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Figure from:Fan Airflow

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par AT? ENULUL GAIU IJLALUL PVULIIAIIILOS Consider the turbomachine stages shown schematically in Figure 2.2.1 and Figure 2.2.2. In each case, the stage consists of two flow domains: the rotor domain, which is rotating at a prescribed angular velocity, followed by the stator domain, which is stationary. The order of the rotor and stator is arbitrary (that is, a situation where the rotor is downstream of the stator is equally valid). In a numerical simulation, each domain will be represented by a separate mesh. The flow information between these domains will be coupled at the mixing plane interface (as shown in Figure 2.2.2 and Figure 2.2.1) using the mixing plane model. Note that you may couple any number of fluid zones in this manner; for example, four blade passages can be coupled using three mixing planes. Note that the stator and rotor meshes do not have to be conformal; that is, the nodes on the stator exit boundary do not have to match the nodes on the rotor inlet boundary. In addition, the meshes can be of different types (e.g., the stator can have a hexahedral mesh while the rotor has a tetrahedral mesh). — Figure 2 par AT? ENULUL GAIU IJLALUL PVULIIAIIILOS Consider the turbomachine stages shown schematically in Figure 2.2.1 and Figure 2.2.2. In each case, the stage consists of two flow domains: the rotor domain, which is rotating at a prescribed angular velocity, followed by the stator domain, which is stationary. The order of the rotor and stator is arbitrary (that is, a situation where the rotor is downstream of the stator is equally valid). In a numerical simulation, each domain will be represented by a separate mesh. The flow information between these domains will be coupled at the mixing plane interface (as shown in Figure 2.2.2 and Figure 2.2.1) using the mixing plane model. Note that you may couple any number of fluid zones in this manner; for example, four blade passages can be coupled using three mixing planes. Note that the stator and rotor meshes do not have to be conformal; that is, the nodes on the stator exit boundary do not have to match the nodes on the rotor inlet boundary. In addition, the meshes can be of different types (e.g., the stator can have a hexahedral mesh while the rotor has a tetrahedral mesh).

Related Figures (57)

vanes) to be converted into pump power output. Depending on the fluid flow pattern in multistage pumps and the impellers’ arrangement on the pump shaft, impeller design and arrangements are categorised as: single-stage, multistage, single-entry, double-entry, multiple-entry, in-line (tandem) or back-to-back arrangement. Axial and radial flow impellers are rotating industrial mixer components designed for various types of mixing. Both types of impellers are primarily constructed from stainless steel. Impellers impart flow. They serve the purpose of transferring the energy from the motor to the substance of a tank as efficiently as possible. Impellers are organized by their flow patterns. = An impeller (also written as impellor ) is a rotor used to increase (or decrease in case of turbines) the pressure and flow of a fluid. It has been used in variety of everyday equipment such as pumps, compressors, medical devices, mixing tanks, water jets and washing machines. More specifically, an impeller is a rotating component equipped with vanes or blades used in turbomachinery (e. g. centrifugal pumps). Flow deflection at the impeller vanes allows mechanical power (energy at the

Figure 1.3.3 Flow Direction of Three Different Pumps/Impellers (Courtesy of Global Spec) TENT BTR a a ae Oe ee A ee ee ee Are best for mixing applications that require stratification or solid suspension. Axial impellers are set up to create effective top to bottom motion in the tank. This motion is highly effective when placed over the center of a baffled tank. Some common types of axial flow impellers include: marine impellers, pitched blade impellers, and hydrofoils. Hydrofoil impellers are also known as high

reiterated into momentum theory. If a feasible propeller geometry is used this method rapidly converges to a solution. The induced flow speed can be obtained by reiteration of the thrust and torque determined from the 2D section into momentum theory. Convergence of the solver can be tested by different variables with common choices: the angle of attack, angle ~ or the induction factor. The method described below solves for axial and tangential induction factors via and vi. a and B are Angle of attack and geometric airfoil section pitch angle, respectively. U. denotes free stream velocity. For more details, please consult the chapter 3 of [Teeuwen ]¢.

Impellers in mixing tanks are used to mix fluids or slurry in the tank. This can be used to combine materials in the form of solids, liquids and gas. Mixing the fluids in a tank is very important if there are gradients in conditions such as temperature or concentration. Figure 1.5.1 shows two types o: impeller used in mixing tanks, namely:

injection, in the energy industry for pumping oil and natural gas or for operating cooling towers. In the medical industry, pumps are used for biochemical processes in developing and manufacturing medicine, and as artificial replacements for body parts, in particular the artificial heart and penile prosthesis. (see Figure 1.6.1), A pump is a device that moves fluids (liquids or gases), or sometimes slurries, by mechanical action. Pumps can be classified into three major groups according to the method they use to move the fluid: direct lift, displacement, and gravity pumps. Pumps operate by some mechanism (typically reciprocating or rotary), and consume energy to perform mechanical work by moving the fluid. Pumps operate via many energy sources, including manual operation, electricity, engines, or wind power, come in many sizes, from microscopic for use in medical applications to large industrial pumps. Mechanical pumps serve in a wide range of applications such as pumping water from wells, aquarium filtering, pond filtering and aeration, in the car industry for water-cooling and fuel

Figure 2.1.1 Single Blade Model with Rotationally Periodic Boundaries

compute a flow field that can be used as an initial condition for a transient sliding mesh calculation. This eliminates the need for a startup calculation. The multiple reference frame model should not be used, however, if it is necessary to actually simulate the transients that may occur in strong rotor- stator interactions, the sliding mesh model alone should be used.

4.1.2.1 Case study — Mixing lank In a mixing tank with a single impeller, you can define a rotating reference frame that encompasses the impeller and the flow surrounding it, and use a stationary frame for the flow outside the impeller region. An example of this configuration is illustrated in Figure 2.1.2. (The dashes denote the interface between the two reference frames). Steady-state flow conditions are assumed at the interface between the two reference frames. That is, the velocity at the interface must be the same (in absolute terms) for each reference frame. The grid does not move. You can also model a problem that includes more than one rotating reference frame. Figure 2.1.3 shows a geometry that contains two rotating impellers side by side. This problem would be modeled using three reference frames: the stationary frame outside both impeller regions and two separate rotating reference frames for the two impellers. (As noted above, the dashes denote the it

In sliding meshes, the relative motion of stationary and rotating components in a rotating machine will give rise to unsteady interactions. These interactions are illustrated in Figure 2.3.1, and generally classified as follows:

In the sliding mesh technique two or more cell zones are used. (If you generate the mesh in each zone independently, you will need to merge the mesh profiles prior to starting the calculation. Each cell zone is bounded by at least one interface zone where it meets the opposing cell zone. The interface zones of adjacent cell zones are associated with one another to form a grid interface. The two cell zones will move relative to each other along the grid interface. Be advised that the grid interface must be positioned so that it has fluid cells on both sides. For example, the grid interface for the geometry

Figure 2.3.3 Initial position and some translation with Sliding Interface

The first item on right-hand side expresses the local acceleration of the velocity field within the rotating frame of reference. The second term and third item denote the angular velocity acceleration and the convective term within the rotating frame of reference, respectively. While, the fourth item and last item are the Coriolis acceleration and the Centrifugal acceleration, respectively, which are fictitious forces produced as a result of transformation from stationary frame to rotating frame of reference. Figure 2.4.1 shows the directions of the velocity and the acceleration, and relationship between the absolute velocity, relative velocity and rotation (Schobeiri, 2005). Substituting the acceleration, for an incompressible flow equations of motion and energy, in rotating frame of reference can be obtained:

OEE oe ee ee we ee rete ee ee ee eee ee ee Sufficient ventilation of a parking garage depends on the correct fan selection. Fans are selected by deciding if the air flow rate determined by intersecting the fans’ H-Q characteristic curves with the system resistance curve is sufficient for the ventilation of the space. For that purpose, database of various fans’ performance curves and static pressure loss vs. airflow rate curve of the ventilated system are needed. Fan databases are setup by testing the performance of fans in test systems that are designed according to the national and/or international standards. One of the commonly used measurement systems is AMCA 210 test rig shown in Figure 3.1.2. The flow rate through the fan is changed by adjusting the flow resistance, i.e. nozzle diameter, which is located in the variable exhaust system of the AMCA chamber. 1e CFD solution domain consists of one slice of the axial fan around a blade, which is extended in th upstream and downstream directions to improve the solution accuracy, as described below. ter CAD geometry cleanup and generation of surface and volume mesh required by the Multiple ference Frame, MRF, method, the grid is read into the computational fluid dynamics (CFD) ftware. Then, the fluid properties, turbulence model, wall functions, fan rpm and boundary nditions are set. Results are obtained by a parallel CFD solver and post-processed to get visual and meric data about the flow field. tA enliuitinn anf 4 74¢c_p ic W1cad ac tha initial qiace far tha navt nno tharahry 4 naramatrir ctidyu anf tha

Table 3.1.1 Mesh Resolution and Number of Cells

Table 3.1.2 Solution Methods and Boundary Conditions where Q is the volumetric airflow rate in m3/s, AP is the fan pressure rise in Pa, n is the fan speed in rpm, and T is the sum of the fan aerodynamic and mechanical friction torques in Nm. In Figure 3.1.5, numerical torque values are almost half of the experimental ones because the losses caused by the bearing and transmission friction torques are ignored. Even though there are significant variations in the previous charts, simulation results of fan static pressure rise vs. volume flow rate curves are quite similar to the test result, because mechanical friction of bearings and the transmission do not affect these parameters. Following formula is used to calculate the electric motor torque in terms of the motor power and fan rpm.

Alternative technologies have been proposed for the improvement of Jetfan installation factors, including deflection vanes (with or without a clearance between the Jetfan outlet and the vanes, Lotsberg (1997)), the Banana Jet (slanted silencers, Witt and Schiitze (2008)) and the MofJet (a development on the original concept using silencers with an inclined outlet, Tarada (2018)). Each of these technologies implies different space requirements, power consumption, installation factors and induced tunnel velocities. Bench thrust tests were simulated for a conventional Jetfan and a MoJet to establish baseline values for the fan mass flow, power absorbed and axial thrust. The technologies were then tested in a tunnel environment and the values for the shaft power, axial thrust, tunnel airflow velocity and installation factor were calculated. The limitations of using the concept of installation factor in a strongly swirling, 3D flow-field in a tunnel are discussed.

Boundary layer meshes were attached to all solid surfaces to resolve sharp velocity gradients. Within the fans and silencers, these boundary layer meshes were calculated via 30-layer prisms with an initial layer approximately 0.1 mm thick within a Jetfan, and with an expansion factor of 1.2 (Figure 3.2.2). Along the tunnel, the boundary layer mesh was 10-layer with an initial layer 0.3 m thick and an expansion factor of 1.1. A typical y+ value for the last cell (downstream) along the (rough) tunnel surfaces was 315, while a typical y+ value on the fan blades and silencers was 15. Table 3.2.1 lists the number of CFD cells used for each type of technology tested with M = million. OTT? Mee ERA MO Na ee: Si A ee MR oes Re ee MA! A Ne at PA See germme: CM Beene tunnel is 6.75 m, and is 16 m wide. rhe detailed geometry of the Jetfan hub, blades and motor was included in the same CFD model as the tunnel geometry. Blade rotation was simulated using the multiple frame of rotation option in ANSYS CFX, with circumferential averaging. The mass flow through the Jetfans was not prescribed put was calculated through the CFD code based on a blade rotational speed of 1500 rpm, corresponding to the speed of a four-pole motor. Reductions in blade rotational speed due to induction motor slippage were not accounted for. The k-w SST model of turbulence (Menter, 1993) was employed, which assisted in identifying any areas of flow separation. The tunnel walls, floor and soffit were assumed to have a uniform sand roughness height of 0.03 m, giving an equivalent friction factor (A = 4f) of 0.024.

the shaft power absorbed by the conventional Jetfan. The deflector vanes included a flat section followed by vanes set at 25° from the horizontal. Following the fan manufacturer’s guidance, the deflector vanes were installed at a distance of 0.504 m from the silencers at both ends of the Jetfan (Fiogure 3.2.3). was assumed for all four Jetfan technologies. The silencers attached to the fans were selected with 100 mm thickness, and were arranged to be 150 mm below the tunnel soffit. The Jetfan internal diameter for the conventional Jetfan, deflection vanes and Mojet was set at 1.25 m. A smaller Jetfan diameter (1 m) had to be selected for the slanted silencers (with a 7° angle to the horizontal) to keep the Jetfan to just within the 1.7 m installation height limit. All four technologies had “2D” lengths of silencers installed on either side of the fan, where “D” is the internal diameter of the fan. A mid-range blade pitch angle of 33.4° was selected for the initial calculations with the 1.25 m diameter Jetfans. A blade pitch angle of 39° was also tested for the MojJet, to approximately match

Tunnel Installation Results

oo The jet discharged from a conventional Jetfan adheres to the tunnel soffit due to the Coanda effect, causing a loss of in-tunnel thrust due to friction between the jet and the soffit. When installed with 150 mm clearance to the tunnel soffit, the mass flow through the conventional Jetfan is reduced by 6.1% from its bench test value, due to the confining effect of the soffit on the inlet silencer. Taking the reduced mass flow into account, the installation factor calculated from the CFD results is very close to the value provided by the Kempf correlation (1965):

Figure 3.2.6 Axial and Circumferential Velocities at Discharge from Conventional Jetfan Silencer with Struts (33.4° Blade Pitch Angle) where Da is the outlet diameter of the Jetfan and z denotes the distance between the center axis of the jet at the outlet and the tunnel wall. This result is somewhat surprising, since Kempf did not

The MoJet results for a 33.4° blade pitch angle show a 24% increase in the in-tunnel thrust compared to a conventional Jetfan, with 15% less absorbed power. In order to compare the MofJet with a conventional Jetfan of equivalent power consumption, the MoJet blade pitch angle was increased from 33.4° to 39°. This enhances the MoJet in-tunnel thrust to 161% of the conventional Jetfan in- tunnel thrust, with slightly less power consumption than the conventional Jetfan. Removing the struts (i.e. replacing them with drop-rods) reduces the installation factor slightly, but installation factor is an unreliable parameter for prediction of this complex swirling flow-field. Using drop-rods EEE, SNe! BM cer em Bayes: The increase in the cross-sectional area of the MoJet silencer leads to a significant pressure recovery downstream of the fan, and to a reduction in the discharge velocity. Approximately 22% of the kinetic energy of the flow is recovered as static pressure in the MoJet outlet silencer. This pressure recovery leads to an increase in the force exerted by the MoJet onto the tunnel air, while the reduction in discharge velocity reduces the aerodynamic friction between the jet and the soffit. The Coanda effect is also reduced by the effect of the bell mouth in turning the flow away from the tunnel soffit. These factors explain the high (near-unity) installation factors reported for the MoJetin Table 3.2.3.

Figure 3.2.9 Velocity Contours with Deflectors (Color Legend as per 3.2.19) The result with deflectors shows that the swirl in the discharge flow is eliminated, and the flow is efficiently directed downwards . However, the jet then attaches to the tunnel floor, and significan’ friction is generated there. The installation factor is therefore rather poor at only 0.60, and the tunne Jetfan with Deflectors reduction in tunnel air velocity. The loss of thrust due to a reduction in fan diameter with slanted silencers would occur in any height-restricted space, and is not a result of the 1.7 m clearance assumed in this study. However, due to the reduction in fan diameter to 1 m, the thrust/power ratio with this option is better than any of the 1.25 m diameter fans tested here.

A combination of anisotropic and isotropic triangles were used in the surface mesh discretization. Areas of high curvature - such as the leading edge, trailing edge, and the tip were resolved by utilizing Pointwise’s T-Rex algorithm. This tool grows anisotropically stretched, right-angled triangles layered in the normal direction to a boundary [5], as shown below. Using this, areas of high curvature are able to be resolved without the need to isotopically refine the area. The result is an accurate adherence to the surface and a reduction in the point count. The interior of the surface mesh was resolved with isotropic triangles created using a modified Delaunay algorithm. (Figure 3.3.2).

After meshing the geometry, the outer boundary of the moving reference frame (MRF) as well as the outer boundary of the computational domain were meshed utilizing isotropic triangles and the Delaunay algorithm. The MRF is cylindrical domain approximately 4.8 D long (4.8 x the propeller diameter) and 1.5 D in diameter (Figure 3.3.3). It starts just upstream of the propeller and extends downstream into the wake. A far field block was generated outside of the MRF corresponding to 10 D and 2.6 D; these were the limits taken from the file provided.

count of the mesh without sacrificing quality. Once the near-wall viscous mesh was generated, the remainder of the volume was populated with isotropic tetrahedral cells. The total cell count was just below 4.1 M8 cells. The average maximum included angle was 101, and the maximum was 170. The average volume ratio was 1.8 with a maximum of 28.

quantities, such as propeller efficiency, thrust coefficient, torque coefficient, forces, force coefficients, flow rates, and much more. All these quantities are evaluated throughout the simulation run, and all the important data is summarized in tabled .csv files as well as in the HTML report, which can be updated anytime during the simulation for every run. Furthermore, visual postprocessing of the volume fields can be done with ParaView. (Figure 3.3.4 - Figure 3.3.5).

Figure 3.3.6 Comparison of TCFD vs. SVA using Pointwise Mesh

ee a resulting in an additional source term qcae besides the continuity source term q., the momentum source term qm, and the energy source term qe. Note that the APE are formulated in a rotating frame of reference. The time-dependent rotation of the reference frame is only accounted for by adding the Coriolis acceleration term in the momentum equation [20]. Because of the extension to a non- uniform mean flow field, convection and refraction effects are taken into account by the left-hand side operator. The results from the large-eddy simulation determine the mean ow quantities and the sources To accurately resolve the acoustic wave propagation described by the APE-4 formulation, the acoustic perturbation equations are discretized in space using the 9-point 6 order dispersion- relation preserving (DRP) scheme proposed by Tam & Webb [21] in the interior domain of integration and a special formulation of the DRP summation by parts scheme (SBP-DRP)[22] on the boundaries. A 4 order alternating 5-6 stage low-dissipation and low-dispersion Runge-Kutta method originally proposed by Hu et al.,[23] and modified by Stanescu and Habashi [24] is applied for the temporal integration. A detailed description of the two-step method and the discretization of the Navier-Stokes equations and the acoustic perturbation equations is given by Ewert and Schroder [25].

Figure 3.4.3 illustrates the performance curve of the axial fan showing the pressure coefficient P= Ap/[(t2/2) pD2.n2] versus the flow rate coefficient = 4V/(m2D3,n) for the tip-gap width s/D, = 0.001, s/Do = 0.005, and s/Do =0.01, where Ap = Pstatout- Po,in is the difference of the static pressure at the outlet pstatourand the stagnation pressure at the inlet pojn.

Figure 3.4.4 Instantaneous contours of the A2-criterion inside the ducted axial fan configuration color coded by the relative Mach number showing the vorticial structures generated by the tip leakage flow at & = 0:195; (a) s/Do=0.001; (b) s/Do=0.005; (c) s/Do=0.01.

Figure 3.4.5 Development of the tip-leakage vortex and Mach number contours in several radial planes ranging from 8 = 20° to 8 = 706; (a) s/Do=0.001; (b) s/Do=0.005; (c) s/Do=0.01

Figure 3.4.6 Tip vortex core trajectory and turbulent kinetic energy contours in several radial planes ranging from @ = 20¢ to 8 = 700; (a) s/Do=0.001; (b) s/Do=0.005; (c) s/Do=0.01

Figure 3.4.7. Radial distributions inside the various tip-gap width for 6 = 45o; (a) relative axial velocity; (b) relative Mach number; (c) turbulent kinetic energy

Figure 3.4.9 The multi-block structured mesh in the acoustic sourc region resolving one out of five blades of the axial fan; (a) overall mesh; (b) detailed topological view of the mesh

Figure 3.4.11 Instantaneous contours of the fully developed acoustic field; (a) s/Do = 0.001; (b) s/Do = ei Barina heli ani 0.005

Figure 3.4.12 Evaluation method for the microphone data; (a) fixed frame (b) and rotating frame; schematic of the virtual microphone positions for the three acoustic configurations; (c) side view; (d) front view. Note that the outer casing wall and the AMCA nozzle are not shown.

Figure 3.4.13 The evaluation method to locate the microphone positions in a fixed and rotating frame of reference frame Simulation time Microphone position

Figure 3.4.16 Sound spectra on the circle C1 for several tip-gap widths; (a) CAA results (b) CAA results compared with measurements

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