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SN Applied Sciences

, 2:82 | Cite as

Experimental and numerical analysis of vibrations in impeller of centrifugal blower

  • Kiran C. MoreEmail author
  • Sachin Dongre
  • Gaurav P. Deshmukh
Research Article
  • 214 Downloads
Part of the following topical collections:
  1. 3. Engineering (general)

Abstract

Centrifugal blowers are widely used in different industrial applications nuclear and petrochemical, which are proficient of as long as restrained to high pressure rise and flow rates. Design of such impellers is important with noise reduction and maximum efficiency. In present work, an experimental and numerical analysis is done to study the effect of impeller blade thickness and rotating speed on the performance of the impeller. Different blade thickness has been considered with different rotating speed. A modal analysis is done for the numerical study using commercial software ANSYS Workbench. The numerical results are validated with experimental results found in test. The six different mode shapes are found in numerical study. The natural frequency and the total deformation is calculated for different blade thickness and rotating speed. The results shows that, the blade with 1.5 mm thickness has a reduced noise and vibrations with maximum rotating speed.

Keywords

FEA Impeller blade Inconel alloy 740 Vibration Weight optimization 

1 Introduction

Impellers are widely used in many centrifugal components of the nuclear and petrochemical industries. These applications requiring the high speed flow, small in size, low noise and low cost. In particular, these impellers are used majorly air flow applications. Hence, there is need of reduction of the noise, while using these impeller in such applications. The fluid kinetic energy generated by the impeller creates the pressure in whole system which causes the noise and vibrations. Centrifugal blowers having mainly two main parts, namely, casing and impeller. The impeller is often considered an integral part of the suction motor since its housings and the motor are assembled as a unit. The design of such impellers is important with balancing for noise reduction. The major reasons of the noise in such impellers are (1) The large inlet gap between the inlet nozzle and impeller shroud, (2) improper balancing of the impeller, and (3) improper matching between the impeller outlet and volute tongue [1]. The designers of such impellers are encouraged by selection of the smaller, less noisy, and efficient. Several studies getting attentions towards the numerical and experimental studies for effect of noise and reduction of the noise level in such applications. Ramamurthi and Balasubramanian [2] did a steady state stress analysis of an impeller blade used in centrifugal fan. Cyclic symmetric structures were considered for the analysis. A finite Element analysis (FEA) tool is used for the analysis. Impeller stresses were measured using strain gauge technique and numerical results were compared with these results. Elyamin et al. [3] investigated the effect of number of impeller blade on the performance of the centrifugal pump numerically. Three different impeller with different number of impeller blades 5, 7 and 9 were considered for the analysis. Authors found that, the highest efficiency for the 7 number of blades and the overall loses deceases by increasing number of blades. Srivastava et al. [4] presented a numerical investigation of the mixed flow pump impeller blade with stress analysis. A design and stress analysis had been done to find out the effect of blades on the pump with different position in meridional annulus. The results of von misses stress were compared among the different blade position. The inclined blades at the inlet were found to be more suitable for the overall performance. Mane et al. [5] did a study on the M.S. impeller used in centrifugal pump using finite element analysis (FEA) method. A simple impeller model had been designed and analysed using the commercial software ANSYS. A simple modal analysis is done to measure the frequency (Hz) with respective rotating speed. From the results, It was found that the vector displacement of the impeller increased by increasing natural frequencies.

Ashiri et al. [6] did a dynamic modal analysis of impeller blade. A natural frequency and the mode shapes were studied for the different operating conditions. A single entry radial vane impeller had been considered for the modal analysis. Modelling and analysis were carried out using ANSYS Bladegen and ANSYS workbench. The results found that, there is a minimal effect of impeller blade thickness and impeller blade on the frequency. Mohonty and Rixen [7] developed a modified SSTD method for measurement of the harmonic excitements in a common modal analysis. Harmonic excitation can happened due to rotating components viz. unbalanced rotor and fluctuating forces etc. The purposed method which were considered a purely harmonic vibrations.

Mane et al. [8] did a FEA study on the design and analysis of the centrifugal pump impeller. A modal analysis had been done for the measurement of the natural frequency. A two different materials viz. Aluminum and Inconel 625 of the impeller were considered for the modal analysis. The results shows that, a maximum deformation were found in aluminum material under similar conditions while there was minimal effect on the natural frequency on these material change. Zhang et al. [9] did a numerical study to find out the effect of weld heat treatment on the residual stress in welded impeller. Based on results they were concluded that, circumferential residual stress in the impeller was found higher than the vertical residual stress. Lu et al. [10] presented numerical optimization on the vibroacoustics of a centrifugal volute impeller. An optimization was carried out using local thickness variation and experiments are conducted and results are compared with numerical results, which were found good agreement. A natural frequency and total deformation was extracted for the optimization of the volute fan. The results are used for the designing of such type of impeller. Fehse and Niese studied a generation mechanism of low frequency noise of a centrifugal fan. Different experiments had been performed with five centrifugal fan impeller and they found reduced flow separation, reduced noise, more uniform flow with less turbulence [11, 12].

Liu et al. [13] did a FEA study on analysis of the impeller life by considering centrifugal load and aerodynamic load. This study was established as a reliable by comparing experimental results with FEA results. Jayapragasan et al. [14] has selected Parameters for optimization is—fan outer diameter, number of blades and fan blade angle. Taguchi’s orthogonal array method helps to find out the optimum number of cases and the modelling has been carried out using SOLIDWORKS. ICEM CFD is used for meshing the blowers and analysing using FLUENT. The results are showing that the optimum combinations are 190 mm outer diameter, 80° blade angle and 8 numbers of blades.

Parshi et al. [15] did the study to increase the life period of centrifugal blower impeller by considering different types of materials, designs and by varying the thickness of base plate of centrifugal blower impeller. Using ANSYS software implemented static analysis under different boundary conditions also implemented Taguchi optimization technique to grab the best material, design and thickness. They found optimum total deformation for 12 number of blades and 8 mm blade thickness.

Present work describes the analysis of the impeller blade with reducing the blade thickness as a part of weight optimization. An experimental and numerical study has been performed to study the effect of the blade thickness and the speed of the impeller on the vibrations. A MS made impeller of 2 mm blade thickness is manufactured and tested experimentally. Different thickness of the blade viz. t = 1.5, 2, and 3 mm are considered to analyse the natural frequency and total deformation. A numerical study is done for the analysis of the different blade thickness and different spindle speed. The numerical results are validated with the experimental study results (Fig. 1).
Fig. 1

Centrifugal blower with impeller

2 Mathematical model

The first step in the impeller design is to select relative speed as per the requirement of head and flow rate conditions. This establishes the specific speed or type of the impeller. Selection of the speed is governed by a number of considerations: (1) Type of driver contemplated for the unit. (2) Higher specific speed results in a smaller blower and cheaper drivers. (3) Optimum hydraulic and total efficiency possible with each type varies with the specific speed.

2.1 Selection of impeller type as per specific speed

In the present design the flow rate and static pressure is taken as an input data based on the industrial requirement of machine [16, 17]. The selected input data are:
$$\begin{aligned} & {\text{Discharge}}\,\left( {\text{Q}} \right) = 0.5\,{\text{m}}^{3} / {\text{s}} \\ & {\text{Static}}\,{\text{Pressure}}\,\left( {\text{PS}} \right) = 981.25\,{\text{Pascal}} \\ & {\text{Specific}}\,{\text{Speed}}\,{\text{for}}\,{\text{N}} = 2850\,{\text{RPM}}\,{\text{for}}\,{\text{available}}\,{\text{motor}} \\ \end{aligned}$$
$$N_{\text{S}} = \frac{{{\text{N}} \times \sqrt {\text{Q}} }}{{\left( {\text{PS}} \right)^{3/4} }}$$
(1)
$$\begin{aligned} & {{N}}_{\text{s}} = \frac{{2850 \times \sqrt {0.5} }}{{\left( {981.25} \right)^{3/4} }} \\ & {\text{N}}_{\text{s}} = 1 1. 4 9 5\,\left( {{\text{RPM}},{\text{m}}^{3} / {\text{s}},\,{\text{Pascal}}} \right) \\ & {\text{N}}_{\text{s}} = 33,181\,\left( {{\text{RPM}},{\text{cfm}},{\text{inch}}\,{\text{of}}\,{\text{WC}}} \right) \\ \end{aligned}$$
where Ns = specific speed, N = impeller speed in RPM, Q = volume flow rate, in m3/s or cfm, PS = static pressure, in Pascal or inch of WC.

2.2 Design steps

Following are steps for the design of impeller of blower. This design steps are to be required to satisfy all the limitations and get flow condition as per theoretical design (Fig. 2).
Fig. 2

Inlet and outlet velocity diagram for impeller

  1. (1)

    Minimum impeller inlet diameter (d1)

     
$${\text{d}}_{1} = \sqrt[5]{{{\text{Q}}/{\text{N}}}}$$
(2)
$$\begin{aligned} & {\text{d}}_{1} = \left( {0.5/2850} \right)1/5 \\ & {\text{d}}_{1} = 0.177\,{\text{m}} \\ \end{aligned}$$
  1. (2)

    Impeller outside diameter (d2)

     
From the optimum performance specification it is stated that the ratio of the internal diameter to the external diameter is to fall between 0.4 and 0.7 as stated in the ASME code.
$$\frac{{{\text{d}}_{1} }}{{{\text{d}}_{2} }} = 0.4\,{\text{to}}\, 0.7$$
$${\text{d}}_{1} /{\text{d}}_{2} = 0.5$$
(3)
$${\text{d}}_{2} = 0.177/0.5 = 0.354\,{\text{m}}$$
  1. (3)

    Shroud diameter (ds)

     
$${\text{d}}_{\text{s}} = 0.94 \times {\text{d}}_{1}$$
(4)
$${\text{d}}_{\text{s}} = 0.94 \times 0.177 = 0.166\,{\text{m}}$$
  1. (4)

    Area of Shroud (As)

     
$${\text{A}}_{\text{s}} = \frac{\uppi}{4} \times {\text{d}}_{\text{s}}^{2}$$
(5)
$$\begin{aligned} & {\text{A}}_{\text{s}} = \left( {\uppi/4} \right) \times \left( {0.166} \right)2 \\ & {\text{A}}_{\text{s}} = 0.0216\,{\text{m}}^{2} \\ \end{aligned}$$
  1. (5)

    Impeller inlet blade width (b1)

     
$${\text{b}}_{1} = 0.221{\text{d}}_{1}$$
(6)
$$\begin{aligned} & {\text{b}}_{1} = 0.221 \times 0.177 \\ & {\text{b}}_{1} = 0.039\,{\text{m}} \\ \end{aligned}$$
  1. (6)

    Impeller inlet area (A1)

     
$${\text{A}}_{1} =\uppi \times {\text{d}}_{1 } \times {\text{b}}_{1}$$
(7)
$$\begin{aligned} & {\text{A}}_{1} =\uppi \times 0.177 \times 0.039 \\ & {\text{A}}_{1} = 0.0216\,{\text{m}}^{2} \\ \end{aligned}$$
By considering discharge, static pressure and specific speed further calculation are summarized in the Table 1.
Table 1

Summary of calculated values

Sr. no.

Parameter

Value

Unit

1

Minimum impeller inlet diameter (d1)

0.177

M

2

Impeller outside diameter (d2)

0.354

M

3

Eye or shroud diameter (ds)

0.166

M

4

Area of shroud (As)

0.0216

m2

5

Impeller inlet blade width (b1)

0.039

M

6

Impeller inlet area (A1)

0.0216

m2

7

Blade peripheral velocity at inlet (U1)

26.413

m/s

8

Absolute velocity at impeller inlet (V1)

23.148

m/s

9

Blade angle at impeller inlet (β1)

41.225

Degree

10

Relative velocity at impeller inlet (W1)

35.125

m/s

11

Impeller outlet blade width (b2)

0.039

M

12

Impeller outlet area (A2)

0.0434

m2

13

Outlet blade velocity (U2)

52.826

m/s

14

Radial component of outlet velocity (Vr2)

11.520

m/s

15

Blade exit angle (β2)

45

Degree

16

Tangential component of outlet velocity (Vu2)

41.306

m/s

17

Absolute velocity at impeller outlet (V2)

42.882

m/s

18

Relative velocity at impeller outlet (W2)

16.292

m/s

19

Diameter ratio (€)

2

20

Number of blades (Z)

12

Nos.

21

Slip factor (µ)

0.852

22

Actual exit velocity peripheral component due to slip (Vu2′)

35.193

m/s

23

Actual absolute exit velocity (V2′)

37.030

m/s

24

Actual relative velocity (W2′)

21.068

m/s

25

Actual blade exit angle (β2′)

33.148

Degree

26

Actual air exit angle (α2′)

18.125

Degree

27

Air velocity at impeller eye (Veye)

23.148

m/s

28

Loss factor (K1)

0.5

29

Pressure loss at impeller entry

164

kpa

30

Air density (ρ)

1.23

kg/m3

31

Loss factor (Kii)

0.2

32

Pressure loss in impeller blade passages

24.325

kpa

33

Pressure/head coefficient (ψ)

0.665

34

Flow coefficient (ø)

0.218

2.3 Selection of parameters

Different parameters were selected to analyse the best parameters while designing the impeller blade by using Taguchi’s design of experiments (DOE). Three different stages and parameters to be selected for the numerical study. For analysing the impeller design we are considering three main parameter like thickness of impeller, RPM of the system and mode number. A L9 orthogonal array has been selected for the analysis. The selected parameters are as shown in Table 2 and L9 orthogonal array parameters are shown in Table 3.
Table 2

Parameters for experimentation

Parameter

Parameter1

Thickness

Parameter2

Mode no.

Parameter3

RPM

Stage (1)

1.5

1

2550

Stage (2)

2

2

2650

Stage (3)

3

3

2750

Table 3

Orthogonal array

Parameter1

Parameter2

Parameter3

1

1

1

1

2

2

1

3

3

2

1

2

2

2

3

2

3

1

3

1

3

3

2

1

3

3

2

2.4 Statistical analysis using ANOVA technique

The Statistical analysis software used to optimize the impeller on the basic of Thickness, Mode Shape, RPM and natural frequency with the help of techniques of Taguchi design and ANOVA. Figure 3 implies that main effect plot for means for different process parameters. From the figure, it shows that, the maximum S/N ratio was found for the lower thickness and 2650 RPM.
Fig. 3

Mean effect plot for optimization of the impeller blade

3 Experimental setup

An impeller has been made with the MS material with the specific dimensions [18, 19, 20]. The properties of the MS material are tabulated in Table 4. Figures 4 and 5 shows the impeller with blades and casing cover. The dimensions of the impeller are mentioned in Table 1. The availability of the Inconel material is difficult and the cost is also high, hence MS material is used for the analysis. The different manufacturing process was used for manufacturing the impeller.
Table 4

Impeller material properties

Properties

Inconel 740

MS (IS 2062)

Young modulus (MPa)

2.21 × 105

2.07 x 105

Density g/cm3

8.05

7.85

Poisson’s ratio

0.37

0.3

Fig. 4

Photograph of actual impeller used for the Analysis

Fig. 5

Photograph of casing of the blower

3.1 Impeller

Figure 4 shows photograph of manufactured impeller which is used in the proposed study.

3.2 Casing of the blower

The component contains the impeller is generally called as the casing [21, 22, 23]. A rotating instrument comprises of suction as well as a discharge penetration for the rotating instrument like pumps in main flow path. The pump or blowers normally has vent fittings and small drain to eliminate gases confined in the casing. The casing of the blower used is shown in Fig. 5. In addition, it also performs few other important functions viz. (1) Provides pressure containment, (2) incorporates the collector, (3) allows rotor installation and removal, (4) maintains the alignment of the pump and its rotor under the action of pressure, (5) supports the pump and reasonable piping loads.

3.3 Motor

A three phase Crompton Greaves make induction motor of maximum 3000 RPM is selected for the analysis. The motor is connected to the impeller though shaft. The shaft is connected to the impeller and transmits the rotations to direct impeller. Figure 6 shows the three phase motor.
Fig. 6

Photograph of three phase motor

3.4 SVAN 974 vibration analyzer

A vibration analyzer SVAN 974 is designed for vibration measurements of various machinery and various tools. The instrument is using the SV 80 accelerometer which is used for vibration measurement of pumps, motors and fans. A non-contact type infrared tachometer is used for the measurement of the RPM while performing analysis. The SVAN 974 provides the parallel vibration acceleration, velocity and displacement results along with frequency analysis and wave recording, all at the same time. Figure 7 shows the photographs of the SVAN 974 instrument. The Different features like Vibration analysis, Charge type accelerometer, RPM measurement, FFT analysis used in the present study.
Fig. 7

Photograph of vibration analyzer SVAN 974

3.5 Experimental procedure

An investigation of the performance of impeller blade with different blade thickness and the rotating speed is carried out using experimental analysis. A MS material is selected for the manufacturing impeller blade with different blade thickness. A whole setup of blower has been manufactured with motor for the analysis of the impeller. Figure 8 shows the photograph of experimental setup with instrument and Fig. 9 shows the photograph of 2 mm blade thickness impeller. Table 4 shows the material properties of the impeller material used for the experimental and numerical analysis. A different process and operating parameters used for the analysis are tabulated in Table 5. An FFT analysis is done for vibration analysis using SVANTEK make SVAN 974 and SV 80 instrument machine vibration analyzer. The SVAN 974 provides the parallel vibration acceleration, velocity and displacement results along with the frequency analysis wave recording. The frequency range of the instrument is 0.7 Hz to 22.6 kHz with a magnetic probe. A digital anemometer is used to the measurement of the rotating speed of the impeller. A three phase motor is connected with the impeller. Casing of MS material provides the flow of the air. Four magnetic sensors of SVAN 974 instrument is connected at different location of the blower casing to find out the vibrations at different locations as a mode shape. The blower runs with a maximum speed 2939 of the motor to figure out the maximum frequency and noise.
Fig. 8

Photograph of the test setup

Fig. 9

Photograph of the impeller

Table 5

Selected performance parameters for the analysis

Process parameters

Experimental conditions

1

2

3

4

Thickness of impeller blade (mm)

1.5

2

3

Rotating speed (RPM)

2550

2650

2750

2850

4 Numerical analysis

A numerical modelling of impeller has been done using FEA and can be divided into different steps viz. 3D modelling, meshing, analyses using different boundary conditions and post processing the results. A 3D model is built with commercial software CATIA and then it is imported into a FEA based model analysis in ANSYS. A numerical domain of the impeller considered for the vibration analysis is shown in Figs. 10, 11, 12 and 13. A total 12 number of blades is considered by changing thickness of the blade. Inconel 740 material is selected for manufacturing as well as numerical analysis of the impeller blade and the properties are listed in Table 4. The meshing is an important part in the analysis and it is done by keeping an optimal ratio between the accuracy of selected performance parameter and the overall analysis time. A simple static modal analysis is done by applying different boundary conditions of loads and moments. Structural loads can be used viz. nodal forces or pressure on different faces or edges of the model. The modal analysis set is created for obtaining natural frequency and total deformation of the impeller blade for different rotational speed. Three different rotational speed (Np = 2550, 2650 and 2750 rpm) are considered for the analysis.
Fig. 10

Geometry of the impeller with backward curved blades

Fig. 11

Backward curved impeller with shroud

Fig. 12

Meshing (No of nodes: 90145, No. of elements: 58870)

Fig. 13

Boundary conditions

In numerical analysis following Boundary Conditions are given,
  • Rotational velocity:

    1. 1.

      Velocity is given in Z axis,

       
    2. 2.

      Initial velocity is given as 100 rpm

       
    3. 3.

      Final velocity is given 2850 rpm

       
  • Remote displacement:

    1. 1.

      X, Y component—free

       
    2. 2.

      Z component—0 mm

       
    3. 3.

      X, Y rotation—free

       
    4. 4.

      Z rotation—0°

       

Figures 10, 11, 12 and 13 are shows the numerical domain of the impeller.

5 Results and discussion

5.1 Validation of the numerical model

Experimental test is performed for the analysis of vibrations and noise at different location of the impeller. A three phase motor is rotating at maximum speed and the magnetic sensors are connected on three different location viz. motor base, impeller front cover and impeller cover top to achieve the natural frequency. The Numerical modal analysis is performed with same rotating speed and material of the impeller. The result found from the experimental test is shown in the Fig. 14, 15 and 16. These figures contain FFT analysis report of acceleration versus frequency at different locations on the Impeller. Computed results compared with the experimental results and found good agreement. The frequency found at motor base, impeller front cover and the impeller cover top are 1080.4 Hz, 347.6 Hz and 1079.6 Hz, respectively. The numerical results can be extracted with six different mode shapes and the results are shown in Fig. 17. The natural frequency found at mode shape 6 is 341.41 Hz at the front of the impeller.
Fig. 14

Experimental frequency at the location of blower motor base

Fig. 15

Experimental frequency at the location at impeller cover top

Fig. 16

Experimental frequency at the location of impeller cover front

Fig. 17

Contours of total deformation and the natural frequency at different location of the impeller

Four different rotating speeds are selected for the analysis, viz. 2550, 2650, 2750 and 2850 rpm and blade thickness is considered to be 1.5 mm. The vibrations are created with the different speed on the different location of the impeller body. Six different mode shapes are drawn for the analysis of the results for different rotating speed. Figure 18 shows the contours for mode shapes for 1.5 mm thickness at different rotating speed. It can be seen form the figure that, the total deformation and the frequency of the impeller blade is different for different mode shapes. The maximum frequency and the total deformation is found for the mode shape six. Rainbow contours indicates that the total deformations presents in the impeller blade.
Fig. 18

Results for 1.5 mm blade thickness with different rotating speed of 2550 to 2850 RPM

The analysis has been performed to study the effect of impeller blade thickness on the natural frequency and total deformation of the impeller blade. Three different impeller blade thicknesses have been chosen viz. 1.5 mm, 2 mm and 3 mm for the weight optimization of the impeller. A natural frequency and total deformation has been calculated for three different blade thicknesses. Six different mode shapes have been plotted for each case of impeller blade. The result shows that, as the increase in thickness of the blade the natural frequency increased, while the total deformation reduced. The maximum natural frequency found 364.05 Hz, while minimum total deformation of 16.28 mm for 3 mm blade thickness and 2850 rpm.

Figure 19 depicts the similar results for the 2 mm blade thickness at different rotating speed. It can be found that, the minimum natural frequency can be found for the 2650 rpm speed. Figure 20 indicates that, there no significant effects with the 3 mm impeller blade thickness at all rotating speed
Fig. 19

Results for 2 mm blade thickness with different rotating speed 2550 to 2850 RPM

Fig. 20

Results for 3 mm blade thickness with different rotating speed of 2550 to 2850 RPM

6 Conclusion

In this work, the impeller of M.S. and INCONEL 740 material are analyzed by the finite element method and experimental study by varying thickness of the blade. A total deformation and frequency of the impeller was obtained for different parameters such as rotating speed and thickness of the impeller blade. FFT test has been performed in a lab scale setup for analyzing the vibrations during rotating speed of the impeller and results were validated with numerical analysis performed in commercial software ANSYS version 14.5. The studied cases show the effect of thickness of the impeller blade and rotating speed on the vibration of the impeller. The following conclusions had been drown,
  1. 1.

    The vibrations are presents in impeller while rotating high speed, found maximum at the motor base of the impeller than the cover front and top of the impeller casing.

     
  2. 2.

    Thickness of the blade has a greate influence on the impeller performance. Vibrations can be controlled with selection of optimum blade thickness. Thickness of impeller blade was reduced from 3 mm to 1.5 mm with changing the rpm of rotation 2550, 2650, 2750.

     
  3. 3.

    The total deformation has reduced up to a certain speed and thickness after that the defoemation and frequency has been increased.

     
  4. 4.

    The maximum deformation of Inconel 740 impeller is 11.522 mm for 2750 RPM, 1.5 mm blade thickness with natural frequency of 8.27E−04 Hz.

     

Notes

References

  1. 1.
    Atre PC, Thundil KR (2012) Numerical design and parametric optimization of centrifugal fans with aerofoil blade impellers. Res J Recent Sci 1(10):7–11Google Scholar
  2. 2.
    Ramamurti V, Balasubramanian P (1987) Steady state stress analysis of centrifugal fan impellers. Comput Struct 25(1):129–135CrossRefGoogle Scholar
  3. 3.
    Elyamin GRA, Bassily MA, Khalil KY, Gomaa MS (2019) Effect of impeller blades number on the performance of a centrifugal pump. Alex Eng J 58:39–48CrossRefGoogle Scholar
  4. 4.
    Srivastava S, Roy AK, Kumar K (2014) Design of a mixed flow pump impeller blade and its validation using stress analysis. Procedia Mater Sci 6(2014):417–424CrossRefGoogle Scholar
  5. 5.
    Mane PR, Firake PL, Firake VL (2017) Finite element analysis of M.S. impeller of centrifugal pump. Int J Innov Eng Sci 2(9):1–4Google Scholar
  6. 6.
    Ashri M, Karuppanan S, Patil S, Ibrahim I (2014) Modal analysis of a centrifugal pump impeller using finite element method. In: MATEC web of conferences, vol 13, p 04030CrossRefGoogle Scholar
  7. 7.
    Mohanty P, Rixen DJ (2004) Modified SSTD method to account for harmonic excitations during operational modal analysis. Mech Mach Theory 39(12):1247–1255CrossRefGoogle Scholar
  8. 8.
    Mane P, Firake PL, Firake VL (2016) Design and analysis of centrifugal pump impeller by using FEA. Int J Eng Technol Sci Res 4(9):1368–1374Google Scholar
  9. 9.
    Zhang Z, Ge P, Zhao GZ (2017) Numerical studies of post weld heat treatment on residual stresses in welded impeller. Int J Press Vessels Pip 153:1–14CrossRefGoogle Scholar
  10. 10.
    Lu FA, Qi DT, Wang XJ, Zhou Z, Zhou HH (2012) A numerical optimization on the vibroacoustics of a centrifugal fan volute. J Sound Vib 331:2365–2385CrossRefGoogle Scholar
  11. 11.
    Fehse KR, Neise W (1999) Generation mechanisms of low-frequency centrifugal fan noise. AIAA J 37(10):1173–1179CrossRefGoogle Scholar
  12. 12.
    Fehse KR, Neise W (1998) Generation mechanisms of low-frequency centrifugal fan noise. In: 4th AIAA/CEAS aeroacoustics conference, p 2370Google Scholar
  13. 13.
    Liu S, Liu C, Hu Y, Gao S, Wang Y, Zhang H (2015) Fatigue life assesment of centrifugal compressor impeller based on FEA. Eng Fail Anal 60:383–390CrossRefGoogle Scholar
  14. 14.
    Jayapragasan CN, Janardhan Reddy K (2017) Design optimization and experimental study on blower for fluffs collection system. J Eng Sci Technol 12(5):1318–1336Google Scholar
  15. 15.
    Parshi B, Kumar A (2017) Design and analysis of centrifugal blower using steels and aluminium alloy. Tech Res Organ India 4(11):93–98Google Scholar
  16. 16.
    Wang P, Wang W, Li J (2017) Research on fatigue damage of compressor blade steel KMN-I using nonlinear ultrasonic testing. Shock Vib 2017:4568460Google Scholar
  17. 17.
    Karanjkar MU, More SH (2017) Vibration analysis and weight optimization of impeller for industrial air blower. Int J Adv Res Innov Ideas Educ 3(3):2945–2955Google Scholar
  18. 18.
    Kay M, Htay W (2014) Design and analysis of impeller for centrifugal blower using solid works. Int J Sci Eng Technol Res 03(10):2138–2142Google Scholar
  19. 19.
    Dinesh Tarel, Vaibhav Bhagat, Basavaraj Talikotti (2016) Static and dynamic analysis of impeller of centrifugal blower. Int J Innov Sci Eng Technol 3(5):547–553Google Scholar
  20. 20.
    Xiaozhang Qu, Liu Guiping, Duan Shuyong, Yang Jichu (2016) Multi-objective robust optimization method for the modified epoxy resin sheet molding compounds of the impeller. J Comput Des Eng 3:179–190Google Scholar
  21. 21.
    Oyelami AT, Olaniyan OO, Iliya D, Idowu AS (2008) The design of a closed-type-impeller blower for a 500 kg capacity rotary furnace. Assumpt Univ J Technol 12(1):50–56Google Scholar
  22. 22.
    Thangarasu VS, Sureshkannan G, Dhandapani NV (2015) Design and experimental investigation of forward curved, backward curved and radial blade impellers of centrifugal blower. Aust J Basic Appl Sci 9(1):71–75Google Scholar
  23. 23.

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© Springer Nature Switzerland AG 2019

Authors and Affiliations

  1. 1.Mechanical Engineering DepartmentD. Y. Patil Institute of Engineering and TechnologyAmbi, PuneIndia

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