit using a high-frequency response total pressure probe. Time accurate measurements provided valuable aerodynamic information quantifying the near tip flow modifications imposed by casing treatments. The influence of casing treatments was studied in a range of tip clearance values. The current study aims to improve our physical understanding of axial turbine tip leakage flow under the influence of casing treatments. Although there have been many studies in understanding the influence of casing treatments on aerodynamic efficiency or losses in turbomachinery systems, our knowledge on the influence of casing treatments in the tip region of an axial turbine is very limited. The main goal of the study is to investigate if the casing treatments facing the rotor blade tips could be used to reduce the tip leakage mass flow rate in an effort to improve turbine stage performance.

97 6.1 Baseline at Tip Clearance of t/h = 2.5% 80 The total pressure coefficient contour map with a smooth casing surface is shown in Figure 6.1. The contour map contains three squealer-tip blades (B24, B25, and B26) with large tip clearance (t/h=2.5%) and two subsequent blades with nominal tip clearance (t/h=0.85 % and 0.81%). Figure 6-1: C pt Distribution at 30 % Downstream of AFTRF Rotor (Squealer Tips)

98 6.2 Casing Treatments at Tip Clearance of t/h = 2.5% 81 As in the previous chapter the results are presented in form of contour plots, radial distributions of passage averaged coefficient, wake plots, and area averaged total pressure coefficient. The comparisons made in this section are with respect to the baseline data shown in Figure Straight Casing Treatment Panel No.S1 The total pressure coefficient distribution shown in Figure 6.2 was obtained after applying straight pattern (S1) to the casing inner surface. It is clear that the pattern treatment has considerable effect on both the tip leakage flow and the tip-side passage vortex. The tip leakage vortices of test blades B24, B25 and B26 shown are greatly reduced in size. Their individual total pressure defect is also substantially reduced. Additionally, the severe gradient observed in the tip leakage vortex due to the large tip clearance of t/h = 2.5% is also eliminated. The tip leakage vortices of test blades moved significantly towards the blade suction surface. It was shown in Chapter 5 that the movement of the tip leakage vortices towards the blade suction surface is observed when the tip clearance was reduced from t/h = 2.5% to t/h = 0.75%. This behavior of tip leakage vortices observed here is obtained without the change in the tip gap height. The tip leakage vortices of blades B27 and B28 are significantly weakened and completely mixed in with the wake fluid. It is of course unclear whether the leakage flow is eliminated or if it is completely mixed in with the wake.

99 82 Figure 6-2: C pt Distribution at 30 % Downstream of AFTRF Rotor (S1) None of the red C pt areas with strong total pressure defect are visible after the straight casing treatment with smaller pattern height. The application of casing treatment is expected to locally increase turbulent kinetic energy within the tip gap, leading to greater total pressure loss. Thus the secondary kinetic energy associated with the tip leakage vortex in the blade passage is reduced and the over tip leakage fluid is relatively weakened.

100 83 The tip-side passage vortices appear to be stronger at large tip gap height. This is partly due to the reduced interaction between the tip leakage flow and the tip-side passage secondary flow Straight Casing Treatment Panel No.S2 Figure 6.3 is a contour plot of the total pressure field downstream of the rotor with straight pattern (S2) applied on to the casing inner surface. It is clear that the straight pattern treatment with higher pattern height has much more effect on the tip leakage flow. The tip leakage vortices are greatly reduced in area and the total pressure field within the vortices has much less defect when compared to smooth wall and casing treatment of S1. The weakened tip leakage vortices at the small gap heights are again mixed in with the blade wakes. The passage flow cores with high momentum are shifted considerably to the left and this is probably due to the reduced blockage from the tip vortex and the new position of the tip-side passage. Figure 6.2 and 6.3 shows a strong beneficial influence of using a straight casing treatment. The weakened tip vortex system and a reduced shear interaction of the tip vortex and the passage vortex system are clearly observed.

101 84 Figure 6-3: C pt Distribution at 30 % Downstream of AFTRF Rotor (S2) Straight Casing Treatment Panel No.S3 Figure 6.4 is a contour plot of the total pressure field downstream of the rotor with straight pattern (S3) applied on to the casing inner surface. The thickness of the applied treatment is same as that of straight pattern S1. The spacing between the patterns is tighter than that of other straight casing treatments. The effect of casing treatment S3 on the tip leakage flow appears to be similar to that of straight pattern S1. The tip leakage vortices of test blades B24, B25 and B26 shown are greatly reduced in size. Their individual total pressure defect is also substantially reduced. The tip leakage vortices of test blades moved again towards the blade suction surface but the movement is less than

102 that of straight pattern S1. The weakened tip leakage vortices at the small gap heights are again mixed in with the blade wakes. 85 Figure 6-4: C pt Distribution at 30 % Downstream of AFTRF Rotor (S3) Curved Casing Treatment Panel No.C1 The total pressure coefficient distribution in the measurement plane with a curved pattern (C1) applied to the casing inner surface is shown in Figure 6.5. The tip gap height of each blade is noted above the casing boundary. Comparing the pitch-wise extent of the vortices to the boundary marking the vortices with smooth casing, it is apparent that curved casing treatment has reduced the

103 86 size of the vortex. The reduction in total pressure defect is less compared to that of straight casing treatment with same pattern height (S1); however the expected movement of the tip leakage vortices to the right is also observed. The tip-side passage vortex region is not distinct, due to the interaction with the tip leakage vortex. The relative locations of passage flow cores with respect to tip leakage flow are about the same as that observed in the baseline (smooth casing). Figure 6-5: C pt Distribution at 30 % Downstream of AFTRF Rotor (C1)

104 Curved Casing Treatment Panel No.C2 The total pressure coefficient distribution shown in Figure 6.6 was obtained after applying curved pattern (C2) to the casing inner surface. The thickness of the applied treatment is same as straight pattern (S2). It is clear that the pattern treatment has considerable effect on both the tip leakage flow and the tip-side passage vortex. The tip leakage vortices of test blades B24, B25 and B26 shown are greatly reduced in size. Their individual total pressure defect is also substantially reduced. The reduction in total pressure defects are almost same compared to that of straight casing treatment with same pattern height (S2) and the gradients across the tip leakage vortices are also smoother. The weakened tip leakage vortices at the small gap heights are again mixed in with the blade wakes. The tip leakage vortices of test blades moved significantly towards the blade suction surface. The movement of the tip leakage vortices towards the blade suction surface and the reduction in the size of the leakage vortices can occur if the gap normal momentum is reduced, similar to when the gap height is reduced. This means that the application of casing patterns does cause reduction in gap normal velocity by reducing gap mass flow rate. The passage flow cores with high momentum have moved back to a location comparably similar to that observed in the baseline and this is probably due to the blockage from the tip vortex and the interaction between the tip-side passage vortex and the tip leakage vortex.

105 88 Figure 6-6: C pt Distribution at 30 % Downstream of AFTRF Rotor (C2) Curved Casing Treatment Panel No.C3 The total pressure coefficient distribution in the measurement plane with a curved pattern (C3) applied to the casing inner surface is shown in Figure 6.7. The thickness of the applied treatment is same as that of curved pattern C1. The spacing between the patterns is tighter than that of other curved casing treatments. It is clear that the curved pattern treatment with tighter pattern spacing has much more effect on the tip leakage flow than that of straight pattern treatment with the same pattern spacing (S3). The tip leakage vortices of test blades B24, B25 and B26 shown are greatly reduced in size. The

106 tip-side passage vortex region is not distinct, due to the interaction with the tip leakage vortex. 89 Figure 6-7: C pt Distribution at 30 % Downstream of AFTRF Rotor (C3) Passage to Passage C pt Variation near the Casing Since the C pt measurements are phase-locked, a comparison of the total pressure (wake) profile for different casing roughness treatments in different passages is illuminating. The wake profiles in Figure 6.8 for r=0.96h show that the casing treatments leads to considerable reduction in wake defect. Those figures also clearly show the movement of the wakes of the test blades towards the suction surface of the blade. The reduction in wake depth of the test blades B24, B25 and B26 is greater than the reduction

107 obtained by the effect of reducing the gap height only (Figure 5.6). The total pressure defect reduction is also greater at the larger tip gap heights. 90 Total Pressure Coefficient, Cpt Straight Pattern - S1, t/h=2.5% Straight Pattern - S2, t/h=2.5% Straight Pattern - S3, t/h=2.5% Baseline, t/h=2.5% Blade Number Total Pressure Coefficient, Cpt Curved Pattern - C1, t/h=2.5% Curved Pattern - C2, t/h=2.5% Curved Pattern - C3, t/h=2.5% Baseline, t/h=2.5% Blade Number Figure 6-8: C pt Distribution at a Fixed Radius r = 0.96h with Significant Casing Treatment Influence

108 Comparison of the Averaged Total Pressure Coefficient The radial distribution of the passage averaged pressure coefficient in Figure 6.9 shows that there is a definite shift in the flow field towards the blade tip, in addition to the increase in total pressure. This confirms that the lower total pressure seen in the wake plots is due to a change in the passage flow structure. 1.0 Fraction of Blade Height from Hub Straight Pattern - S1, t/h=2.5% Straight Pattern - S2, t/h=2.5% Straight Pattern - S3, t/h=2.5% Baseline, t/h=2.5% 1.0 Fraction of Blade Height from Hub Curved Pattern - C1, t/h=2.5% 0.6 Curved Pattern - C2, t/h=2.5% Curved Pattern - C3, t/h=2.5% 0.5 Baseline, t/h=2.5% Passage Averaged Total Pressure Coefficient Figure 6-9: Effect of Casing Treatments on the Passage Averaged Coefficient of Blade 25

109 6.3 Effect of Casing Surface Roughness 92 The original casing surface was artificially roughened by applying two grades of sandpaper to the inner surface of the casing with double-sided adhesive tape. Two grades of sandpaper used were a coarse 60 grit paper and a fine 150 grit paper. A smooth, thermoplastic layer was also separately applied in a few experiments to obtain a smooth reference surface. This was done to isolate the effect of artificially roughening the casing surface from the effect of tip gap height reduction due to sand paper thickness. Rao [39] showed that an artificially roughened casing in relative motion clearly reduces momentum deficit that is in the core of each tip vortex. The measurements also show that the significant shearing between the tip vortex and the passage vortex is weakened in the last 15 % of the blade height. The radial distribution of the rotor averaged pressure coefficient in Figure 6.10 shows that the artificial introduction of surface roughness appears to have no observable effect in the region of the tip leakage vortex at the tip clearance of t/h=0.9%. On the other hand, there is a definite shift in the flow field towards the blade tip, in addition to the increase in total pressure at the tip clearance of t/h=2.5%. The effect of coarse roughness on the tip leakage flow appears to be much better than that of fine sandpaper at larger tip clearances.

110 93 1 Fraction of Blade Height from Hub Smooth Casing Surface, t/h=0.9% Coarse Roughness (60 Grit), t/h=0.9% Fine Roughness (150 Grit), t/h=0.9% Fraction of Blade Height from Hub Smooth Casing Surface, t/h=2.5% Coarse Roughness (60 Grit), t/h=2.5% Fine Roughness (150 Grit), t/h=2.5% Passage Averaged Total Pressure Coefficient Figure 6-10: Rotor Averaged Coefficient with Different Casing Roughness Treatments

111 Chapter 7 Inclined Squealer Tip The tip leakage flow that exists between the stationary casing and the rotor tip of a shroudless high-pressure turbine remains a major source of aerodynamic loses. Almost 1/3 of the losses in a turbine rotor stage can be associated with the tip leakage flow. Further, the tip leakage flow contributes to migration of hot gases over the tip resulting in increased thermal distress in the tip region. Consequently, research on tip desensitization techniques is the main topic on turbine tip clearance in recent years. Most of the tip desensitization techniques reported involve modification of the tip surface geometry in an effort to decrease the leakage mass flow rate, thereby reducing pressure losses and thermal loads on the blade tip. One of the most common methods to reduce the flow on the tip is to use a recessed tip, also known as a squealer tip. A squealer tip concept has been discussed in the previous sections. This chapter focuses on a new blade tip concept, which is a variant of a conventional squealer tip and discuss the effects of the new blade tip concept on tip leakage flow with and without casing treatments. The blade tip investigated is a GE patented concept and shown in Figure 7.1. It is referred to as an inclined shelf because the squealer wall over the shelf is inclined towards the pressure side [56].

112 95 Figure 7-1: Inclined Squealer Tip Manufactured by SLA Technique 7.1 Squealer Tip vs. Inclined Squealer Tip As in the previous chapters the results are presented in form of contour plots, radial distributions of passage averaged coefficient, wake plots, and area averaged total pressure coefficient Smooth Casing Treatment The total pressure coefficient distribution in the measurement plane with a smooth casing surface is shown in Figure 7.2. Boundaries at the hub, the mid-span and the

113 casing, are marked in the contour plot. Each blade with inclined squealer tip and the corresponding tip gap height are noted. 96 Smooth Casing B (0.75%) (1.5%) Figure 7-2: Total Pressure Coefficient Contours (Smooth Casing) The contour map shown in Figure 7.3 contains three inclined-squealer-tip blades (B24, B25, and B26) with tip clearance of t/h=1.5% and a subsequent blade with nominal tip gap (t/h=0.85%). The comparisons made in this section are with respect to the baseline data shown in Figure 5.6.

114 97 There is reasonable reduction in the tip leakage vortices of the test blades (B24, B25, and B26), as measured by the total pressure downstream of the rotor exit. Inclined squealer tip arrangement has severe effects on both passage core flow and the interaction between the leakage vortex and the tip side passage vortex. It also affects the tip leakage vortex in the neighboring passage of B27. Although tip gap height of B27 (t/h=0.85%) is greater than that of B12 (t/h=0.76%), the tip vortex of B27 has a lower value of total pressure compare to that of B12 shown in Figure 5.6. Inclined Squealer Tip U Hub 0.5 h Passage Core Tip leakage vortex Casing # 24 (1.5%) # 25 (1.5%) # 26 (1.5%) # 27 (0.85%) Tip-side passage vortex U 0.75 h 0.5 h Hub 0.85 h Smooth Casing Figure 7-3: Total Pressure Coefficient Contours with Inclined Squealer Tips (Smooth Casing)

115 98 CFD analysis of inclined squealer tip performed by Parakash et al [56] shows that as the leakage fluid flows over the pressure surface, the incline forces it to turn by a large angle resulting in separation that serves to reduce leakage. Figure 7.4 shows the flow over inclined shelf. Figure 7-4: Flow over Inclined Shelf [56]

116 Curved Casing Treatment The total pressure coefficient distribution with a curved casing treatment is shown in Figure 7.5. Boundaries at the hub, the mid-span and the casing, are marked in the contour plot. Each blade with inclined squealer tip and the corresponding tip gap height are noted. Figure 7-5: Total Pressure Coefficient Contours (Curved Casing Treatment)