Abstract
Axial-flow fan with advantages such as large air volume, high head pressure, and low noise is commonly used in the work of air-conditioner outdoor unit. In order to investigate the internal flow mechanism of the axial-flow fan with different trailing edge structures of impellers, four kinds of impellers were designed, and numerical simulation and experiment were deployed in this article. The pressure distribution on the blades surface and distribution of vorticity in impellers were obtained using numerical simulation. Distribution of blade loading and velocity at the circumference are discussed. The relationship between the wideband noise and the trailing edge was established based on the experiment results. The results show that after the optimization of the trailing edge structure, the distribution of vorticity near the trailing edge of the blade is more uniform, especially at the trailing edge of 80% of the chord length of the suction surface. From the blade height position of 70% to the impeller tip, the pressure on the surface rapidly increases due to the tip vortex and the vortex shedding on the blade edge occurred in the top region of impeller. The pressure fluctuation amplitude at the trailing edge structure of the tail-edge optimization structure is smaller. In the distribution of blade loading, the three tail-edge optimization structures have smaller pressure fluctuations and pressure differences at the trailing edge structure. It is extremely important to control the fluctuation amplitude at the trailing edge. The amplitude of low-frequency sound pressure level of optimizing the trailing edge structure decreases obviously in the range of 50–125 Hz, and the optimization structure of trailing edge has an obvious effect on low-frequency wideband noise.
Introduction
In the air conditioning system, the axial-flow fan is generally used in the air-conditioner outdoor unit because of it has many advantages, such as compact structure, large air volume, high head pressure, and low noise. Among them, minimal noise is a significant indicator. Due to the shroud only exists near the trailing edge of the tip, it is similar to the semi-open axial-flow fan. There are many differences between the design method and the internal flow phenomenon compared with the traditional industrial axial-flow blades, especially in the region near the blade tip and the trailing edge of the blade.
With the continuous development of industrial technology, higher requirements are put forward for noise of axial-flow fans. The broadband component of the noise spectrum is primarily induced by the pulsating force with random characteristics, such as vortex shedding at the tail of the blade, tip vortex, and turbulent flow. Furthermore, the vortex shedding at the blade trailing edge is one of the principal causes of eddy noise. How to deal with the trailing edge of the axial-flow fan blade reasonably and effectively improve the pressure gradient distribution in the air duct and restrain the influence of the secondary flow and the wakes on the downstream flow can effectively reduce the eddy noise of the fan. And, improving the axial-flow fan flow characteristics and reducing aerodynamic noise help to improve and get higher an aerodynamic performance, to provide a basis for improving the aerodynamic characteristics.
Study of the literature1,2 has shown that the serrated edge of the axial-flow fan blade can reduce the wake of the axial-flow fan blades downstream, which reduce aerodynamic noise and improve sound quality. The contribution of aerodynamic sound sources in the sound field on various parts of outdoor units of air conditioners is predicted and the noise reduction effect of serrated axial-fan blades is analyzed. 3 The predicted result is in good agreement with the experiment qualitatively, which shows that the numerical analysis method of noise has good engineering application value. Gong et al. 4 designed two groups of saw tooth blade trailing edges with different shapes and sizes. The influence of serrated trailing edge of the aerodynamic noise of the fan was fabricated and tested. The results show that the trailing edge of a saw tooth has significant noise reduction effect and sinusoidal serrated triangle better. The structure of blade trailing edge improves the distribution of the pressure gradient in the flow path and it restrains the influences of both secondary flows and weakened wakes in the downstream flow.5,6 The leading edge and trailing edge of axial-flow fan are reasonably designed and the internal flow is improved, widening the wake area and speeding up the breakup of large vortices.7–9 The eddy shedding noise of the leading edge of the axial wind turbine blades was analyzed and discussed in detail. 8 This study pointed out that the middle chord of the axial wind turbine was serrated, which could make the boundary layer become turbulent, thus effectively reducing eddy noise. The noise reduction effect of serrated blades at the outlet of centrifugal fan blades was studied based on experimental method. 10 The serration at the outlet of centrifugal wind turbine is able to reduce the noise by 2.5 dB(A). The study shows that the non-smooth surface of the fan can increase the adsorption capacity of the suction surface (SS), and the flanging of the fan’s outer edge is helpful to reduce the noise. 11 Perforation design of the blade was performed. 12 Static and aerodynamic acoustics characteristics of the perforated blade were analyzed by means of simulation. The perforated design was proposed to effectively reduce the noise. Spectrum distribution characteristics of radiant noise on different shapes and surfaces of blades were investigated to find out the main influencing factors of aerodynamic noise and to provide reference for the design of low-noise fan. 13 The trailing edge depression can effectively improve the pressure distribution on the blade surface, reduce the pressure difference between the pressure surface (PS) and the SS, and effectively improve the internal flow of the impeller.14,15 Most scholars only focus on the structural factors that affect the aerodynamic noise of wind blades and the research on the causes of noise.16–19 But the location of noise sources and effective solutions is still not deep.
Some common theories have been explored, but a few of means and mechanisms do not apply to internal flow characteristics and pressure fluctuation characteristics of the trailing edge of the axial-flow fan blades. In this article, the numerical analysis research on the trailing edge structure of low-pressure axial-flow fans for air conditioners is carried out for the sake of knowing the distribution of blade loading and velocity at the circumference; characteristics of pressure fluctuation and the trend of circumferential velocity distribution are discussed, and the effect of blade trailing edge structure on broadband noise is experimentally studied.
Numerical methods
Calculation model
In this study, the research object is the air-conditioner outdoor unit. Four impellers of axial-flow fans were designed to study the effects of the trailing edge of the blade on the fan’s performance and internal flow characteristics. Figure 1 shows the specific shape impellers with different trailing edges. Impeller 1 has the original trailing edge; while the impellers 2 and 3 have tooth-shaped blade trailing edge 1 and bionic blade trailing edge, respectively; impeller 4 has large tooth-shaped blade trailing edge. Among them, the bionic trailing edge of the blade refers to the wings of a flying bird at high altitude, while the trailing edge of the larger tooth-shaped blade resects the low-efficiency site at the trailing edge of the blade. The blades rotate in clockwise direction. In order to obtain the accurate flow in impeller, the three-dimensional (3D) flow domains including the inlet section, impeller, flow passage of the fan, condenser, motor bracket, outlet section, and structural simplification were made irrespective of the effect of the downstream exit grille, but the effect of side inlet airflow was taken into account. The structure of the outdoor unit of air conditioner is shown in Figure 2. The fan impeller is designed to rotate at n = 850 r/min. Specifications of the fan are given in Tables 1 and 2.
The trailing edge of the impellers: (a) impeller 1, (b) impeller 2, (c) impeller 3, and (d) impeller 4.
Geometric specifications of the fan.
Aerodynamic specifications of the fan.
The structure of the outdoor unit of air conditioner and computational domain.
Grid independence
After being modeled in Creo 3.0, the calculation domains are imported to ICEM CFD 14.5 for mesh generation. The mesh grids are not only expressive forms of geometry but also the carrier of simulation and analysis. 20 With the increase in the grids number, the error caused by the grids will be reduced gradually. Grid independent study is conducted by solving the flow of nine sets of grids with different numbers varying from 2.32 to 8.85 million, as shown in Figure 3. The flow rate changes are compared under nine different grid numbers. It can be seen that the flow rate ascends with the increase in the grid number, but almost keeps constant after the grid number exceeds 2.96 million. Considering the simulating precision and designing period, the total number of grids in this article is 2.96 million.
Mesh independence analysis.
Turbulence model
Owing to the airflow of shedding and recycling in transient flow process, the selection of reliable turbulence models is highly important for simulating the performance more accurately.
3D steady and unsteady internal flow analyses were performed by numerical simulation using the software ANSYS FLUENT. The realizable k-ε model was applied to the steady-state calculation, instead of the unsteady calculation to select the large eddy simulation and the sub-lattice Smagorinsky–Lilly model. In order to speed up the calculation of the convergence rate, a polyhedral mesh is used. The steady and transient calculations are carried out in the same computational domain, and the steady results are used as the initial conditions for transient simulation.
Boundary conditions
The whole flow passage of the air-conditioner outdoor unit is taken as the computational flow domains. The surface of the flow channel is configured as a no-slip wall. The zone in the impeller is defined as a rotating zone. The link between the blade and the stationary components is the interface. Given a boundary condition, the total pressure given to the inlet is 0 Pa and the corresponding static pressure at the outlet is 0 Pa.
Monitoring points
In order to make a clearer analysis of the flow characteristics of the fan, there are 12 monitoring points evenly arranged in the shroud as shown in Figure 4.
Distribution of the monitoring point.
Experiment
Noise test rig
Noise test rig was built based on international standards GB 6884-1986 “acoustic noise source sound power level determination anechoic chamber and semi-anechoic chamber precision method.” The sound-absorbing chamber was tested with a glass wool tipped sound-absorbing material, and the floor was in a semi-anechoic chamber with a vibration-reducing floating structure. The background noise was 18 dB(A).
The noise signal was obtained using LMS acoustic measurement and analysis system (Figure 5). The system includes hardware and software in two parts. Hardware components include front-end data acquisition, sound pressure sensor, sensor preamplifier, and calibrator. The software uses LMS test, a lab spectral acquisition noise analysis module. The sound pressure spectrum was obtained in one-third octave bands by acoustic sensors: GRAS Type46AE and Type 26CA. And then the sound pressure was converted to A-weighted sound level. The sampling time and frequency were 600 s and 4096 Hz, respectively.
Noise measurement equipment and real impellers.
Air volume test rig
The flow rate was obtained based on GB/T 1236-2000 “industrial fan with a standardized duct performance test” standards. The test system mainly consists of a spherical probe, hydrostatic meter, rectifier, temperature and humidity sensor, differential pressure gauge, nozzle, pneumatic plug, and other components (Figure 6). Multi-nozzle flow test system is suitable for air-based centrifugal or axial fan for aerodynamic performance experiments.
Air volume equipment.
Results and discussion
Comparison of external characteristics between simulation and experiment
The impeller rotational speed is 850 r/min, which is driven by the motor, adjusting the speed to get the flow rate of bonus points, and the total pressure predicted by numerical simulation results are in agreement with the experimental results, as shown in Figures 7 and 8, respectively. And, the error of flow rates between experiment and simulation is below 5%. The simulation and experimental errors are within the range of acceptability. It shows that numerical simulation is more accurate for axial-flow fan performance prediction.
Flow rate comparison between simulation and experiment.
Pressure comparison between simulation and experiment.
Inflow characteristic analysis
Distribution of velocity at the circumference in different times
The last period of the rotation cycle of the original impeller is divided into six moments, from T1 to T6. Velocity expansion can be used to observe vortex changes at the rear of blades. At 90% of the blade height on the circumferential surface, the speed expansion is shown in Figure 9. The top side of the figure is the inlet of the impeller, the lower side is the outlet of the impeller, the impeller on the side close to the exit is the PS, and impeller on the other side is a SS. From the figure, the impeller is clockwise, and the three runners are numbered as runners 1–3 from left to right. As the fan inlet includes two rectangular entrances on the front and sides, and the exit is circular, the asymmetry of the import and export causes the internal flow asymmetry. The distribution of velocity at the trailing edge of the blade flow path gradually increases. With the rotation of the blade, a confined vortex is established. T5 at a time when the flow channel at nearly 80% of the blade trailing edge which the high-speed area increases, the flow past the blade trailing edge gradually dissipated into a number of the high-speed vortices. The study found that a part of the high-speed vortex collapses in the flow channel dissipated. It shows that shedding eddies trained at the trailing edge of the blade to eventually evolve into inefficient stall eddies after one cycle of enhancement and gradual decrease. It affects the flow in the adjacent channels and creates new vortices in the adjacent channels. Lift ripples brought about by the trailing edge vortex shedding are important causes of vortex noise.
The distribution of velocity at 90% blade height at different time.
Distribution of pressure on the surface of the impellers
Comparing the pressure distribution on the PS and SS of the four groups of impellers, as shown in Figures 10 and 11, the pressure gradient changes greatly in the blade tip region. The maximum pressure rise of the blade surface is right on the blade starting from the blade leading edge 80% of the direction of the chord line, especially near the top of the shroud inlet area that the most intense pressure. After optimization of the trailing edge structure, the variation of pressure gradient on the blade surface is fairly uniform, especially in the shape of the trailing edge which the area of the low-pressure area is larger. The pressure difference between the SS and the PS is more uniform, and the stress distribution of the blade is improved.
Distribution of static pressure on the PS for different impellers: (a) impeller 1, (b) impeller 2, (c) impeller 3, and (d) impeller 4.
Distribution of static pressure on the SS for different impellers: (a) impeller 1, (b) impeller 2, (c) impeller 3, and (d) impeller 4.
Distribution of vorticity at the impellers and shroud
Figures 12 and 13 show vorticity distribution for different impellers in the last moment of the impeller cycle. The vortices in the flow field cause larger pressure fluctuations and become noise sources. Corresponding to the pressure fluctuation region, the large eddy region is also located near the tip of the blade. The SS of the front and trailing edge has larger vorticity area, and the vorticity distribution is irregular, which at other positions on the blade surface is significantly reduced. This indicates that the tip vortex and wake shedding vortex are the main factors that cause the change of the vorticity of the axial-flow fan. Especially in the circumferential direction, the trailing edge is squeezed by the shroud and the vorticity is obviously increased. In the middle of the impeller, the absolute value of the trailing edge vortex strength decreases gradually. The gradient decreases and the flow becomes more uniform. It shows that the optimization of the trailing edge structure improves the flow state in the blade tail and reduces the wake strength.
Distribution of vorticity on the SS for different impellers: (a) impeller 1, (b) impeller 2, (c) impeller 3, and (d) impeller 4.
Distribution of vorticity on the PS for different impellers: (a) impeller 1, (b) impeller 2, (c) impeller 3, and (d) impeller 4.
Through the optimization of the trailing edge structure, the eddy current distribution on the surface of the blade has been changed, and the distribution of vorticity near the trailing edge of the blade is more uniform, especially at the trailing edge of 80% of the chord length of the SS.
Distribution of blade loading
Distribution of blade loading at different spans
Blade load distribution of unsteady calculation on the blade height positions of 10%, 30%, 50%, 70%, and 90% of the original impeller surface are shown in Figure 14. The load distribution is very similar to that along the blade height direction from the hub to the blade tip. The load distribution on the PS and the SS at different blade heights is gradually increased. 10%~50% of blade height, the lift, and total pressure values produced by the blades are basically the same. From the blade height position of 70% to the top, due to the tip vortex and vortex shedding on the blade edge, the pressure difference rapidly increases on the blade surface. The results demonstrate that at the trailing edge of the blade, the pressure fluctuation amplitude, and the turbulence in the flow field are enhanced. It is extremely important to control the fluctuation amplitude at the trailing edge.
Blade loading at impeller 1 on different spans.
Distribution of blade loading for different times
Load distribution of the original impeller at different times corresponding to the three blade heights are shown in Figures 15–17. The time interval is equal to the time of impeller rotating 60°, so the load distribution is extracted a total of six times for a cycle. It is can be seen that the load distribution is varied for different times. When the blade rotates at different times, the pressure fluctuation is weak at the blade height of the 10% and 50%, and the difference of the total pressure is small. At 90% blade height, the pressure fluctuation is firm. The tip clearance leakage vortex occurs at the leading edge of the blade, resulting in a rapid increase in the pressure fluctuation at the position. The pulsating force with random characteristics is the major component of the wideband noise in the noise spectrum that includes the vortex shedding at the trailing edge of the blade and the turbulence of the tip vortex. The results show that the amplitude of the pressure fluctuation of the same blade height is similar during the rotation of one blade. However, due to the squeezing of the shroud, the pulsation near the blade height of 90% is noticeable. The optimized position of the tail-edge structure will be obtained near the blade tip for the better effect.
Blade loading at 10% blade height for different times.
Blade loading at 50% blade height for different times.
Blade loading at 90% blade height for different times.
Distribution of blade loading on different impellers
The blade surface pressure distribution at 70% and 90% blade heights was extracted, and four different blades were distributed together, as shown in Figures 18 and 19. At 70% and 90% blade heights, it can be observed that the difference of the blade edge trailing results in the differences of the surface pressure distribution at the blade. As the flow separation leads to a larger pressure value of the leading edge of the blade, the pressure rapidly decreases to a certain value and changes smoothly along the chord of the blade. Then, along the blade chord direction, the total pressure gradually increases. In the overlapping position of the blade and the shroud, the pressure difference greatly increases. At 80% chords, the pressure difference reaches the maximum value. It shows that the pressure changes at 90% blade height, and the tail shedding vortex emerges from the trailing edge of the blade and then is further squeezed by the shroud to cause the pressure on the surface of the blade to change.
Blade loading at 70% blade height on different impellers.
Blade loading at 90% blade height on different impellers.
Compared with the original blade, the three tail-edge optimization structures have smaller pressure fluctuation and pressure differences at the trailing edge structure. In particular, the pulsation amplitude of the tooth-shaped trailing edge is more stable and the optimization effect is better.
Static pressure fluctuation characteristics at shroud
The static pressure of different monitoring points in the shroud for different impellers is shown in Figure 20. The picture shows the change of pressure fluctuation of M01–M12 over time. Due to the periodic impact of vortex on the wall of the shroud, the static pressure changes of these 12 points are a common characteristic:
The static pressure at each point changes periodically with time. During one cycle of the impeller, the tail vortices of the three blades periodically impact each point, so the static pressure at each point has three periodic pressure fluctuations.
The static pressure at each point is reached when the vortex trail of the vane sweeps over the point, so the phase of the static pressure fluctuates with time at each point.
After the optimization of the trailing edge structure, the lower limit of fluctuation is effectively lifting, the pressure difference is reduced and the pressure fluctuation changes in the wake vortex.
When the structure is optimized, the pressure fluctuation amplitudes are smaller, especially the fluctuation amplitude of the pulse width of the tooth-shaped blade trailing edge scheme is more stable and the optimization effect is better. The optimization of the trailing edge structure of the internal flow field has changed in the wake vortices to make it produce, develop, and collapse more regularly.
Distribution of static pressure in the shroud at different impellers: (a) impeller 1, (b) impeller 2, (c) impeller 3, and (d) impeller 4.
Comparison spectrum of the broadband noise
The constant percentage bandwidth noise spectra of the structural optimization of the tail of the three blades are compared with the original blades. As shown in Figures 21–23, it is obvious that the changes in the sound pressure level at the corresponding frequency band are in the low-frequency range. After optimizing the trailing edge structure, the amplitude of low-frequency sound pressure level decreases obviously in the range of 50–125 Hz, while the 212 Hz frequency corresponds to the blade passage frequency. It also further illustrates that the optimized structure of the tail edge has an obvious effect on the low-frequency band of wideband noise. Impeller 4 of the tail edge in the sound pressure level shows a certain increase, especially in the blade passage frequency as shown in Figure 24. The pressure reaches to the maximum value when the vortex trail of the vane sweeps over this shroud. It is very important to reduce the noise by controlling the amplitude of the vortex pulsations at the trailing edge.
Comparison spectrum of impeller 1 and impeller 2.
Comparison spectrum of impeller 1 and impeller 3.
Comparison spectrum of impeller 1 and impeller 4.
Comparison of impeller 1 and impeller 4 at spectrum of sound pressure level.
Conclusion
In this article, the transient process of four kinds of impellers is studied by numerical simulation and experimental method. According to the results, the aerodynamic characteristics and internal flow patterns in the transient process are analyzed:
The trend of the experimental result is consistent with the simulated results below 700–850 r/min. The transient flow pattern is well indicated by the numerical simulation.
The area at the top of the shroud inlet is most strongly pressed. Vortex pulsation which caused by vortex shedding at the trailing edge of the blade is an important reason for vortex noise. The tip vortex and the wake shedding vortex are the most important factors that cause the vorticity of the axial-flow fan blades. In the circumferential direction, trailing edge of the blade is squeezed by the shroud and the vorticity is obviously increased.
Analysis of the blade load shows that the pressure fluctuation amplitude of the same blade height is similar in one blade rotation period. From the blade height position of 70%, due to the influence of the tip vortex and vortex shedding on the blade surface, the pressure difference and lift increase rapidly. Owing to the optimized structure of the blade trailing edge, the pressure fluctuation amplitudes are smaller at the trailing edge of the blade.
The static pressure of monitoring points on shroud shows, so the periodic impact of vortex on the wall of the shroud, three periodic changes with time. After the optimization of the trailing edge structure, the lower limit of fluctuation is effectively lifting, the pressure difference is reduced, and the pressure fluctuation changes in the wake vortex.
The amplitude reduction of the low-frequency sound pressure level in the range of 50–125 Hz is more obvious at the optimized trailing edge of the blade. It is demonstrated by the characteristics of the spectrum that the optimized structure of the tail edge has a significant impact on the wideband noise of the low frequency.
Footnotes
Handling Editor: Zengtao Chen
Declaration of conflicting interests
The author(s) declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article.
Funding
The author(s) disclosed receipt of the following financial support for the research, authorship, and/or publication of this article: This work was supported by the Natural Science Fund No. BK20160539, China.
