experiment and simulation analysis on noise attenuation of

Research ArticleExperiment and Simulation Analysis on Noise Attenuation ofAlMF Cylindrical Shells

Bin Li1 Jian Li2 Shilin Yan1 Wenjie Yan13 and Xu He1

1Hubei Key Laboratory of Theory and Application of Advanced Materials Mechanics Wuhan University of TechnologyWuhan Hubei 430070 China2Guangxi Key Laboratory of Automobile Components and Vehicle Technology Liuzhou Guangxi 545006 China3Department of Mechanical Engineering Henan Mechanical and Electrical Vocational College Xinzheng Henan 451191 China

Correspondence should be addressed to Shilin Yan yanshl408126com

Received 15 June 2017 Accepted 28 September 2017 Published 26 October 2017

Academic Editor Tai Thai

Copyright copy 2017 Bin Li et al This is an open access article distributed under the Creative Commons Attribution License whichpermits unrestricted use distribution and reproduction in any medium provided the original work is properly cited

For the issue concerning internal noise reduction of Al-made cylindrical shell structure the noise control method of layingmelamine foam (MF) layer is adopted for in-shell noise attenuation experiments of Al and AlMF cylindrical shells andcorresponding internal noise response spectrograms are obtained Based on the VirtualLab acoustics software a finite elementmodel is established for the analysis of noise in the AlMF cylinder shell and numerical simulation computation is conductedfor the acoustic mode and in-shell acoustic response the correctness of the finite element model is verified via comparison withmeasured data On this basis influence rules of different MF laying rate and different laying thickness on acoustic cavity resonanceresponse within the low and medium frequency range of 100ndash400Hz are studied It is indicated that noise reduction increases withMF laying rate but the amplification decreases along with the rising of MF laying rate noise reduction per unit thickness decreaseswith the increase of laying thickness while noise reduction per unit area increases

1 Introduction

During the launching process of a carrier rocket a fairingsuffers function of engine jet noise and surrounding aerody-namic noise which lead to a total sound pressure level (SPL)of 120sim140 dB in the fairing [1] In recent years with the largethrust and high load-oriented development of carrier rockettechnologies and application of light high-strength and low-damping composite materials precise instruments protectedby a fairing and other effective loads are under increasinglyhostile noise environment A laying acoustic blanket one ofthe most effective measures for noise attenuation in a fairingcavity has been a focus and difficulty for passive noise controlresearches and has been extensively studied by researchersboth at home and abroad

Pirk and Souto [2] established a simulation model for aninternal laying fiberglass acoustic blanket in the fairing ofBrazilian carrier rocket by utilizing statistical energy anal-ysis software but only discussing noise attenuation effectsof sound absorption materials under different thicknesses

densities and laying rates in medium and high frequencyconditions Gautam [3] laid distributed vibration absorberand a heterogeneous acoustic blanket inside a fairing the het-erogeneous acoustic blanket was made up of heterogeneousmass particle foam materials and showed favorable noiseattenuation effect in low frequencies Similarly Idrisi [4]embedded multiple mass blocks in foam materials to designheterogeneous blanket and created multiple resonant fre-quencies by controlling embedding depths of the individualmass blocks based on which a more extensive low frequencyrange was achieved with noise attenuation effect Hugheset al [5] reduced the noise (frequency range 200ndash250Hz)for additional 3 dB by optimizing density and thickness ofcotton fiber and avoided high cost for battery modificationHowever both mass and volume increase significantly in theabovementioned methods compared to the original acousticblanket Lane et al [6] conducted an experiment in compositefairing to study denoising effects of the thermal protectionsystem for MF material under three tested bandwidths butrarely involved denoising rules of MF materials in fairing

HindawiShock and VibrationVolume 2017 Article ID 6980501 8 pageshttpsdoiorg10115520176980501

2 Shock and Vibration

under low and medium frequency There are relatively fewerresearches concerning noise attenuation technologies forsound field in a fairing with acoustic blanket Sun and Pan[7] established a statistical energy model for rocket fairingand suggested that the SPL in an acoustic cavity was highlyreduced after laying urea plastic foam in a fairing

Based on research status concerning noise issue in afairing in both China and other countries it can be viewedthat porousmaterials or porousmaterialothermaterial com-binations are mostly adopted for an acoustic blanket Amongcommon porous materials MF material shows advantagesof lightness high flexibility and high acoustical absorptioncoefficient for medium and high frequencies [8 9] andis applied by US AFRL in design of noise attenuation incomposite fairings [6] and is taken by NASA Glenn ResearchCenter as one of the optional noise attenuation technologiesfor future large carrier rockets [10] There are a few Chineseresearches on in-cavity noise attenuation performance ofMFShen et al [11 12] selectedMF as their experimental materialset clearance thicknesses in a cavity filled the clearances andachieved favorable noise attenuation effects Liu et al [13]studied impacts of MF density porosity and thickness oncavity noise attenuation performances

It can be viewed from Chinese and foreign research sta-tuses that there are many studies concerning noise reductionperformance of acoustic blanket under high frequency anda rare study involving noise reduction characteristics underlow frequency However low and medium frequency is akey frequency band for combined application of internalporous material and resonance sound absorber when theresonant absorption coefficient is close to themaximumvalueof 09 the combined resonant frequency is not more than400Hz [14] Therefore the paper refers to cylindrical sectionstructure of a carrier rocket fairing to construct a scale cylin-der finite element model and noise experimental devicesTypical porous material MF was adopted as an acousticblanket laid on internal walls of cylinder shells Accordingto objective experimental conditions and application rangeof the simulation software studying noise reduction rulesin MF material cavity under the range of 100ndash400Hz willprovide beneficial reference for studying application rulesand characteristics of acoustic blanket for passive noisecontrol

2 Experimental Research

In order to obtain internal and external sound pressureresponse values of a typical cylinder shell the cylindershell was placed in the center of a reverberation room Anexperimental device as shown in Figure 1 was constructedFor the cylinder cell the body is made of 1mm thicknessaluminum sheet the diameter is 10500mm the height is11150mm the weight is 157 g there are 6 internal transversereinforcing ribs and 20 longitudinal reinforcing ribs and thethickness of every reinforcing rib is 15mm The upper andlower covers aremade ofmedium-density sheet the thicknessis 150mm and the average sound insulation volume is 20 dBAs the volume of the reverberation room is small a rubberband or soft spring-hoisting structure could not be simply

Poweramplier Equalizer Computer Noise

measurement

MicrophoneCylindershell

Reverberationchamber Audio

Figure 1 Sketch of the noise experiment

Figure 2 Sketch of external sound field test

applied and four pieces of 50mm thickness shock-reducingrubber blocks were adopted between the lower cover plate ofthe cylinder shell and the floor to simulate the free state of acylinder shell

During practical measurements Prosig noise and vibra-tion measurement was adopted for objective noise test ofthe cylinder shell When setting acoustic microphones insideand outside the cylinder shell as shown in Figure 2 a totalof 3 acoustic microphones were uniformly distributed at30mmsites of extracavitymiddle layer from the externalwallexternal sound pressure values were measured consistencyof external load was guaranteed As shown in Figure 3 in-cavity area was divided into four planes with 90 degrees apartin radial direction and every plane was divided into uppermiddle and lower layers one acoustic microphone was set atthe radial center and radial endpoints of every layer internalsound pressure values were measured against studied rule ofin-cavity sound pressure response under different conditionsIn conclusion a total of 3 acoustic microphones and 24acousticmicrophoneswere set respectively outside and insidethe cylinder shell cavity During the measurement the signalsampling frequencywas 32000Hz and the sampling timewasset as 8 s the uniformity scope of indoor noise loadwasplusmn3 dBand the deviation range of the noise control spectrum wasplusmn5 dB

Shock and Vibration 3

Middle layer

Upper layer

Lower layer

Innermicrophone

270∘ 180

∘

90∘

0∘

Figure 3 Sketch of arrangement of in-cavity acoustic microphones

3 Numerical Simulation Analysis

31 Simulation Model A finite element model for an entirecylinder shell was established on the basis of Hypermeshfinite element software for the model the cylinder shelland reinforcing rib structure were simulated via CQUAD4surface element the acoustic cavity and the upper and lowercover plates were simulated via CHEXA volume meshingthe connection between the reinforcing ribs and the wallswas simulated via RBE2 element and MF was simulated viaCHEXA volume meshing with node merging with the inter-nal acoustic cavity mesh

The established finite element model was introduced inthe VirtualLab Acoustic simulation software for acousticmeshing pretreatment which defined the structures fluidand material properties while the cylindrical shell and thereinforcing rib material are aluminum whose density is2700 kgm3 Youngrsquos modulus is 70GPa and Poissonrsquos ratiois 033 The upper and lower covers are treated as isotropicmaterials whose density is 800 kgm3 Youngrsquos modulus is2Gpa and Poissonrsquos ratio is 04 The MF acoustic parametersuse the test results from Italian SCS Institute as shown inTable 1 Meanwhile established MPC wire jointing elementsbetween the upper and lower cover plates with the shelldefined coupling between internal and external cavity acous-tic elements and structural elements (including couplingbetween the internal acoustic cavity with MF and internalsurfaces of the upper and lower cover plates and between theexternal acoustic cavity with the shell and external surfaces ofthe upper and lower cover plates) and set different tolerancesaccording to clearances existing on coupling surfaces duringmodeling Meanwhile the plane where the lower surface ofthe external acoustic cavity was located was defined as thereflecting surface to simulate the floor and the remainingexternal surfaces were for free ins and outs of sound wavesand were defined as automatch layer (AML) See Figure 4 fordetails

With the simulation software distributed plane waveswere adopted to stimulate reverberation room environment

Exterior acoustic cavities

Cylindrical shell Interior acoustic cavities

Cover plate

Figure 4 Finite element model map

Table 1 Acoustic parameters of MF

Acoustic parameters ValueFlow resistivity(Pasdotssdotmminus2) 10925Sound speedmsdotsminus1 346Air densitykgsdotmminus3 1185Porosity 099Tortuosity 102Viscous characteristic lengthmm 01Thermal characteristic lengthmm 013

the power spectrum inputs of external loads set in thesoftware were actual parameters obtained via experimentalmeasurement the polarization amount was set as 2 thereverberation sound source was evenly divided into 24plane waves which were uniformly distributed around thesimulationmodel in a circular shape superposition of severalplane waves formed a reverberation filed direct sound andvibration coupling calculationwas then conducted accordingto randomized posttreatment solvers the random frequencyresponse of the acoustic cavity was analyzed The range ofoutput frequency was 0ndash400Hz and the step length was10Hz the output form was power spectral density functionthe SPL curves were achieved at simulation filed pointswhich were located at same positions of the experimentalmeasurement sites and rules of soundpressure responsewerestudied at corresponding measurement sites

32 Simulation Results and Comparative Analysis On thebasis of the established simulation model the correctness ofthe finite element modeling and simulation computation wasverified from three aspects acoustic mode cavity acousticresponse and in-cavity acoustic response of laying MFmaterial

The mode of cylinder cavity determines internal soundfield distribution According to the analytical formula forintrinsic frequency of a cylinder sound cavity the theoretical


Table 2 Comparison of acoustic modes achieved by analytical method and simulation computation

Mode 119897 119898 119899 (1 0 0) (0 1 0) (1 1 0) (2 0 0) (0 2 0) (1 2 0) (2 1 0) (0 0 1)Analytical value Hz 1511 1922 2445 3022 3179 3521 3582 3980Simulation value Hz 1511 1913 2438 3022 3172 3514 3577 3981Error 000 051 031 000 023 019 015 001

value of the acoustic mode in the cylinder shell is calculatedspecifically and it is shown as

119891119897119898119899 = 1198882120587radic( 119897119871120587)2 + 1198962119898119899 (1)

In the formula 119888 refers to sound velocity 119897 refers to axialmode 119898 refers to circumferential mode 119899 refers to radialmode 119871 refers to the axial length of the cylinder acoustic cav-ity 119896119898119899 represents the number 119899 solution of 1198951015840

119898(119896119898119899119903) = 0 119895119898

refers to 119898 order of Bessel function and 119903 refers to radius ofthe cylinder acoustic cavity

The results of the analytical values are compared withsimulation values as shown in Table 2 It can be viewedfrom the following table that the maximal error of analyticalresults is 051 verifying the correctness of the finite elementmodeling

Due to limitation of experimental objective conditionsthe measurement values obtained from the 24 acousticmicrophones in the cylinder shell show no entire consistencyTherefore the comparison is conducted globally betweensimulation values and experimental values of internal acous-tic response before and after laying MF The comparison ofinternal acoustic sound pressure response averages obtainedvia numerical computation and experimental measurementbefore and after laying the material is shown in Figure 5

It can be viewed from the figure that the experimentalresults are basically identical with simulation computationresults within the frequency range of 100ndash400Hz showingthe correctness of the finite element method for the acous-tic response computation within low-mid frequency rangeMeanwhile the curve within the range of 0ndash100Hz showsseveral peaks it can be known by comparing with struc-tural mode that the fact is caused by structural resonancestimulated by external sound filed as shown in Figure 5 thepeak response frequency near amodality shows left shift afterlaying material For example the peak response frequencynear the first-order axial mode (1 0 0) after laying MFmate-rial shifts to the left the difference of response peaks near thefirst-order axial modal frequency before and after laying thematerial is caused by higher wave velocity of sound wave inthe air than in MF After laying MF and in order to guaranteethe continuity at the contact surface between the acousticcavity and MF the axial wave number of the sound wave inthe two shall maintain consistency which leads to reductionof the axial modal frequency of the cavity and further verifiesthe correctness of the simulation model

4 Results and Discussion

To study noise reduction rules of MF acoustic blanket theratio of MF laying area to inner wall area of cylindricalstructure is defined as laying rate and the ratio of MF layingmass to total structural mass is defined as specific gravityratio Within the frequency range of 100ndash400Hz numericalcomputation and analysis are conducted for the 5 resonantfrequency orders (1 0 0) (0 1 0) (1 1 0) (0 2 0) and(2 1 0) in the cylindrical acoustic cavity (Figure 5) to studyinfluence rules of different laying rates laying thicknessesand laying areas on acoustic resonance response and toprovide beneficial reference for in-depth study on applicationrules and characteristics of acoustic blanket for passive noisecontrol

41 Impacts of Laying Rate To explore influence rules oflaying rate on noise reduction performance of MF acousticblanket numerical computation and analysis of noise reduc-tion rules are conducted for 40mm MF with five differentlaying rates (0 25 50 75 and 100)

Figure 6 shows average SPL response spectrum of 12measurement points in a cylindrical acoustic cavity understimulation of external noise It can be viewed from the figurethat under an empty-cavity state resonance points of differ-ent orders are stimulated to different extents due to soundand vibration coupling effect frequencies of different ordersin the acoustic cavities are enhanced to different degreeswithin the frequency range of 100ndash150Hz acoustic responseof internal cylindrical cavity is determined by structuralstiffness additionally within the range response of internalacoustic cavity changes with the variation of noise reductionthe noise reduction is higher and internal acoustic response ishigher with a higher laying rate therefore internal responseof the empty cavity is lowest while that of the acoustic cavitywith full laying ranks at the top [15] Within the frequencyrange of 150ndash400Hz the internal SPL curvilinear trends afterlaying MF materials are basically consistent a higher layingrate results in better noise reduction outcome however whenthe laying rate is higher than 25 the overlap ratio of internalSPL curves increases indicating that the amplification ofnoise reduction within the frequency range decreases withthe increase of MF laying rate

It can also be viewed from Figure 6 that there are severalobvious resonance orders within the frequency range of100ndash400Hz among those the response of the (1 1 0) orderacoustic cavity resonance is not prominent which is majorlyrelated to the higher thickness of upper and lower covers thansidewall whenMFacoustic blanket is applied the order reso-nance is difficult to be identified For the five resonance orders


Simulation(No laying material)Experiment(No laying material)

Simulation (laying 40 mm melamine material)Experiment (laying 40 mm melamine material)

50

60

70

80

90In

tern

al ac

ousti

c res

pons

e (db

)

50

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

db)

100 200 300 4000Frequency (Hz)

100 200 300 4000Frequency (Hz)

(1 0 0)(0 1 0)

(1 1 0)

(0 2 0) (2 1 0)

(1 0 0)(0 1 0) (1 1 0) (0 2 0)

(2 1 0)

Figure 5 Comparison of internal response spectrum achieved from experiment and simulation

No laying materialLaying rate 25Laying rate 75

Laying rate 50Laying rate 100

200 250 300 350 400150Frequency (Hz)

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

dB)

(1 0 0) (0 1 0)

(1 1 0)

(0 2 0)(2 1 0)

Figure 6 Average sound pressure level inside cylinder with different area coverage


indicated in Figure 6 acoustic cavity resonance decreaseswith the increase of material laying rate suggesting that MFacoustic blanket shows significant denoising performanceit can also be viewed that 75 and 100 laying rates showrelatively small difference of denoising performance butbetter performances than 25 and 50 laying rate Withthe increase of frequency medium-high frequency acousticalabsorption coefficient of MF increases and a higher layingrate leads to better noise reduction performance In additionfor high-order acoustic cavity resonance the noise reductionswith the increase of laying rate are higher than low-orderresonance

To further explore the influence rules of laying rate andspecific gravity ratio on noise reduction performance ofMF acoustic blanket numerical analysis and discussion areconducted for noise reductions of 40mm MF material inresonance frequency orders of (1 0 0) (0 1 0) (1 1 0)(0 2 0) and (2 1 0)with the frequency range of 100ndash400HzNoise transmission and acoustic blanket performance weremeasured by computing the noise reduction which is definedhere to be the ratio of the spatially averaged external soundfield impinging on the cylindrical shell to the spatiallyaveraged interior acoustic response The computation is asfollows

NR (dB) = 20 log10 (External rmsInternal rms

) (2)

To estimate the ldquoExternal rmsrdquo (where rms denotes root-mean square) of the external sound field 3 microphonesmeasurements were taken at different locations around thecylindrical cavity exterior and spatially averaged The ldquoInter-nal rmsrdquo was estimated in a similar way by taking 24 micro-phones measurements at many locations throughout thecylindrical cavity interior

In Figures 7 and 8 the noise reduction is presented bothas a function of laying rate and as a function of specificgravity ratio That is to show that curves of noise reductionchange with laying rate and specific gravity ratio of 20mmand 40mm MF materials Values of the points in thecurves are obtained via formula (2) For each order linearregressions were computed for the corresponding data andare superimposed on the data points In each case therewere similar trends and the noise reduction appears to besomewhat linear with respect to the amount of acoustictreatment The midpoint of the curve is the average noisereduction while the slope of the curve indicates influenc-ing degree It can be viewed from Figures 7 and 8 thatfor (1 0 0)- (0 1 0)- (1 1 0)- (0 2 0)- and (2 1 0)-ordercenter frequency points the average noise reductions are395 dB 152 dB 233 dB 455 dB and 489 dB respectivelyunder a specific gravity ratio of 225 and 494 dB 261 dB336 dB 572 dB and 635 dB under a specific gravity of 450It can be viewed that increasing MF thickness significantlyincreases low and medium frequency sound absorptioneffects However the increase of noise reduction does notshow an exact direct proportional relation with the increaseof specific gravity ration Except for the (1 0 0) order theaverage noise reduction and the slope of curve of centralfrequency points both increase with frequency suggesting

40 60 80 10020Surface area coverage ()

0

2

4

6

8

10

Noi

se re

duct

ion

(dB)

0 09 18 27 36 45Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)

Figure 7 Curves of noise reductionwith changing of laying rate andproportion (20mm)


0

2

4

6

8

12

10

Noi

se re

duct

ion

(dB)

0 18 36 54 72 90Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)


that noise reduction performance of MF material graduallyimproves with the increase of frequency and laying rate andthat laying rate poses significantly high impacts on noisereduction of (1 0 0)-order central frequency point Whenthe low frequency acoustical absorption coefficient of MF isrelatively low improving laying rate effectively reduces lowfrequency noise

42 Noise Reductions per Unit Thickness and Unit AreaIn addition to satisfying requirements of noise reduction


Table 3 Noise reductions as a function of thickness and surface area coverage

Order Noise reductions per unit thickness (dB10mm) Noise reductions per unit area (dBm2)20mm 60mm 100mm 20mm 60mm 100mm

(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319

performance structural lightness shall also be considered forrealization of engineering application Therefore with MF asthe study subject numerical computation is used to studynoise reduction characteristics per unit thickness and unitarea of MF acoustic blanket

The total area of internal cylindrical shell wall is 3709m2the noise reductions per unit thickness and unit area areshown as in Table 2 For cases with same laying thicknessmaximal noise reductions per unit thickness and unit areaare both shown in (1 1 0) It can be viewed from the tablethat with the increase of thickness the noise reduction perunit thickness shows a gradually declining tendency whilethe noise reductions per unit area gradually increase withthickness

It can also be viewed from SPL in Table 3 that the thinnertreatment had the higher reduction per unit thickness relativeto the other two treatments which is reasonable since 20mmMF covers more surface area than the same amount of theother two groups For a given mass limit the 20mm foamoffered the highest noise reduction per unit thickness whichis 396 dB In fact the 20mm material offered nearly 2 to3 times the performance as the other two groups The dataalso indicate that if the surface area available for treatmentis limited but the thickness of the treatment is not then the100mm treatment offers the best noise control solution andthe noise reduction per unit area for 100mm MF is 319 dBThis is reasonable considering that 1m2 of 100mm foamis a considerable amount of acoustic treatment Thereforefor a given mass limit smaller laying thickness leads tohigher noise reduction per unit thickness When the layingarea is limited higher laying thickness leads to higher noisereduction per unit area however with the increase of layingmaterial thickness the amplification of noise reduction perunit area gradually decreases

5 Conclusion

In the paper the noise control method of laying acousticblanket is adopted to establish a finite element model ofcylinder shell and noise experiment device to test internalnoise responses of structural cavity of the cylinder shell andcavity with layingMFmaterialThe results obtained via finiteelement method and experimental method are comparedto verify the correctness of the finite element modelingBased on the finite element model within the frequencyrange of 100ndash400Hz numerical computation and analysis

are conducted for cylindrical acoustic cavity (1 0 0) (0 1 0)(1 1 0) (0 2 0) and (2 1 0) orders to study impacts of differ-ent laying rate laying thickness and laying areas on acousticcavity resonance response The following conclusions aremade

(1) Within the frequency range of 100ndash400Hz noisereduction increases with the MF laying rate butthe amplification decreases In addition within thefrequency range of 100ndash150Hz a higher MF layingrate leads to a higher internal acoustic response forthe frequency range of 150ndash400Hz the internalacoustic response decreases with the increase of MFlaying rate

(2) Within the frequency range of 100ndash400Hz noisereduction performance of MF material graduallyincreases along with frequency and laying rate noisereduction performance improves faster with thickerMF materials In addition the impact of laying rateis relatively significant for (1 0 0)-order acousticcavity resonance and the average noise reductionsare 395 dB (20mm MF) and 494 dB (40mm MF)respectively

(3) For a givenmass limit a smaller layingmaterial thick-ness leads to a higher noise reduction per unit thick-ness When the area of laying material is limited alarger laying material thickness leads to higher noisereduction per unit area however with the increase oflaying thickness the amplification of noise reductionper unit area decreases gradually

Conflicts of Interest

The authors declare that there are no conflicts of interestregarding the publication of this article

Acknowledgments

This work is supported by Guangxi Natural Science Foun-dation (2016GXNSFAA380211) and Fundamental ResearchFunds for the Central Universities (WUT 2017IB016) Thiswork was finished at Wuhan University of Technology(WUT) Wuhan


References

[1] J-L Rong X-Y Chen B Li X-Y Cheng and H-B LildquoA method for noise attenuation in cylindrical cavity withmelamine foam liningrdquo Journal of Astronautics vol 37 no 10pp 1271ndash1278 2016

[2] R Pirk and C D Souto ldquoImplementation of acoustic materialsto the VLS-1 fairing - A sensitivity analysis using SEArdquo Journalof the Brazilian Society of Mechanical Sciences and Engineeringvol 34 no 1 pp 82ndash88 2012

[3] A GautamDesign and Development of Advanced Vibration andNoise Control Devices Using Finite Element Analysis VirginiaPolytechnic Institute and State University Blacksburg VAUSA 2005

[4] K Idrisi Heterogeneous (HG) Blankets for Improved AircraftInterior Noise Reduction Virginia Polytechnic Institute andState University Blacksburg VA USA 2005

[5] W O Hughes AMMcNelis andH Himelblau ldquoInvestigationof acoustic fields for the cassini spacecraft Reverberant versuslaunch environmentsrdquo in Proceedings of the Aeroacoustics Con-ference and Exhibit AIAACEAS 1999 pp 1193ndash1203 BellevueWashington USA 1999

[6] S A Lane S Kennedy and R Richard ldquoNoise transmissionstudies of an advanced grid-stiffened composite fairingrdquo Journalof Spacecraft and Rockets vol 44 no 5 pp 1131ndash1139 2007

[7] M Sun and Z W Pan ldquoNoise Environment Prediction andAnti-acoustic Design of Payload Fairingrdquo Missiles and SpaceVehicles vol no 4 pp 6ndash10 2008

[8] G H Yuan X C Wang P Z Hou and C L Li ldquoAbsorbingProperty of Open-cell Melamine Foamrdquo International Journalof Mechanical and Materials Engineering vol 31 no 9 pp 55ndash57 2007

[9] J P Arenas and M J Crocker ldquoRecent trends in porous sound-absorbingmaterialsrdquo Sound and Vibration vol 44 no 7 pp 12ndash17 2010

[10] W O Hughes A M Mcnelis and M E McNelis ldquoAcousticTest Characterization of Melamine Foam for Usage in NASAsPayload Fairing Acoustic Attenuation Systemsrdquo in Proceedingsof the in Proceedings of the 28th Aerospace Testing SeminarSponsored by the Aerospace Corporation Los Angeles CalifUSA 2014

[11] Y P Shen Application of The New Sound-Absorbing Materialin Ship Cabin Noise Reduction Ocean University of ChinaQingdao 2014

[12] Y P Shen Y C Yang C L Chen andCWang ldquoThe effect of thespace between trimpanel and bulkheadplate onnoise reductionof cabinrdquo Ship Ocean Engineering vol 43 no 3 pp 37ndash40 2014

[13] K Liu W G Wu and B Qiu ldquoFull spectrum simulationprediction of high-speed vessel cabin noiserdquo TransportationScience and Technology vol 2 pp 109ndash112 2010

[14] L He H C Zhu X J Qiu and G H Du Acoustic Theory andEngineering Application Science Press Beijing 2006

[15] D Li Vibro-Acoustic Behavior and Noise Control Studies ofAdvanced Composite Structures University of Pittsburgh Pitts-burgh PA USA 2003

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under low and medium frequency There are relatively fewerresearches concerning noise attenuation technologies forsound field in a fairing with acoustic blanket Sun and Pan[7] established a statistical energy model for rocket fairingand suggested that the SPL in an acoustic cavity was highlyreduced after laying urea plastic foam in a fairing

Based on research status concerning noise issue in afairing in both China and other countries it can be viewedthat porousmaterials or porousmaterialothermaterial com-binations are mostly adopted for an acoustic blanket Amongcommon porous materials MF material shows advantagesof lightness high flexibility and high acoustical absorptioncoefficient for medium and high frequencies [8 9] andis applied by US AFRL in design of noise attenuation incomposite fairings [6] and is taken by NASA Glenn ResearchCenter as one of the optional noise attenuation technologiesfor future large carrier rockets [10] There are a few Chineseresearches on in-cavity noise attenuation performance ofMFShen et al [11 12] selectedMF as their experimental materialset clearance thicknesses in a cavity filled the clearances andachieved favorable noise attenuation effects Liu et al [13]studied impacts of MF density porosity and thickness oncavity noise attenuation performances

It can be viewed from Chinese and foreign research sta-tuses that there are many studies concerning noise reductionperformance of acoustic blanket under high frequency anda rare study involving noise reduction characteristics underlow frequency However low and medium frequency is akey frequency band for combined application of internalporous material and resonance sound absorber when theresonant absorption coefficient is close to themaximumvalueof 09 the combined resonant frequency is not more than400Hz [14] Therefore the paper refers to cylindrical sectionstructure of a carrier rocket fairing to construct a scale cylin-der finite element model and noise experimental devicesTypical porous material MF was adopted as an acousticblanket laid on internal walls of cylinder shells Accordingto objective experimental conditions and application rangeof the simulation software studying noise reduction rulesin MF material cavity under the range of 100ndash400Hz willprovide beneficial reference for studying application rulesand characteristics of acoustic blanket for passive noisecontrol

2 Experimental Research

In order to obtain internal and external sound pressureresponse values of a typical cylinder shell the cylindershell was placed in the center of a reverberation room Anexperimental device as shown in Figure 1 was constructedFor the cylinder cell the body is made of 1mm thicknessaluminum sheet the diameter is 10500mm the height is11150mm the weight is 157 g there are 6 internal transversereinforcing ribs and 20 longitudinal reinforcing ribs and thethickness of every reinforcing rib is 15mm The upper andlower covers aremade ofmedium-density sheet the thicknessis 150mm and the average sound insulation volume is 20 dBAs the volume of the reverberation room is small a rubberband or soft spring-hoisting structure could not be simply

Poweramplier Equalizer Computer Noise

measurement

MicrophoneCylindershell

Reverberationchamber Audio

Figure 1 Sketch of the noise experiment

Figure 2 Sketch of external sound field test

applied and four pieces of 50mm thickness shock-reducingrubber blocks were adopted between the lower cover plate ofthe cylinder shell and the floor to simulate the free state of acylinder shell

During practical measurements Prosig noise and vibra-tion measurement was adopted for objective noise test ofthe cylinder shell When setting acoustic microphones insideand outside the cylinder shell as shown in Figure 2 a totalof 3 acoustic microphones were uniformly distributed at30mmsites of extracavitymiddle layer from the externalwallexternal sound pressure values were measured consistencyof external load was guaranteed As shown in Figure 3 in-cavity area was divided into four planes with 90 degrees apartin radial direction and every plane was divided into uppermiddle and lower layers one acoustic microphone was set atthe radial center and radial endpoints of every layer internalsound pressure values were measured against studied rule ofin-cavity sound pressure response under different conditionsIn conclusion a total of 3 acoustic microphones and 24acousticmicrophoneswere set respectively outside and insidethe cylinder shell cavity During the measurement the signalsampling frequencywas 32000Hz and the sampling timewasset as 8 s the uniformity scope of indoor noise loadwasplusmn3 dBand the deviation range of the noise control spectrum wasplusmn5 dB


Middle layer

Upper layer

Lower layer

Innermicrophone

270∘ 180

∘

90∘

0∘








Cover plate











119891119897119898119899 = 1198882120587radic( 119897119871120587)2 + 1198962119898119899 (1)


119898(119896119898119899119903) = 0 119895119898













50

60

70

80

90In

tern

al ac

ousti

c res

pons

e (db

)

50

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

db)

100 200 300 4000Frequency (Hz)

100 200 300 4000Frequency (Hz)

(1 0 0)(0 1 0)

(1 1 0)

(0 2 0) (2 1 0)

(1 0 0)(0 1 0) (1 1 0) (0 2 0)

(2 1 0)




200 250 300 350 400150Frequency (Hz)

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

dB)

(1 0 0) (0 1 0)

(1 1 0)

(0 2 0)(2 1 0)






) (2)




0

2

4

6

8

10

Noi

se re

duct

ion

(dB)

0 09 18 27 36 45Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)



0

2

4

6

8

12

10

Noi

se re

duct

ion

(dB)

0 18 36 54 72 90Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)







(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319




5 Conclusion








Acknowledgments



References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation










Middle layer

Upper layer

Lower layer

Innermicrophone

270∘ 180

∘

90∘

0∘








Cover plate











119891119897119898119899 = 1198882120587radic( 119897119871120587)2 + 1198962119898119899 (1)


119898(119896119898119899119903) = 0 119895119898













50

60

70

80

90In

tern

al ac

ousti

c res

pons

e (db

)

50

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

db)

100 200 300 4000Frequency (Hz)

100 200 300 4000Frequency (Hz)

(1 0 0)(0 1 0)

(1 1 0)

(0 2 0) (2 1 0)

(1 0 0)(0 1 0) (1 1 0) (0 2 0)

(2 1 0)




200 250 300 350 400150Frequency (Hz)

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

dB)

(1 0 0) (0 1 0)

(1 1 0)

(0 2 0)(2 1 0)






) (2)




0

2

4

6

8

10

Noi

se re

duct

ion

(dB)

0 09 18 27 36 45Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)



0

2

4

6

8

12

10

Noi

se re

duct

ion

(dB)

0 18 36 54 72 90Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)







(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319




5 Conclusion








Acknowledgments



References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation













119891119897119898119899 = 1198882120587radic( 119897119871120587)2 + 1198962119898119899 (1)


119898(119896119898119899119903) = 0 119895119898













50

60

70

80

90In

tern

al ac

ousti

c res

pons

e (db

)

50

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

db)

100 200 300 4000Frequency (Hz)

100 200 300 4000Frequency (Hz)

(1 0 0)(0 1 0)

(1 1 0)

(0 2 0) (2 1 0)

(1 0 0)(0 1 0) (1 1 0) (0 2 0)

(2 1 0)




200 250 300 350 400150Frequency (Hz)

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

dB)

(1 0 0) (0 1 0)

(1 1 0)

(0 2 0)(2 1 0)






) (2)




0

2

4

6

8

10

Noi

se re

duct

ion

(dB)

0 09 18 27 36 45Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)



0

2

4

6

8

12

10

Noi

se re

duct

ion

(dB)

0 18 36 54 72 90Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)







(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319




5 Conclusion








Acknowledgments



References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation












50

60

70

80

90In

tern

al ac

ousti

c res

pons

e (db

)

50

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

db)

100 200 300 4000Frequency (Hz)

100 200 300 4000Frequency (Hz)

(1 0 0)(0 1 0)

(1 1 0)

(0 2 0) (2 1 0)

(1 0 0)(0 1 0) (1 1 0) (0 2 0)

(2 1 0)




200 250 300 350 400150Frequency (Hz)

60

70

80

90

Inte

rnal

acou

stic r

espo

nse (

dB)

(1 0 0) (0 1 0)

(1 1 0)

(0 2 0)(2 1 0)






) (2)




0

2

4

6

8

10

Noi

se re

duct

ion

(dB)

0 09 18 27 36 45Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)



0

2

4

6

8

12

10

Noi

se re

duct

ion

(dB)

0 18 36 54 72 90Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)







(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319




5 Conclusion








Acknowledgments



References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation













) (2)




0

2

4

6

8

10

Noi

se re

duct

ion

(dB)

0 09 18 27 36 45Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)



0

2

4

6

8

12

10

Noi

se re

duct

ion

(dB)

0 18 36 54 72 90Specic gravity ()

(1 0 0)

(0 1 0)

(1 1 0)

(0 2 0)

(2 1 0)







(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319




5 Conclusion








Acknowledgments



References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation












(1 0 0) 362 148 103 195 239 270(0 1 0) 349 237 175 188 383 419(1 1 0) 551 302 224 297 430 502(0 2 0) 313 294 202 269 401 434(2 1 0) 238 236 176 204 338 397SPL 396 175 118 246 283 319




5 Conclusion








Acknowledgments



References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation










References
















RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation









RoboticsJournal of





Journal of



RotatingMachinery



Journal of

Volume 201


VLSI Design



Shock and Vibration







Journal of



Volume 2014


SensorsJournal of





Propagation









experiment and simulation analysis on noise attenuation of

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