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Text | Dovinci
One of the problem areas of fluid flow in turbomachinery is its inlet area, which manifests itself as flow deformation due to induced fluid vortices, accompanied by incorrect flow incident onto the impeller.
In addition, the hub is one of the main components of many turbomachinery, and it has been found that no significant research has been done on the geometric modifications used to enhance performance in centrifugal fans. This is partly due to the fact that the designers tried to reduce the cost of the entire machine.
The flow behavior in the inlet area of the impeller of a centrifugal fan is generally considered to be very complex, with pre-rotation, vortices, and strong fluid boundary interactions. These behaviors not only lead to incorrect incidence,
It also results in a loss of undesirable flow at the impeller inlet, resulting in a reduction in the overall efficiency of the fan. In order to capture the flow loss phenomenon in the vortex zone at the inlet of the fan, a detailed experimental analysis is often required.
Due to inlet flow deformation, appropriate modifications must be made to the inlet flow geometry to improve the flow characteristics. Various researchers have earlier made many attempts at inlet geometry interventions to alter flow behavior and thus potentially improve wind turbines across the board.
Research methods
Description of the centrifugal fan experimental test bench
A schematic diagram of the centrifugal fan scale used in this study is shown in the figure. 1. The geometry of the fans used closely matches the fan model used by Meakhail and Park (2005), so the results of this study can be verified.
The fan consists of four structural parts, namely the inlet part, the blade impeller, the blade diffuser and the rectangular volute. The inlet section with a diameter (D) of 200 mm is used as a reference size to represent all other dimensions of the fan. A 1.15 D inlet duct is provided to ensure that the speed curve at the fan inlet meets the requirements.
The shroud impeller swept backwards and had thirteen equidistant blades. The inlet and outlet diameters of the impeller are 1.2 D and 2 D, respectively. The blade angle at the impeller inlet is set to 30° and the blade angle at the outlet is set to 76°.
Each blade has a rounded leading edge with a thickness of 0.025 D. The impeller channel is 0.175 D wide in the axial direction. The impeller is followed by a diffuser with the same number of blades, width and blade thickness as the impeller.
The inlet diameter of the diffuser domain is 2.3 D, the outlet diameter is 3 D. The inlet blade angle at the diffuser inlet is maintained at 23°, while the inlet angle at the vane outlet is 38°.
The width of the diffuser channel is the same as the width of the axial impeller. A small gap space of up to 0.15 D is provided between the rotating impeller and the fixed diffuser to partially reduce transient turbulence before the fluid enters the diffuser channel.
The channel height of the volute is 0.45 D. The flange width is maintained at 2.25 D at the volute outlet. Figures 2 and 3 show the experimental setup of the centrifugal fan test bench and its close-up view, respectively.
The experimental process includes measuring air flow (Q), static pressure, power input to obtain overall efficiency. The experiment is repeated several times to obtain consistent fan output data. The standard performance is then plotted using the experimental values.
for reliable experimental data. Both static and stagnant pressures are recorded at the inlet section of the fan and at the outlet of the diffuser, using a digital manometer with a resolution of 0.01 mm water. The speed of the fan is measured by a non-contact digital tachometer with a resolution of 1 RPM.
The experiment was initially conducted on a test bench and there were no abbreviated attachments. This arrangement is considered to be the hubless basic configuration in this study. The fan driven by a 3HP DC motor operates at a rated speed of 1500 RPM.
A butterfly valve set at the outlet pipe is connected to the end flange of the volute chamber to regulate the air flow rate, as shown in the figure. 4. Static and stagnant pressure is measured using a pitot tube located at the fan outlet pipe. It is also used to obtain the outlet speed by means of a removable traversing mechanism.
Determine the design point operation of the fan
The various volume factors corresponding to the hubless base configuration are shown in Figure 5. Due to system limitations, the volume factor cannot be increased beyond the maximum value of 0.051.
Overall efficiency (ηov) obtained from experimental work is defined as the ratio of the static energy transfer rate from the rotor to the fluid to the actual electrical energy rate of the input fan.
where Δp is the difference between the static pressure of the fan outlet (P4) and the fan inlet (P1), ηt is the transmission efficiency, ηm is the motor efficiency, and P is the power input of the motor.
It can be noted that the overall efficiency obtained from experimental work depends on the input energy provided by the motor. It is not enough to measure the electric drive energy of the input, as bearing and transmission losses are considered. Therefore, in order to solve these losses, the motor and transmission efficiency are increased by 80% and 90%, respectively, according to the panel specifications on the motor.
An increase in overall efficiency was observed for the range of volume factors selected for the experiment. At a maximum volume factor of 40.5, the peak total efficiency is approximately 0.051%. Therefore, the corresponding mass flow rate of 0.236 kg/s is used as the design point of the fan.
A description of the performance parameters used in the analysis
The performance variables, i.e. volume and head coefficients, are calculated experimentally using equations. (2) and (3) are shown below, respectively.
where the outlet of the p3 diffuser is static pressure, U2 is the tangential velocity at the exit of the impeller, ρ is the air density, and R0 is the outlet radius of the impeller.
Similarly, the relative theoretical efficiency (ηRt) is derived using equation (4) and is defined as the ratio of the specific energy obtained from experimental work to the specific energy obtained by Euler's equation.
Uncertainty analysis of experimental work
It is clear that uncertainty is inevitable for any measurement in an experimental study. In order to estimate the measurement error in this experimental work,
The overall efficiency of the dependent variable selected for uncertainty analysis is determined experimentally. The pressure head (y1 mm water) used to measure the static pressure rise of the fan, the air flow rate (y2 mm water), the power input (IP) of the fan driven by the DC motor are considered to be measured variables for uncertainty analysis. Table 1 shows the uncertainty values corresponding to the selected measurement variable.
Describe the various hub geometries used in the experiment
Typically, the direction of fluid entering the impeller changes sharply and abruptly from axial to meridian, followed by a loss of flow in a hubless base configuration, as shown in the figure. 6.
The purpose of this analysis is to examine the flow characteristics of variable hub geometries by providing a hemispherical or elliptical hub configuration, as shown in the figure
Experiments with two hemispherical hubs (Figure 7) and oval hubs (Figure 8) are performed to highlight the basic flow characteristics in each case relative to the hubless base configuration. Figures 7b and 8b show the respective wooden models of the hubs manufactured for experimental work. This wooden hub is completely fixed to the back plate of the impeller and rotates with the impeller.
The geometric parameters used to influence the change are expressed as either ball hub ratio (RS) for hemispherical hub configurations or elliptical hub ratio (RE) for ellipsoidal hub configurations. These ratios are non-dimensionalized in terms of the radius (r) of the inlet pipe given in EQS.
The spherical and elliptical hub ratios used in the analysis are shown in Table 2. The height of the elliptical hub in the meridian direction (d) is obtained by optimizing the hemispherical hub geometry, as further explained in the section "Effect of hemispherical hub configuration on fan performance".
Table 2 Hub layout configuration
and discuss
This section details the results of the experimental study. This performance is performed to obtain the main and operational characteristics of the different hub geometry configurations. Therefore, the following discussion considers the total efficiency, head coefficient, and relative theoretical efficiency corresponding to the volume factor selected at rated speed for the hubless base configuration and the dual-hub geometry configuration.
The effect of the hemispherical hub configuration on fan performance
Figure 9 shows the change in overall fan efficiency for the various dome hub configurations shown in Table 2. The study found that the presence of the hub improved the overall efficiency performance of the fan at all volume factors.
This is obviously because basically the hub provides better guidance for air flow at the fan inlet compared to hubless configurations at the corresponding volume factor. Overall, it can be seen from the graph that the overall efficiency of the various hemispherical hub configurations is on average about 2.6% higher than that of the hubless base configuration under the design volume factor.
However, it can be seen that the hemispherical wheel configuration S3 can relatively improve the overall efficiency of the fan at all volume factors. Presumably, this configuration helps to recharge the flow at the fan inlet,
This significantly reduces inlet flow loss and improves static pressure. Therefore, the hemispherical wheel configuration S3 is considered a useful optimal configuration to improve overall efficiency performance.
In addition, configuring S5 and S3 produces lower overall efficiency compared to S4 configurations. This is because the inlet at the entrance can be restricted due to the protruding hub, which can lead to accelerated flow, resulting in a loss of flow.
The main features of the hemispherical hub configuration versus the hubless base configuration are shown in the figure. 10. It was found that as the volume factor increased, the decrease in the head coefficient confirmed the theoretical trend required for a typical backcurved impeller.
The study found that the head factor appears to be higher at all volume factors for any hub configuration compared to the hubless base configuration. In general, it is observed that at the design point volume factor, the head coefficient value is relatively high relative to the hubless base configuration, about 0.6%.
Resources:
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