The Influence of the Blade Outlet Angle on the Flow Field And

processes

Article The Inﬂuence of the Blade Outlet Angle on the Flow Field and Pressure Pulsation in a Centrifugal Fan

Hongchang Ding *, Tao Chang and Fanyun Lin College of Mechanical and Electronic Engineering, Shandong University of Science and Technology, Qingdao 266590, China; [email protected] (T.C.); [email protected] (F.L.) * Correspondence: [email protected]; Tel.: +86-178-5272-8669

Received: 29 September 2020; Accepted: 5 November 2020; Published: 8 November 2020

Abstract: This paper takes centrifugal fan as the research object and establishes five impeller models with different blade outlet angles. By means of computational fluid dynamics (CFD), the external characteristics of the centrifugal fan and the internal characteristics, including the velocity, pressure, and turbulent energy distribution, at the middle span plane of the impeller or fan were obtained and compared. In addition, the pressure fluctuations surrounding the impeller outlet were also analyzed. The results showed that the change of the blade outlet angle of the centrifugal fan had a great influence on the performance; the total pressure and efficiency of the fan were the highest when the outlet angle of the blade was increased to 29.5◦ under the design flow rate; and the influence of the outlet angle on the fan performance was different in off-design conditions. On the other hand, at different flow rates, the change of the internal flow field with the increase of the outlet angle was different. For the pressure fluctuation of the fan, by increasing the blade outlet angle properly under high flow conditions, the fluctuation amplitude of the fan at the blade frequency and its frequency multiplication could be reduced, which is conducive to decreasing the impeller noise. The research results have good guiding significance regarding the design of the pneumatic performance and noise reduction performance of centrifugal fans.

Keywords: centrifugal fan; blade outlet angle; aerodynamic performance; numerical simulation

1. Introduction Fans belong to the category of general machinery and are widely used in various industries of national economy. They are indispensable equipment for industrial and agricultural production. According to statistics, the power consumption of wind turbines accounts for 8–10% of the total power generation in China (Chen [1]). At present, centrifugal fans occupy a large proportion in China’s energy system. Therefore, it is of great significance to research and improve the pneumatic performance of fans for energy saving. However, in the process of optimizing the performance of fans, the traditional experimental methods have long cycles and high cost, and it is difficult to visually display the gas distribution inside a fan. Therefore, a CFD (Computational Fluid Dynamics) numerical simulation technology that can effectively reduce the sample size of experimental design and capture the details of flow inside the fan more specifically, has increasingly caught the interest of scholars. Many scholars have used numerical simulation technology to study centrifugal fans (Zhang et al. [2]; Zhou et al. [3]; Lin et al. [4]; Yu et al. [5]; Kishokanna et al. [6]). The impeller is the main moving part of a centrifugal fan, and the structural parameters of the impeller include the blade shape, blade profile, outlet width, number of blades, inlet and outlet diameter, etc. An excellent impeller design is helpful to improve the aerodynamic performance of the fan. At present, many scholars have studied the influence of certain impeller parameters on the aerodynamic performance of fans.

Processes 2020, 8, 1422; doi:10.3390/pr8111422 www.mdpi.com/journal/processes Processes 2020, 8, 1422 2 of 14

Li et al. [7] studied the influence of blade shapes on the performance of high specific speed centrifugal fans, and found that the blockage phenomenon at the blade outlet of impeller with plate blade and the turbulent kinetic energy inside the volute were weakened under the condition of a large flow rate, so that the performance of the fan with plate blade was better than that of the airfoil blade. Wu et al. [8] compared the performance of a centrifugal fan with different blade profiles, and found that the centrifugal fan with a double-arc blade was higher in the efficiency and total pressure under the design condition, but the axial power consumed by equal deceleration blade was smaller. However, under the condition that other parameters of the fan remain unchanged, the internal flow of the fan with an equal deceleration blade was more uniform under the condition of low flow. Jian et al. [9] found that when the blade outlet width changed, various losses of the fan increased, and the efficiency decreased. With the decrease of the blade outlet width, the flow-pressure curve of the fan shifted to the lower left and the pressure decreased with the increase of the flow rate. This provides a reference for the design of the outlet width of the impeller and the reconstruction of the impeller. Liu et al. [10] found that the aerodynamic performance of the fan can be improved by increasing the number of blades and the diameter of the blade outlet. The optimized fan with a 12-blade number and increased blade outlet diameter was better than the prototype fan in terms of the total pressure and efficiency. In addition, some scholars studied the interaction among several structural parameters of the impeller. For example, Esra et al. [11] used a neural network method to determine the optimal family of the impeller inlet and outlet radius and inlet and outlet angle to reduce the noise level in the early stage of fan design. Meng et al. [12] studied the influence of three blade structure parameters on the performance of a centrifugal fan, which were inlet angle, outlet angle, and blade number, and obtained that the best combination of the three blade structure parameters using the response surface model (RSM) optimization method. The maximum efficiency was 93.7%. Shi et al. [13] optimized the combination of the impeller and guide vane by studying the internal flow law of fans and combining this with numerical calculation results. They found that the streamline in the optimized guide vane was uniform and that the vorticity was reduced compared with the original guide vane. HEO et al. [14] analyzed the aerodynamic characteristics of a centrifugal fan with additional splitter blades in the impeller by using three-dimensional Reynolds-averaged Navier–Stokes (RANS). The global Pareto optimal frontier for centrifugal fan design was obtained by using a hybrid multi-objective evolutionary algorithm and response surface approximation model. As the main parameter of impeller, the blade outlet angle of centrifugal fans has also been studied by scholars. Wang et al. [15] optimized the fan design based on an orthogonal experimental design method and CFD numerical simulation and obtained the relative optimal combination model of the impeller geometric parameters and speed. Through the range analysis of the calculation results, they found that the blade outlet angle had the greatest impact on the fan efficiency. Recently, Swe et al. [16] discussed the flow characteristics of centrifugal fans with different blade outlet angles using the CFD numerical simulation method. The results showed that when the blade outlet angle was 25◦, the steady flow and unsteady flow were more uniform, and the flow distribution in the circumferential direction had little change. Yu et al. [17] also studied the effect of the blade outlet angle on the performance of a multi-blade centrifugal fan, and found that the wind pressure and efficiency of the fan increased in the flow range of 420–725 m3/h with the increase of the blade exit angle, and properly increasing the blade outlet angle can reduce the pulsation amplitude of the fan at the blade frequency and its multiplier frequency. Although many scholars have studied the blade outlet angle on the centrifugal fan, few have studied the effect of the blade outlet angle on the internal flow field and pressure fluctuation of the straight blade fan. In this paper, a type of straight plate blade fan was selected as the research model. Then, on the basis of the original blade outlet angle, five outlet angles (26.5◦, 28◦, 29.5◦, 31◦, and 32.5◦) were obtained, and the geometric model was established using SolidWorks software. ANSYS CFX software was used for the numerical simulation, and the feasibility of the calculation results was Processes 2020, 8, 1422 3 of 14 Processes 2020, 8, x FOR PEER REVIEW 3 of 16 verifiedangles on by the experiments. internal flow Through field and the performance software post-processing, of the centrifugal the influencefan was studied. of different The bladeamplitudes outlet anglesof fan blades on the internalwith different flow field outlet and angles performance were obtained of the centrifugal and analyzed fan wasusing studied. fast Fourier The amplitudes transform. ofThe fan research blades withresults di ffcanerent provide outlet certain angles reference were obtained values and for analyzedthe efficient, using safe, fast and Fourier stable transform. operation Theof a researchcentrifugal results fan. can provide certain reference values for the efficient, safe, and stable operation of a centrifugal fan. 2. Research Model and Simulation Method 2. Research Model and Simulation Method

2.1.2.1. Research Model ThisThis paperpaper tooktook aa certaincertain typetype ofof centrifugalcentrifugal fanfan asas thethe researchresearch model,model, andand thethe mainmain designdesign parametersparameters ofof thethe fanfan areare shownshown inin TableTable1 .1.

TableTable 1.1. MainMain designdesign parametersparameters ofof thethe centrifugalcentrifugal fan.fan.

ParametersParameters ValuesValues Rated air flow 6200 m³/h Rated air ﬂow 6200 m3/h Rated totalRated pressure total pressure 2700 Pa 2700 Pa Rated powerRated power 30 kW 30 kW Rated speedRated speed 1490 r/ 1490min r/min

TheThe strategystrategy ofof thisthis paperpaper waswas toto changechange thethe bladeblade outletoutlet angleangle ofof thethe originaloriginal fanfan model,model, andand itsits speciﬁcspecific impellerimpeller parametersparameters areare shownshown inin TableTable2 .2. The The overall overall geometric geometric model model was was established established in in SolidWorksSolidWorks asas shownshown inin FigureFigure1 .1.

Figure 1. Geometric model of the original fan.

Table 2. Specific parameters of model fan impeller. Figure 1. Geometric model of the original fan. Parameters/Marks Value Table 2. Specific parameters of model fan impeller. Impeller outer diameter D2 810 mm Number of blades Z 16 Parameters/Marks Value Impeller inner diameter D1 368 mm ImpellerBlade outer outlet diameter width b2 D2 43 mm 810 mm NumberBlade outlet of blades angle Zβ2 29.5◦ 16 Blade thickness δ 3 mm Impeller inner diameter D1 368 mm Blade outlet width b2 43 mm 2.2. Mesh and Check of Grid IndependenceBlade outlet angle β2 29.5° In this paper, SolidWorks wasBlade used forthickness the three-dimensional δ modeling3 mm of the centrifugal fan. In the process of modeling, the geometric model of the original fan was simplified, and some unimportant chamfers,2.2. Mesh and fillets, Check and of gaps Grid were Independence ignored. Based on this, the flow-path model of the centrifugal fan was established.In this paper, The flow-path SolidWorks model was was used divided for the into three-dimensional three parts: air inlet modeling flow-path, of the impeller centrifugal flow-path fan. In the process of modeling, the geometric model of the original fan was simplified, and some unimportant chamfers, fillets, and gaps were ignored. Based on this, the flow-path model of the

Processes 2020, 8, x FOR PEER REVIEW 4 of 16 Processes 2020, 8, 1422 4 of 14 centrifugal fan was established. The flow-path model was divided into three parts: air inlet flow- path, impeller flow-path and volute flow-path, and they were combined to form the centrifugal fan model.and volute To ensure flow-path, the real and working they were condition combined of toth forme air theflow, centrifugal the inlet fanand model.outlet flow-path To ensure themodel real wereworking extended condition properly. of the air flow, the inlet and outlet flow-path model were extended properly. AfterAfter completing completing the the establishment establishment of of the the fan fan model, model, the the mesh mesh generation generation of of the the model model began. began. DueDue to to the the simplicity simplicity of of the the axisymmetric axisymmetric model model of of the the fan fan inlet inlet section, section, the the hexahedron hexahedron structure structure waswas used used for for mesh mesh generation. generation. However, However, the the impeller impeller and and volute volute were were meshed meshed with with the the tetrahedron tetrahedron structurestructure due due to thethe complexitycomplexity of of their their flow flow passages, passages, and and the the meshing meshing of impeller of impeller and volute and volute is shown is shownin Figure in 2Figure. To make 2. To the make numerical the nu simulationmerical simulation as close toas theclose actual to the situation actual situation as possible, as allpossible, boundary all boundarylayer grids layer were grids refined were to refined meet the to requirementsmeet the requirements of the wall of function the wall method. function Themethod. minimum The minimumorthogonal orthogonal angle of the angle mesh of wasthe greatermesh was than greater 27◦, and than the 27°, maximum and the extensionmaximum ratio extension was less ratio than was 2.1, lesswhich than will 2.1, ensure which high will meshensure quality. high mesh quality.

(a) (b) (c)

FigureFigure 2. 2. MeshMesh generation generation of of the the impeller impeller and and volute. volute. (a ()a )Impeller. Impeller. (b (b) )Volute. Volute. (c ()c )Local Local refinements. reﬁnements.

TableTable 3 shows the grid grid independence independence verification verification results results of of the the centrifugal centrifugal fan fan at at design design points, points, inin which which the the total total pressure pressure and eefffificiencyciency were were obtained obtained under under five five di differentfferent numbers numbers of grids.of grids. When When the thenumber number of gridsof grids increased increased from from 3,284,561 3,284,561 to to 495,126, 495,126, the the total total pressure pressure and and effi efficiencyciency tended tended to to be bestable. stable. Based Based on on the the accuracy accuracy of the of simulationthe simulation results results and theand cost theof cost the of computing the computing time, the time, scheme the schemeof No. IIIof No. was III finally was adopted.finally adopted.

TableTable 3. 3. SchemeScheme for for grid grid independence independence check. check.

No. No. NumberNumber of Grids of Grids Total TotalPressure Pressure (Pa) (Pa)Efficiency Efficiency I I1,325,601 1,325,601 2820.9 2820.9 0.849 0.849 II II2,578,452 2,578,452 2686.6 2686.6 0.838 0.838 III III3,284,561 3,284,561 2608.4 2608.4 0.831 0.831 IV 4,002,563 2606.5 0.828 IV V4,002,563 4,951,236 2606.5 2606.9 0.828 0.830 V 4,951,236 2606.9 0.830 2.3. Simulation Settings 2.3. Simulation Settings At present, the standard k-ε turbulence model is the most widely used turbulence model, which At present, the standard k-ε turbulence model is the most widely used turbulence model, which has good stability and fast calculation speed compared with the zero equation model and the single has good stability and fast calculation speed compared with the zero equation model and the single equation model, and it is very suitable as the research object of this paper. equation model, and it is very suitable as the research object of this paper. The numerical simulation was carried out in ANSYS CFX14.5, and the simulation process was The numerical simulation was carried out in ANSYS CFX14.5, and the simulation process was based on the following settings: based on the following settings: • The inlet and volute were set in the static domain, and the impeller was set in the rotating domain. • The inlet and volute were set in the static domain, and the impeller was set in the rotating The default fluid was 25 C air, the reference pressure was 1 atm, and the flow of air was steady. • domain. ◦ • The rotation axis was set to be the z axis, and the rotational speed was set as 1490 r/min. • The default fluid was 25 °C air, the reference pressure was 1 atm, and the flow of air was steady. • The rotation axis was set to be the z axis, and the rotational speed was set as 1490 r/min.

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• The inlet boundary condition was set to the normal speed, and the outlet boundary condition The inlet boundary condition was set to the normal speed, and the outlet boundary condition was • was set to the average static pressure. set to the average static pressure. • Interface models were set to general connection, and the frozen rotor method was used for Interface models were set to general connection, and the frozen rotor method was used for • interface connection. interface connection. • The SIMPLE (Semi-Implicit Method for Pressure Linked Equations) algorithm was used for The SIMPLE (Semi-Implicit Method for Pressure Linked Equations) algorithm was used for • pressure-velocity coupling. The continuity equation, momentum equation, and dissipation rate pressure-velocity coupling. The continuity equation, momentum equation, and dissipation rate equation were discretized using the second-order upwind method. The solution step was 1000, equation were discretized using the second-order upwind method. The solution step was 1000, and the convergence precision of all residuals was less than 10−5. and the convergence precision of all residuals was less than 10 5. • The steady simulation served as the initial condition for the unsteady− simulation. The time step The steady simulation served as the initial condition for the unsteady simulation. The time step • was set at 0.00011186 s, which corresponded to a rotating angle of 1° for each time step at a was set at 0.00011186 s, which corresponded to a rotating angle of 1 for each time step at a rotating speed of 1490 r/min. The time for one cycle was 0.04027 s,◦ and five rounds were rotating speed of 1490 r/min. The time for one cycle was 0.04027 s, and five rounds were simulated. simulated. The final round was selected for analysis due to the relatively more stable flow field. The final round was selected for analysis due to the relatively more stable flow field. To study the possible influence of different blade outlet angles on the pressure fluctuation characteristicsTo study of the the possible impeller influence outlet and of volute different of fan, blade monitoring outlet angles points onwere the set pressure on the outlet fluctuation of the impellercharacteristics with mid-section of the impeller (z =outlet 79 mm) and to volute monitor of fan,the monitoringunsteady flow. points The were layout set onof the monitoring outlet of the pointsimpeller is withshown mid-section in Figure (z3, =and79 mm)the impellers to monitor with the different unsteady blade flow. Theoutlet layout angles of thehad monitoring the same monitoringpoints is shown position. in Figure In Figure3, and the3(b), impellers P1 to P4 with are distributed di fferent blade clockwise outlet angles along hadthe thecircumference, same monitoring and P5position. and P6 Inare Figure arranged3b, P1 in tothe P4 volute are distributed tongue and clockwise volute outlet along section, the circumference, respectively. and P5 and P6 are arranged in the volute tongue and volute outlet section, respectively.

(a) (b)

Figure 3. Schematic: (a) mid-section (z = 79 mm); (b) location of the monitoring point. Figure 3. Schematic: (a) mid-section (z = 79 mm); (b) location of the monitoring point. 3. Experimental Veriﬁcation 3. Experimental Verification 3.1. Experimental Equipment 3.1. Experimental Equipment To verify the feasibility of the established three-dimensional model of a centrifugal fan in numerical simulation,To verify the the external feasibility characteristics of the established experiment three-dimensional of the fan model was model carried of a out. centrifugal The fan testfan wasin numericalconducted simulation, according tothe the external GB / T1236–2000 characteristics standard, experiment the air intakeof the testfan systemmodel was carried adopted, out. and The the fan test was conducted according to the GB / T1236–2000 standard, the air intake test system was

Processes 2020, 8, 1422 6 of 14

Processes 2020, 8, x FOR PEER REVIEW 6 of 16 testadopted, device wasand the type test C device (pipe was inlet type and C free(pipe outlet) inlet and speciﬁed free outlet) by specified the national by the standard. national standard. The schematic diagramThe schematic of the experimental diagram of the device experimental is shown device in Figure is shown4. in Figure 4.

Processes 2020, 8, x FOR PEER REVIEW 6 of 16

adopted, and the test device was type C (pipe inlet and free outlet) specified by the national standard. The schematic diagram of the experimental device is shown in Figure 4.

FigureFigure 4.4. Schematic diagram diagram of ofcentrifugal centrifugal fan fanexperimental experimental device. device.

TheThe centrifugal centrifugal fan fan testtest platform was was primarily primarily composed composed of the of fan the inlet fan inletpipe, pipe,main body main of body the of the centrifugalcentrifugal fan, fan, and and motor. motor. Among them, them, the the inlet inlet pipe pipe part part mainly mainly included included the current the current collector, collector, compensationcompensation micromanometer, micromanometer, rectifier rectiﬁer network, network, an andd U-tube U-tube liquid liquid pressure pressure gauge. gauge. Table 4 Table shows4 shows thethe instrument instrument used used in in the the testtest and its its specific speciﬁc parameters. parameters. The The actual actual drawing drawing of the of fan the is shown fan is shownin in FigureFigure5, and5, and Figure Figure1 shows 1 shows the the geometric geometric model of of Figure Figure 5.5 .

Table 4. Instruments for the experiment.

Test Parameter Name Instrument Name Model/Speciﬁcation Minimum Scale Accuracy

Atmospheric pressure Aneroid barometer YM3 100 Pa 200 Pa ≤ Atmospheric temperature hygrothermograph TA298 0.1 C 1 C Figure 4. Schematic diagram of centrifugal fan experimental device.◦ ± ◦ Atmospheric humidity hygrothermograph TA298 0.10% 5.0% ± Flowmeter diﬀerential Compensation YJB 2500 0.01 mm 8Pa Thepressure centrifugal/∆P fan test platformmicromanometer was primarily composed− of the fan inlet pipe, main body± of the centrifugal fan, and motor. AmongU-tube them, liquid the inlet pipe part mainly included the current collector, Gauge pressure/Pe3 1 mm 0.5% compensation micromanometer,pressure rectifier gauge network, and U-tube liquid pressure gauge. Table 4± shows the instrumentTest speed used in the test and its specific parameters. The actual drawing0.1 r/min of the fan is shown in Shaft powerTorque meter 100 N m 0.1 kW 0.1% Figure 5,Torque and Figure 1 shows the geometric model of Figure· 5. 0.1 N m ± ·

Figure 5. The actual drawing of the fan.

Table 4. Instruments for the experiment.

Test Parameter Minimum Instrument Name Model/Specification Accuracy Name Scale Atmospheric Aneroid barometer YM3 100 Pa ≤200 Pa pressure Atmospheric hygrothermograph TA298 0.1 °C ±1 °C temperature

FigureFigure 5. 5.The The actualactual drawing of of the the fan. fan.

Table 4. Instruments for the experiment.

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3.2. Experimental Results Figure6 shows the results comparison between the numerical simulation and experiment. The longitudinal coordinate is the total pressure coefficient P and total pressure efficiency η, and the calculation equation is expressed as: P P = , (1) ρu2 πD n u = 2 , (2) 60 QP η = , (3) Mω where P, ρ, u, n, Q, M and ω respectively represent the total pressure, Pa; air density, kg/m3; impeller outlet diameter, mm; rotational speed, r/min; experimental flow rate m3/s; torque, N m; angular · velocity, rad/s. From Figure6, the experiment results are close to the numerical results. Q /Qdes is the ratio of the experimental flow rate to the design flow rate. As the numerical simulation simplifies the model and neglects the loss, the simulation value was higher than the experimental value. The error between the tested and calculated total pressure was the smallest near the work point but gradually increased with the flow away from the design point, and the maximum deviation value did not exceed 8%. There were two main influencing factors: One was the complex internal flow rule of fans under a small flow rate; the turbulence model selected in this paper could not effectively simulate the flow separation in the flow channel due to time and condition limitations. The other factor was that the actual flow rate of the fan was transient, which is ignored by the frozen rotor method. Although there wereProcesses some 2020 errors, 8, x FOR between PEER REVIEW the experimental and the numerical results, the general changing8 of trend16 of the total pressure and efficiency curves was the same, and the reliability of the numerical simulation can be confirmed according to the experimental results.

Figure 6. Curve of the total pressure coefficient and efficiency of the centrifugal fan model. Figure 6. Curve of the total pressure coefficient and efficiency of the centrifugal fan model. 4. Results and Discussion 4. Results and Discussions 4.1. Contrast of External Characteristics 4.1.From Contrast Figure of External7, the totalCharacteristics pressure of the fan decreased with the increase of the flow rate. The efficiencyFrom Figure of the 7, fan the increased total pressure first andof the then fan decreased,decreased with and the it reached increase the of the highest flow valuerate. The near the ratedefficiency operating of the point. fan increased At Q/Qdes first< and1.4, then the totaldecrea pressuresed, and ofit reached the fan the increased highest value first and near then the rated decreased operating point. At Q/Qdes < 1.4, the total pressure of the fan increased first and then decreased with with the rise of the blade outlet angle at the same flow rate. When β2 = 29.5◦, the total pressure was the highest.the rise of In the addition, blade outlet the changeangle at trendthe same of eflowfficiency rate. When was di βff2 =erent 29.5°, from the total the totalpressure pressure was the change. highest. In addition, the change trend of efficiency was different from the total pressure change. The The outlet angle of the blade with the highest efficiency was different under different flow rates. At a outlet angle of the blade with the highest efficiency was different under different flow rates. At a high flow rate, the efficiency increased with the increase of the outlet angle. At the design or a low flow rate, the efficiency was the highest when the blade outlet angle was 26.5° and reached the maximum of 84.85% at the designed flow rate.

(a) (b)

Figure 7. External characteristic curve of fan with different blade outlet angles: (a) Comparison of the total pressure coefficient. (b) Efficiency comparison.

4.2. Velocity Vector Distribution To further describe the internal flow of the fan, the relative velocity distributions in the middle section of the flow passage were obtained, as shown in Figure 8. There were vortices of different degrees in the impeller channel between 90° and 270° (see Figure 3) at different flow rates, and the

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Figure 6. Curve of the total pressure coefficient and efficiency of the centrifugal fan model.

4. Results and Discussions

4.1. Contrast of External Characteristics From Figure 7, the total pressure of the fan decreased with the increase of the flow rate. The efficiency of the fan increased first and then decreased, and it reached the highest value near the rated operating point. At Q/Qdes < 1.4, the total pressure of the fan increased first and then decreased with the rise of the blade outlet angle at the same flow rate. When β2 = 29.5°, the total pressure was the Processeshighest.2020, 8 ,In 1422 addition, the change trend of efficiency was different from the total pressure change. The8 of 14 outlet angle of the blade with the highest efficiency was different under different flow rates. At a high flow rate, the efficiency increased with the increase of the outlet angle. At the design or a low flow high flow rate, the efficiency increased with the increase of the outlet angle. At the design or a low flow rate, the efficiency was the highest when the blade outlet angle was 26.5° and reached the maximum rate, theof 84.85% efficiency at the was designed the highest flow rate. when the blade outlet angle was 26.5◦ and reached the maximum of 84.85% at the designed flow rate.

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(a) (b) number of vortices gradually decreased with the increase of the flow rate. At a low flow rate, the FigureoutletFigure 7. angleExternal 7. Externalhad little characteristic characteristic effect on the curve curve relative of of fan fanvelocity with with di differentdistribution;fferent blade blade however, outlet outlet angles: angles: the velocity (a) (Comparisona) Comparison near the of volute the of the totaltongue pressuretotal increasedpressure coeffi coefficient. significantlycient. (b)E (bffi) Efficiencywithciency thecomparison. increase comparison. of the outlet angle.

4.2. Velocity4.2. VelocityAt Vector the Vector design Distribution Distribution flow rate, when the blade outlet angle was 31°, the number of impeller channels with a vortex was the least, and the low-speed area of the impeller outlet was the least. Under the Tocondition furtherTo further describeof a describelarge theflow the internal rate, internal there flow flowwere of of low-speed the the fan,fan, thethe regions relative of varying velocity velocity degrees distributions distributions near the in impellerthe in middle the middle sectionsectionoutlet of the ofbetween flowthe flow passage 90° passage and 270°, were were and obtained, obtained,the area of as as the shownshow low-speedn in Figure area decreased 8.8. There There werewith were thevortices vorticesincrease of ofdifferent of the di fferent degreesblade outletin the angle. impeller However, channel the be flowtween state 90° of and the 270°impeller (see channelFigure 3)near at differentthe volute flow tongue rates, became and the degrees in the impeller channel between 90◦ and 270◦ (see Figure3) at di fferent flow rates, and the numberincreasingly of vortices unstable. gradually When decreased β2 = 32.5°, with the low the velocity increase region of the and flow flow rate. separation At a low appeared flow rate, on the the outlet suction surface of the trailing edge of the blade, resulting in flow blockage. It may be that the outlet angle had little effect on the relative velocity distribution; however, the velocity near the volute tongue angle was too large to match the volute and that the fluid impact on the volute tongue caused this increasedphenomenon. significantly with the increase of the outlet angle.

β2 = 26.5° β2 = 28° β2 = 29.5° β2 = 31° β2 = 32.5°

4103.5 m³/h

6200 m³/h

8545.7 m³/h

FigureFigure 8. Velocity 8. Velocity vector vector distributions distributions at at the the middle middle span span plane plane of the of fan. the fan.

4.3. Pressure Distribution Figure 9 shows the static pressure distribution of the middle span plane of impellers with different blade outlet angles and different flow rates. Under the condition of a low flow rate, the low- pressure region of the impeller inlet increased first and then decreased with the increase of the outlet angle. At the design flow rate, the area of the low-pressure zone increased first and then decreased with the increase of the impeller outlet angle between 90° and 270°. Under a large flow rate, there

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At the design flow rate, when the blade outlet angle was 31◦, the number of impeller channels with a vortex was the least, and the low-speed area of the impeller outlet was the least. Under the condition of a large flow rate, there were low-speed regions of varying degrees near the impeller outlet between 90◦ and 270◦, and the area of the low-speed area decreased with the increase of the blade outlet angle. However, the flow state of the impeller channel near the volute tongue became increasingly unstable. When β2 = 32.5◦, the low velocity region and flow separation appeared on the suction surface of the trailing edge of the blade, resulting in flow blockage. It may be that the outlet angle was too large to match the volute and that the fluid impact on the volute tongue caused this phenomenon.

4.3. Pressure Distribution Figure9 shows the static pressure distribution of the middle span plane of impellers with di fferent blade outlet angles and different flow rates. Under the condition of a low flow rate, the low-pressure region of the impeller inlet increased first and then decreased with the increase of the outlet angle. Processes 2020, 8, x FOR PEER REVIEW 10 of 16 At the design flow rate, the area of the low-pressure zone increased first and then decreased with the increasewas a reverse of the pressure impeller gradient outlet angle in the between impeller 90◦ chandannel 270 ◦between. Under a90° large and flow 270°, rate, and there this phenomenon was a reverse pressurewas alleviated gradient with in the the increase impeller of channel the blade between outlet 90 angle.◦ and 270◦, and this phenomenon was alleviated with the increase of the blade outlet angle. The static pressure increased with the increase of thethe flowflow rate.rate. The maximum static pressure occurred at the impeller outlet near the volute tong tongue,ue, which increased with the increase of the blade outlet angle, and the area of highhigh pressurepressure areaarea becamebecame larger.larger. In the static pressure distribution diagram of the impeller, we observed that the static pressure value of the fan gradually increased due to the continuous workwork ofof thethe impellerimpeller bladeblade rotation rotation on on the the air air flow. flow. At At the the impeller impeller inlet, inlet, there there was was a cleara clear low low pressure pressure area area on on the the suction suction surface surface of ofthe the blade,blade, andand reversereverse flowflow occurred near the wall, wall, resulting in separation loss. At high flow flow rates, this phenomenon was alleviated with the increase of the outlet angle.

β2 = 26.5° β2 = 28° β2 = 29.5° β2 = 31° β2 = 32.5°

4103.5 m³/h

6200 m³/h

8545.7 m³/h

Figure 9. StaticStatic pressure pressure distributions at the middle span plane of the impeller.

As can be seen from Figure 10,10, the total pressure distribution distribution of the fan fan gradually gradually increased increased from the impeller inlet to the outlet. As As the the flow ﬂow rate rate in increased,creased, the the total total pressure pressure increased. increased. Under Under different diﬀerent outlet angles, the pressure changes in the impellerimpeller passage mainly occurred between 90° 90◦ andand 270°. 270◦. At a low flow rate, the total pressure did not change greatly with the increase of the blade outlet angle. At the design flow rate, the low-pressure region appeared at the outlet of the single impeller, which increased first and then decreased with the increase of the outlet angle.

Combined with the velocity vector diagram, the low-pressure area appeared in the low speed region. At high flow rates, the total pressure increased with the increasing of the outlet angle. With the increase of the blade outlet angle, the effective working area of the impeller increased, the total impeller pressure increased overall, especially in the circumferential area of the outlet of the blade. In addition, the pressure distribution of different flow paths between blades was different in the range of 90°−270°, which indicates that the flow characteristics inside the fan were asymmetric.

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At a low flow rate, the total pressure did not change greatly with the increase of the blade outlet angle. At the design flow rate, the low-pressure region appeared at the outlet of the single impeller, which increased first and then decreased with the increase of the outlet angle. Combined with the velocity vector diagram, the low-pressure area appeared in the low speed region. At high flow rates, the total pressure increased with the increasing of the outlet angle. With the increase of the blade outlet angle, the effective working area of the impeller increased, the total impeller pressure increased overall, especially in the circumferential area of the outlet of the blade. In addition, the pressure distribution of different flow paths between blades was different in the range of 90 270 , ◦− ◦ which indicates that the flow characteristics inside the fan were asymmetric. Processes 2020, 8, x FOR PEER REVIEW 11 of 16

β2 = 26.5° β2 = 28° β2 = 29.5° β2 = 31° β2 = 32.5°

4103.5 m³/h

6200 m³/h

8545.7 m³/h

Figure 10. TotalTotal pressure distributions at the middle span plane of the impeller.impeller.

4.4. Turbulence Kinetic Energy Distribution 4.4. Turbulence Kinetic Energy Distribution It can be seen from Figure 11 that the high turbulent kinetic energy region was mainly distributed It can be seen from Figure 11 that the high turbulent kinetic energy region was mainly near the volute tongue at the impeller outlet and gradually diffused toward the surrounding area. distributed near the volute tongue at the impeller outlet and gradually diffused toward the When the gas flowed through this area, boundary layer separation easily occurred. In general, surrounding area. When the gas flowed through this area, boundary layer separation easily occurred. the turbulent energy was the smallest when the blade outlet angle was 29.5 . With the increase of In general, the turbulent energy was the smallest when the blade outlet angle◦ was 29.5°. With the the blade outlet angle, the turbulent energy of the volute section first decreased and then increased. increase of the blade outlet angle, the turbulent energy of the volute section first decreased and then The outlet angle of the blade had a suitable value, which can reduce the flow loss of the volute section increased. The outlet angle of the blade had a suitable value, which can reduce the flow loss of the and improve the efficiency of the fan. volute section and improve the efficiency of the fan.

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Processes 2020, 8, x FOR PEER REVIEW 12 of 16

β2 = 26.5° β2 = 28° β2 = 29.5° β2 = 31° β2 = 32.5°

4103.5 m³/h

6200 m³/h

8545.7 m³/h

FigureFigure 11. Turbulence11. Turbulence kinetic kinetic energy energy distributions distributions at the the middle middle span span plane plane of a of fan. a fan.

4.5. Pressure4.5. Pressure Pulsation Pulsation Analysis Analysis To betterTo better analyze analyze the the influence influence of of the the blade blade outlet outlet angle on on the the pressure pressure pulsation pulsation in the in impeller the impeller outletoutlet area, area, the time-domainthe time-domain characteristics characteristics ofof pressurepressure pulsation at at each each monitoring monitoring point point of the of the impellerimpeller were were transformed transformed by by fast fast Fourier Fourier transform, transform, andand the the frequency frequency domain domain of pressure of pressure pulsation pulsation was obtained,was obtained, as shown as shown in Figurein Figure 12 12.. It It can can bebe seenseen from the the diagram diagram that that the the main main frequency frequency of of pulsation at the monitoring point of the impeller outlet of the centrifugal fan was the blade passing pulsation at the monitoring point of the impeller outlet of the centrifugal fan was the blade passing frequency (fBPF) and its frequency multiplication, and the amplitude of pulsation reached the frequency (f ) and its frequency multiplication, and the amplitude of pulsation reached the maximum maximumBPF at the blade passing frequency. Due to rotor–stator interactions, pressure fluctuations near at thethe blade volute passing tongue frequency. (P1, P5) were Due large. to rotor–stator However, from interactions, the view of pressure the frequency fluctuations domain near map theof the volute tonguemonitoring (P1, P5) werepoints, large. the change However, rules of from each the monitori viewng of thepoint frequency were different domain at different map of outlet the monitoringangles, points,sothe further change analysis rules is ofrequired. each monitoring The blade passing point werefrequency different was calculated at different by outletEquation angles, (4). so further analysis is required. The blade passing frequency was calculated by Equation (4). = NZ fBPF , (4) NZ60 fBPF = , (4) Where N is the speed, r/min; Z is the number of blades.60 where N isFor the further speed, analysis, r/min; Z we is theuse numbermean pressure of blades. amplitude cp to calculate pressure pulsation Forenergy further of all analysis, the six measuring we use mean points pressure for different amplitude outlet angles.cp to calculate pressure pulsation energy of all the six measuring points for different outlet angles.

Pn 1 cp i c = − (n = 6), (5) p 6 where cp i represents the pressure amplitudes at fBEF for diﬀerent measuring positions. − Processes 2020, 8, x FOR PEER REVIEW 13 of 16

n c ==1 pi− cnp ()6 (5) 6 , where c pi− represents the pressure amplitudes at fBEF for different measuring positions. Figure 13 shows the average pressure amplitude at different outlet angles. In general, with the increase of the blade outlet angle, the pressure fluctuation amplitude of the fan increased. However, the amplitude of pressure fluctuation decreased with the increase of the blade outlet angle at 26.5°−28° and 29.5°−31°. Therefore, overall, the outlet angle of the blade increasing in a certain range reduced the pressure pulsation on the impeller, which is beneficial to reduce the noise of the impeller. Processes 2020, 8, 1422 12 of 14

(b) (a)

(e)

Figure 12. Frequency domain of the pressure pulsation for diﬀerent blade outlet angles: (a) 26.5 ; ◦ (b) 28◦;(c) 29.5◦;(d) 31◦; and (e) 32.5◦.

Figure 13 shows the average pressure amplitude at different outlet angles. In general, with the increase of the blade outlet angle, the pressure fluctuation amplitude of the fan increased. However, the amplitude of pressure fluctuation decreased with the increase of the blade outlet angle at 26.5 28 ◦− ◦ and 29.5 31 . Therefore, overall, the outlet angle of the blade increasing in a certain range reduced ◦− ◦ the pressure pulsation on the impeller, which is beneficial to reduce the noise of the impeller. Processes 2020, 8, x FOR PEER REVIEW 14 of 16

Figure 12. Frequency domain of the pressure pulsation for different blade outlet angles: (a) 26.5°; (b) 28°; (c) 29.5°; (d) 31°; and (e) 32.5°.

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Figure 13. Average pressure amplitude at different angles. Figure 13. Average pressure amplitude at different angles. 5. Conclusions 5. ConclusionsIn this paper, five models of centrifugal fan impellers with different blade outlet angles were established,In this and paper, the influencefive models of theof centrifugal blade outlet fan angle impellers on centrifugal with different fan performance blade outlet was angles studied were usingestablished, CFD software. and the The influence conclusions of the can blade be drawnoutlet angle as follows: on centrifugal fan performance was studied using CFD software. The conclusions can be drawn as follows: 1. The blade outlet angle had different effects on the total pressure and efficiency. At Q/Qdes < 1.4, 1. theThe total blade pressure outlet coeangleffi cienthad different of the fan effects first increasedon the total and pressure then decreased and efficiency. withthe At riseQ/Qdes of < the 1.4, bladethe total outlet pressure angle at coefficient the same flow of the rate. fan The first outlet increased angle and of the then blade decreased with the with highest thee risefficiency of the wasblade diff outleterent underangle at di thefferent same flow flow rates. rate. AtThe a outl highet flowangle rate, of the the blade efficiency with the increased highest with efficiency the increasewas different of the outletunder angle. different At the flow design rates. or At a lowa hi flowgh flow rate, rate, the ethefficiency efficiency was theincreased highest with when the theincrease blade outletof the angleoutlet wasangle. 26.5 At◦ theand design reached or thea low maximum flow rate, of the 84.85% efficiency at the was designed the highest flow rate. when 2. Throughthe blade the outlet simulation angle was and 26.5° analysis and ofreached the flow the field maximum inside aof centrifugal 84.85% at fan,the designed with the increaseflow rate. 2. ofThrough the outlet the angle, simulation the flow and speed analysis near of the the volute flow field tongue inside increased a centrifugal and the fan, low with pressure the increase area atof the the impeller outlet angle, inlet the first flow increased speed andnear thenthe volu decreasedte tongue at aincreased low flow and rate. the With low apressure blade outlet area at anglethe impeller of 29.5◦ atinlet the first design increased flow rate, and the then swirl decreased in the blade at a low passage flow wasrate. minimal. With a blade In the outlet case ofangle a largeof 29.5° flow at rate, the thedesign increase flow ofrate, the the blade swirl outlet in the angle blade decreased passage thewas low minimal. speed areaIn the at case the impeller of a large outlet,flow rate, decreased the increase the reverse of the pressure blade gradientoutlet angle area decreased at the blade the leading low speed edge, area and at increased the impeller the totaloutlet, pressure. decreased The the turbulent reverse kinetic pressure energy gradient decreased area at first the andblade then leading increased edge, with and theincreased increase the oftotal the bladepressure. outlet The angle. turbulent kinetic energy decreased first and then increased with the increase 3. Throughof the blade the analysis outlet angle. of the frequency domain diagram, we found that the pressure fluctuation 3. ofThrough a centrifugal the analysis fan was of the the smallest frequency when domain the outlet diagram, angle we of found the blade that wasthe pressure 28, and wefluctuation found thatof a properly centrifugal increasing fan was the the outlet smallest angle when of the the blade outlet reduced angle of the the fluctuation blade was amplitude 28, and we of found the fanthat at properly the blade increasing frequency the and outlet its frequency angle of th multiplication,e blade reduced which the fluctuation is conducive amplitude to reducing of the impellerfan at the noise. blade frequency and its frequency multiplication, which is conducive to reducing impeller noise. Author Contributions: Conceptualization, H.D.; Data curation, T.C.; Methodology, H.D. and T.C.; Software, F.L.; Supervision,Author Contributions: H.D.; Validation, Conceptualization, F.L.; Writing–original H.D.; Data draft, curation, T.C.; Writing–review T.C.; Methodology, and editing, H.D. and H.D. T.C.; All authorsSoftware, haveF.L.; read Supervision, and agreed H.D.; to the Validation, published F.L.; version Writing–orig of the manuscript.inal draft, T.C.; Writing–review and editing, H.D. Funding:Funding:This This research research has has been been supported supported by by key key research research and and development development project projec oft of Shandong Shandong province province (2017GGX203005),(2017GGX203005), Natural Natural Science Science Foundation Foundation Project Project of of Shandong Shandong province province (ZR2019MEE068) (ZR2019MEE068) and and Design Design and and performance optimization of magnetic levitation high speed centrifugal pump(201810424020). The supports are gratefully acknowledged. Conflicts of Interest: The authors declare no conflict of interest. Processes 2020, 8, 1422 14 of 14

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