The first step in the investigation is to characterize the flow when the second blade of the tandem blades is positioned in the "no gap nozzle effect", Figure 13. The interaction between the two blades is only evident in terms of the wake of the front blade that affects the flow behavior along the arrangement. To characterize the influence of the second blade position, different values are given for the axial and tangential displacements of the second blade. Figure 14 shows a graphical representation of the front blade's trailing edge with the possible second blade locations in terms of the axial and tangential displacements. There are a range of three axial displacements that varies between 0 and 0.2 and a range of five tangential displacements that varies between 0.15 and -0.15. These two ranges represent relative position of the rear
The effect of rotor tip clearances in turbomachinery applications has been a primary research interest for nearly 80 years. Over that time, studies have shown increased tipclearance in axialflowcompressors typically has a detrimental effect on overall pressure rise capability, isentropic efficiency, and stall margin. With modern engine designs trending toward decreased core sizes to increase propulsive efficiency (by increasing bypass ratio) or additional compression stages to increase thermal efficiency by increasing the overall pressure ratio, blade heights in the rear stages of the high pressure compressor are expected to decrease. These rear stages typically feature smaller blade aspect ratios, for which endwall flows are more important, and the rotor tipclearance height represents a larger fraction of blade span. As a result, data sets collected with large relative rotor tipclearance heights are necessary to facilitate these future small core design goals.
Pullan et al.  show that the cause of the blue holes in compressors with small clearance is vorticity shed from a flow separation at the leading edge of the rotor tip.
Their work also found that the blue holes propagate and grow into a spike stall cell.
The CFD work suggests that, in a compressor with small tip-clearance, the vortex is swept across the blade passage and triggers a new separation at the next blade leading to the formation of a new blue hole. However, instead of convecting down- stream and out of the passage with the bulk of the flow, as is the case in a large clearance machine, the flow sweeps the original blue hole on a near-tangential trajec- tory around the leading edge of the next blade. This means that the two blue holes combine and move tangentially towards the next blade. This process is illustrated in Diagram 8.2. As fluid is being added to the blue hole, there is the potential for it to grow rapidly into a stall cell.
The following discussion will briefly review past stall inception investigation methodology in light of the new observations made in the present study. As was discussed in Section 2.3.2 of the Literature Review, most previous experi- mental work has relied on measurements taken upstream (or downstream) of the blade row to locate the point of stall inception. Apart from a few overtip pres- sure measurements, there are almost no measurements in the path of the rotor blades. The current work shows that the inception process grows from a small disturbance inside the blade passage and that sparse instrumentation upstream of the blade row is unlikely to detect the very first signs of flow breakdown. In most cases, upstream instrumentation would only detect the stall cell after it has propagated some distance and grown in size. CFD studies also seem to suffer from the same problem in that the stall event is often analyzed where the spike is first clearly observed rather than where it is embryonic and difficult to detect. Experimental and computational investigations supporting the forward spillage hypothesis (e.g., Deppe et al. (2005) and Chen et al. (2008)) provide the spatial or temporal resolution equivalent to observing the spike formation at one circum- ferential location or on a once per rotor revolution basis. This resolution is too low to observe the progression of spike formation described in the current work. Instead, the spike examined in these and other past studies was likely already an established stall cell.
for the unsteady simulation was closer to the experiment.
3.4 Wall modeling approach
The choice of how to resolve the boundary layer was shown in Paper I to play an important part in the performance prediction of an axial compressor stage. Two approaches were analyzed: the k- turbulence model using wall functions and the k- turbulence model using Chien’s low-Reynolds formulation, where the first grid node away from the wall is placed in the viscous sub-layer to accurately resolve the boundary layer. Using wall functions was shown to predict a thicker boundary layer compared to the low-Reynolds formulation. A thicker boundary layer decreases the turning over the rotor (increase the deviation angle), which reduces the work done by the blade on the fluid. Performance parameters such as total pressure ratio and polytropic efficiency predicted using the low-Reynolds model show a closer agreement to the experimental measurements in Paper I compared to wall functions.
A numerical study of the effect of discrete micro tip injection on unsteady tipclearanceflow pattern in an isolated axial compressor rotor is presented, intending to better understand the flow mechanism behind stall control measures that act on tipclearanceflow. Under the influence of injection the unsteadiness of self-induced tipclearanceflow could be weakened. Also the radial migration of tipclearance vortex is confined to a smaller radial extent near the rotor tip and the trajectory of tipclearanceflow is pushed more downstream. So the injection is beneficial to improve compressor stability and increase static pressure rise near rotor tip region. The results of injection with different injected mass flow rates show that for the special type of injector adopted in the paper the effect of injection on tipclearanceflow may be different according to the relative strength between these two streams of flow. For a fixed injected mass flow rate, reducing the injector area to increase injection velocity can improve the effect of injection on tipclearanceflow and thus the compressor stability. A comparison of calculations between single blade passage and multiple blade passages validates the utility of single passage computations to investigate the tipclearanceflow for the case without injection and its interaction with injected flow for the case with tip injection.
Before the investigation of blade damage effects near stall boundary, another case with the medium damaged blade was investigated which was performed away from stall boundary at 60% speed. Compared with the solution obtained in the case operating at the same condition without any damaged blade, the overall performance for mass flow, temperature ratio and isentropic efficiency are the same. Pressure ratio differs the most however the difference is less than 1%. It could be concluded that the global performance was not significantly affected by the damaged blade. Flow blockage was only found in two passages including the damaged blade and flow seems to be compensated in other passages after its redistribution. Figure 5.3 indicates the circumferentially averaged axial velocity profile comparison away from stall boundary. It can be seen that the averaged axial velocity near the tip region from the casing to approximately 90% span was slightly lower for the case with one damaged blade which is as expected. The reduction of flow in that region was caused by the flow blockage due to the damaged blade. The mass flow for the case with one damaged blade seemed to be compensated in the region between the mid span to 90%. Since there was only one damaged blade in the rotor assembly, it result in no significant change in mass flow.
3 Professor, Dept of Mechanical Engg/ Vagdevi College of Engineering, Warangal, T.S India
4 Assistant Professor, Dept of Mechanical Engg/ Vagdevi College of Engineering, Warangal, T.S India
Abstract: In this Paper analytical investigations are done to determine the effect of unsteady tipclearanceflow in axial compressor rotor used in an industrial gas turbine. Thermal analysis is performed by changing the tipclearance 0.333 mm, 0.999 mm and 1.665 mm for different materials Monel K – 500 and Titanium alloy in air as fluid medium. 3D model of axial compressor rotor is drawn in Pro/Engineer and thermal analysis is done in Ansys to obtain total temperature and heat flux for Monel – k 500 and titanium alloy materials using air as a fluid. And Comparative analysis also has done by results of both materials to evaluation of heat flux by varying tipclearanceflow in axial compressor rotor using thermal analysis.
Effects of a large rotor tip gap on the performance of a one and half stage axial compressor are investigated in detail with a numerical simulation based on LES and available PIV data. The current paper studies the main flow physics, including why and how the loss generation is increased with the large rotor tip gap. The present study reveals that when the tip gap becomes large, tipclearance fluid goes over the tipclearance core vortex and enters into the next blade’s tip gap, which is called double-leakage tipclearanceflow. As the tipclearanceflow enters into the adjacent blade’s tip gap, a vortex rope with a lower pressure core is generated. This vortex rope breaks up the tipclearance core vortex of the adjacent blade, resulting in a large additional mixing. This double-leakage tipclearanceflow occurs at all operating conditions, from design flow to near stall condition, with the large tip gap for the current compressor stage. The double-leakage tipclearanceflow, its interaction with the tipclearance core vortex of the adjacent blade, and the resulting large mixing loss are the main flow mechanism of the large rotor tip gap in the compressor. When the tipclearance is smaller, flow near the end wall follows more closely with the main passage flow and this double-leakage tipclearanceflow does not happen near the design flow condition for the current compressor stage. When the compressor with a large tip gap operates at near stall operation, a strong vortex rope is generated near the leading edge due to the double-leakage flow. Part of this vortex separates from the path of the tipclearance core vortex and travels from the suction side of the blade toward the pressure side of the blade . This vortex is generated periodically at near
The effect of the rotor tip gap size on the compressor flow field has been widely investigated. Several previous studies (Mailach et al. , Maertz et al. , Kiel et al. , Inoue et al. , etc.) reported that rotating instability occurs when the tip gap is increased in axialcompressors. As the resulting non-synchronous blade vibration affects engine safety, many studies have been reported. Maertz et al.  reported that the movement of the instability vortex along the leading edge plane is the main cause of instability during the compressor operation at near stall operation. Inoue et al.  and Yamada et al.  reported formation of a tornado-type vortex near the suction surface and subsequent vortex breakup as the main mechanism of the flow at near stall operation with a large tip gap. Although significant progress has been made in the experimental investigation of tip gap flow lately, direct measurement of the unsteady flow field inside the rotor tip gap has not been possible. A new program to investigate the unsteady tipclearanceflow in a low speed one and a half stage axial compressor was initiated under the auspices of the NASA Fixed Wing Project to understand and mitigate losses associated with large rotor tip gaps of N+3 relevant small, high overall pressure ratio compressor aft stages. Detailed measurements of the unsteady tipclearanceflows at two tip clearances (0.5 mm and 2.4 mm, 0.49% and 2.34% rotor tip chord) have been performed with three-dimensional PIV in a refractive index-matched test facility at the Johns Hopkins University. Details of the PIV
and δ c is the blade tipclearance. The staggered spacing, g, is a function of the blade pitch, s, and stagger angle, γ, where average values are used,
(8-2) Hunter and Cumpsty (1982) report similar results obtained with an isolated rotor. Koch and Smith use the empirical curves shown in Fig. 8-1 to estimate the sum of the displacement thicknesses or blockage. This requires that ψ max values be supplied by some unspecified stall criterion. The displacement thicknesses plus the tangential force defect are then used to estimate the efficiency reduction due to end-wall losses. There is no doubt that Smith made a substantial contri- bution to our knowledge of end-wall boundary layers. But there is little reason to believe that the empirical models outlined can be used for general application. To recognize that the empirical curves shown are far from correlations of experi- mental results, one need only note that the tipclearance-to-staggered spacing ratios for all experimental data in Fig. 8-1 lie between 0.028 and 0.062. Careful study of the original reference shows that the experimental data trends contra- dict the empirical curves about as often as they are in agreement. And the exces- sive data scatter in Fig. 8-2 provides no real basis for any empirical correlation of the tangential force defect. Although far from a complete end-wall flow model, these references provide important insight into the end-wall boundary layer problem. It is clear that any end-wall boundary layer theory must address the
Available online: https://pen2print.org/index.php/ijr/ P a g e | 537 of tipclearance is smaller once the diffuser
dimension is reduced, or the pinch is accumulated. This means that to the effect of tipclearance pinched diffusers are less sensitive. Antecedently Backmanet al.(2007) over that vaneless pinched diffusers compressors are less sensitive to changes in tipclearance Because pinched diffuser is a smaller amount sensitive to tipclearance, it's doubtless that the pinch encounters some losses created by tipclearance. The losses some part are depicted as a rise within the secondary flow region, that the pinch suppresses, resulting in associate improved mechanical device potency.
Numerical study is performed to investigate the swirling flow around a rotating disk in a cylin- drical casing. The disk is supported by a thin driving shaft and it is settled at the center of the cas- ing. The flow develops in the radial clearance between the disk tip and the side wall of the casing as well as in the axialclearance between the disk surfaces and the stationary circular end walls of the casing. Keeping the geometry of the casing and the size of the radial clearance constant, we compared the flows developing in the fields with small, medium and large axial clearances at the Reynolds number from 6000 to 30,000. When the rotation rate of the disk is small, steady Taylor vortices appear in the radial clearance. As the flow is accelerated, several tens of small vortices emerge around the disk tip. The axial position of these small vortices is near the end wall or the axial midplane of the casing. When the small vortices appear on one side of the end walls, the flow is not permanent but transitory, and a polygonal flow with larger several vortices appears. With further increase of the rotation rate, spiral structures emerge. The Reynolds number for the onset of the spiral structures is much smaller than that for the onset of the spiral rolls in rotor-stator disk flows with no radial clearance. The spiral structures in the present study are formed by the disturbances that are driven by a centrifugal instability in the radial clearance and they are pene- trated radially inward along the circular end walls of the casing.
Wahiba Yaici , Mohamed Ghorab, Evgueniy Entchev et al  In the present paper a numerical study was performed to predict the influences of the inlet air flow distribution on the performance of heat exchangers. The ranges and values of the geometry of the heat exchanger are studied by the CFD simulation. The comparisons between experimental and the software data implies that the model used in the present study is reliable and can predict the thermal performances satisfactorily for heat exchanger. This study has significant contribution on the optimum design of header and distributor configuration of heat exchanger to minimize misdistribution.
performance parameters are no longer sensitive to grid nodes when the number of nodes is more than 2.9 million. Therefore, 2.9 million grid nodes are applied for the calculation in the paper under comprehensive consideration of computational accuracy and time.
No-slip and adiabatic conditions are applied to all wall surfaces; the total temperature and total pressure of standard atmosphere as well as axial inflow are specified as the inlet boundary condition; while the static pressure is specified at the outlet assuming radial equilibrium. The stage method is used to process the interface. Due to the extreme complexity of fan flow field at near-stall point, so far there is no accurate numerical simulation method to capture the fan stall point. Therefore, at current stage, numerical divergence point is applied as fan stall point by most researchers.
abstracted geometric parameters could be directly used for efficient optimisation of profile thermodynamic performance.
Of the leakage paths shown in Figure 2-5, the radial gap and interlobe gap are directly related to the design clearances. The discharge end face gap - also called the axial gap, is an assembly feature while the blow-hole area is an inherent feature of the rotor profile geometry. Not all leakages paths are present for the entire duration of a compression cycle and the size and shape of the leakage path will vary. Defining each of the leakage areas as a function of the cycle angle sufficiently captures the most important leakage characteristics in a very efficient manner. This efficiency comes at the cost of losing details such as how the area evolves along the path of the leakage flow. When describing leakage areas with non-dimensional area curves the simple objective is to calculate the minimum instantaneous area along each flow path as accurately as possible.
The performance of a conventional Savonius rotor (that has a half circle shaped of 180° blade arc angle, without shaft and with zero clearance) was experimentally tested under the same experimental conditions (at a constant wind speed 5.8 m/sec and Reynolds number of 2.56 x 105. In order to make a comparison between the performance of the conventional Savonius rotor and the new specially designed with clearance in radial and axial axis. Figs. (16) and (17) illustrate the power coefficient and the torque coefficient for the conventional Savonius, The Cp.max is found as 0.066 at Tip speed ratio λ= 0.46 and the CT.max is 0.24 at Tip speed ratio λ= 0.12.
CHAPTER 12. CONCLUSIONS
agreement with measured results. As with the third quadrant, however, the choice of initial conditions had a significant effect on the solution towards which the model converged.
Finally, if the compressor rotates in the opposite direction to that for which it was de- signed, but flow is forced through it in the design direction, then the compressor operates in the second of the two modes possible in the fourth quadrant. Under this mode of operation, the pressure difference across the compressor has the opposite sign to design conditions, as the outlet pressure is lower than the inlet. The performance characteristic curves for this mode of operation are not continuous with third quadrant operation, just as a dis- continuity exists between the characteristic for second quadrant with positive rotation and first quadrant operation. This mode of operation has considerable similarities with second quadrant operation with positive rotation. Like that mode, flow within the compressor is similar to that encountered in a mixer: flow in the test compressor was dominated by the circumferential velocity component, and and areas of recirculating flow occur downstream of each stator blade row, giving rise to large radial velocity components, which are evident even after circumferential averaging: tipward flow occurs in stators, while hubward flow occurs in rotors. Virtually all throughflow occurs in the hub and tip regions downstream of rotors, and in the tip region exclusively downstream of stators. At midspan, very large sep- aration bubbles containing recirculating flow patterns are attached to the pressure surfaces of rotor and stator blades, almost entirely blocking flow through the blade passages. The small amount of throughflow occurring within the blade passages at this span takes the form of a thin jet attached to the suction surface. This jet impinges on the suction surface of downstream blades, influencing the formation for the jet of the downstream blade row.
Irfan Syarif Arief 1 , Tony Bambang Musriyadi 2 ,Ahmad Dwi Arta Je Mafera 3 ,
Abstract one type of ship propeller is a ducted propeller. Ducted propellers are propellers with sheath or duct that can increase thrust on the propeller and useful for directing the flow of water that will pass through the Propeller. In addition to improving thrust, ducted propellers can also increase torque compared with no duct. The basic theory of momentum for this ducted propeller operation has been used by Horn (1940). In order for the efficiency of the thrust to be of good value, the volume of water passing through the propeller should be as large as possible, with the smallest possible flow velocity.