STRUCTURAL VIBRATION AS A NOISE SOURCE ON VIBRATING SCREENS 

David S. Yantek 

National Institute for Occupational Safety and Health 

Pittsburgh, PA, USA 

Hugo R. Camargo 

National Institute for Occupational Safety and Health 

Pittsburgh, PA, USA 

ABSTRACT 

A-weighted sound levels around vibrating screens in coal 

preparation plants often exceed 90 dB(A). The National 

Institute for Occupational Safety and Health (NIOSH) is 

actively developing noise controls to reduce noise generated by 

horizontal vibrating screens. A 121-microphone, 3.5-meterdiameter array was used to perform beamforming to determine 

locations of significant noise radiation on the screen. Below 

about 1 kHz, the screen body was found to be the most 

significant noise source. The beamforming contour maps 

showed several key locations on the sides of the screen and the 

feedbox are the most significant contributors. Operating 

deflection shape (ODS) analysis was used to examine the 

screen behavior under actual operating conditions. This 

information is helpful in determining how to modify the screen 

body to reduce the noise radiated by the screen below 1 kHz. 

The analysis showed modal vibration patterns on the sides and 

feed box were the main contributors to noise. The results show 

several areas on the screen sides and feedbox that can be 

modified to reduce noise. 

INTRODUCTION 

In 1996, NIOSH published the National Occupational 

Research Agenda (NORA), which identified hearing loss as the 

most common job-related disease in the United States [1]. 

Approximately 30 million workers are exposed to hazardous 

sound levels alone or to hazardous sound levels in conjunction 

with ototoxic agents [2]. Despite more than 30 years of noise 

regulation in the mining industry, mine workers develop 

hearing loss at a significantly higher rate compared to the nonnoise exposed population. An analysis of audiograms 

conducted by NIOSH in 1996 shows that by the age of 50, 

nearly 90% of coal miners had a hearing impairment [3]. In 

contrast, only 10% of those who are not exposed to 

occupational noise experienced a hearing loss by the same age. 

The Mine Safety and Health Administration (MSHA) 

modified its rules regarding noise exposure in 1999 in an effort 

to reduce the occurrence of noise-induced hearing loss [4]. 

Rather than relying solely on hearing protection devices, 

MSHA’s new rule requires mine operators to use all feasible 

engineering and/or administrative controls to reduce the noise 

exposures of overexposed miners’. However, for many 

machines, such as vibrating screens, noise controls that reduce 

the operator’s noise exposure below the MSHA Permissible 

Exposure Level (PEL) are not currently available. 

In 2000, there were 212 preparation plants in operation in 

the US and 129 of these plants were located in three states: 

Kentucky, Pennsylvania, and West Virginia [5]. NIOSH studies 

have shown that workers who spend a significant portion of 

their shift working in a coal preparation plant can experience 

noise exposures which exceed the MSHA PEL for noise. 

NIOSH data show that 20 out of 46 coal preparation plant 

workers had noise exposures that exceeded the MSHA PEL 

noise dose [6]. MSHA PEL noise doses up to 220% have been 

recorded for preparation plant workers in jobs with titles such 

as stationary equipment operator, froth cell operator, plant 

operator, plant controls man, third floor operator, wet plant 

attendant, sump floor operator, plant backup, and plant 

mechanic. These job classifications require the worker to spend 

a significant portion of a shift in the plant while working 

around slurry pumps, dryers, centrifuges, and vibrating screens. 

A horizontal vibrating screen (see Figure 1) is a large 

machine used to process clean coal that has been separated 

from refuse materials using a water-magnetite mixture. This 

magnetite is recovered because the magnetite lowers the 

heating value of coal and it can be re-used in the processing 

plant. The screen body has four sides made of steel plates with 

a bottom screening surface made of steel wire welded to a 

frame with small gaps between the wires. The body of the 

screen is supported on a steel coil spring suspension. One or 

more vibration mechanisms are mounted to a steel beam that 

spans the width of the screen. These vibration mechanisms, 

which use rotating eccentric shafts to generate vibration, are 

belt-driven using an electric motor. The screen is designed such 

that it vibrates on roughly a 45-degree angle. Coal flows into 

 1 
the feed end of the screen from a delivery chute. As the screen 

the material moves along the deck and under a water spray that 

rinses the magnetite from the coal. The liquid and fine coal 

particles pass through the gaps in the screening deck as the 

material flows toward the discharge end of the screen. Finally, 

the rinsed coal falls off the discharge end of the chute to 

continue with further processing. 

suspension

springs

electric

motor

FEED

END

eccentric

mechanisms

rib

feedbox

H-beam

suspension

springs

electric

motor

FEED

END

eccentric

mechanisms

rib

feedbox

H-beam

(a) 

left side

electric

motor

DISCHARGE

END

eccentric

mechanisms

belt guard

frame rail

screening

deck

H-beam

left side

electric

motor

DISCHARGE

END

eccentric

mechanisms

belt guard

frame rail

screening

deck

H-beam

(b) 

Figure 1. A horizontal vibrating screen as viewed 

from (a) feed end and (b) discharge end. 

Since they are used to size, separate, and dewater both coal 

and refuse (rock) of various sizes, screens may be located on 

many floors within a preparation plant. The number of screens 

in a processing plant can range from a single screen to more 

than a dozen. Consequently, preparation plant workers can be 

exposed to high sound levels generated by screens multiple 

times during a shift as they move and work throughout the 

plant. Vibrating screens are a major noise problem in most coal 

preparation plants because screens are used extensively in 

plants, are usually located in high traffic areas, and can 

generate high noise levels [7]. The objective of this paper is to 

examine the noise radiated by horizontal vibrating screens 

below 1 kHz. 

vibrates, 

NIOSH performed sound level measurements near a group 

of eight horizontal vibrating screens used to process clean coal 

[8]. These measurements indicated that the sound levels ranged 

from 94 to 98 dB(A) with the plant processing coal. With the 

coal flow turned off and the screen vibration mechanisms 

turned on, the sound levels ranged from 89 to 97 dB(A). The 

sound levels decreased significantly with increasing distance 

from the screens, indicating that the screens dominate the 

overall A-weighted sound level in this area of the preparation 

plant. In order to reduce the potential for overexposing 

preparation plant workers to noise, noise controls must be 

developed to address dominant noise sources on the screen. 

The sound power level of a screen was measured in the 

NIOSH reverberation chamber. The data show that the onethird-octave-band sound power levels below 1 kHz account for 

roughly 80% of the overall A-weighted sound power level. 

NIOSH applied the beamforming technique using a 42-

microphone, 1.9-meter-diamteter array to identify noise sources 

on a vibrating screen [9]. Data were collected with the array 

positioned to the front, rear, and sides of the machine with the 

screen in the NIOSH hemi-anechoic chamber. The results 

showed that the vibration mechanism housings, belt guard, and 

steel coil springs are the dominant noise sources from 1 kHz 

through 3.15 kHz. The 42-microphone array was too small to 

examine noise radiated below 1 kHz. 

The noise radiated by the screen at frequencies below 1 

kHz was examined because this frequency range accounts for 

about 80% of the overall A-weighted sound power level. The 

dominant noise sources at these frequencies must be addressed 

to reduce the radiated noise. Since the screen is relatively large, 

sound intensity and near field acoustic holography would 

require numerous measurements to examine the sources from 

all sides of the screen. Beamforming enables researchers to 

perform relatively few measurements to examine noise 

radiation. However, a large array must be used to achieve 

sufficient spatial resolution. Operating deflection shape (ODS) 

analysis was used to examine the vibration patterns on the 

screen. This information is essential to understand how to 

modify the screen to reduce noise. In addition, the sound power 

level radiated by the sides and feed box was estimated using 

the ODS data. 

NOMENCLATURE 

[X] Response matrix 

[H] Frequency response function matrix 

[F] Force matrix 

L sound power level W

v

2

 surface averaged mean-squared velocity s ,t

S surface area 

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σ radiation efficiency

BEAMFORMING MEASUREMENTS AND RESULTS 

NIOSH contracted Acoustical and Vibration Engineering 

Consultants (AVEC) to perform beamforming measurements 

using their 121-microphone, 3.5-meter-diameter array. The 

vibrating screen was positioned in the center of the NIOSH 

hemi-anechoic chamber. The chamber dimensions are 

approximately 16.7 meters long, 10.1 meters wide, and 7.0 

meters high. AVEC’s phased array is an 11-arm star array with 

a proprietary microphone arrangement. The array was mounted 

to a movable truss to position the array for measurements from 

each screen surface (see Fig. 2).

(a) 

(b) 

Figure 2. AVEC array mounted to a truss in the (a) 

vertical configuration and (b) horizontal 

configuration. 

 It was mounted in two 

configurations: vertical (microphone plane perpendicular to the 

floor) and horizontal (microphone plane parallel to the floor). 

With the array in the horizontal configuration, data were 

collected at three positions above the screen. Data was also 

collected from each side of the machine with the array in the 

vertical configuration. 

A 128-channel acquisition system with proprietary signal 

conditioning and anti-aliasing filtering was used to collect the 

microphone data. All microphone signals were sampled 

simultaneously using a sampling frequency of 51.2 kHz. For 

each measurement, 16 seconds of continuous data were 

collected. This data was post-processed using AVEC 

beamforming analysis software. The results were examined in 

one-third-octave bands. Contour maps were computed for each 

array position using conventional beamforming with diagonal 

removal [10]. Multiple source plane, or calculation plane, 

distances were used to process the data for each measurement. 

This provides the ability to examine noise sources at different 

depths relative to the array. Due to space limitations, only the 

most significant results will be discussed. 

Prior to examining screen data, a test was conducted with a 

loudspeaker to examine the spatial resolution. The researchers 

confirmed the spatial resolution of the array was adequate 

down to 250 Hz. For the 250 Hz one-third-octave band, the 

contour maps appear to indicate the screen sides were the most 

significant source (see Fig. 3). Most of the noise radiation 

appears to occur from approximately the center of the screen 

side. 

Figure 3. Beamforming results for the rt. side for the 

250 Hz 1/3-OB. 

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Figure 4 shows the contour maps with the array positioned 

above the discharge end of the screen and with the array 

centered above the screen for the 315 Hz one-third-octave 

band. 

(a) 

(b) 

Figure 4: Beamforming results for the 315 Hz 1/3-OB 

with array (a) over discharge end and (b) over center. 

The images in Fig. 4 appear to indicate several locations 

of high noise radiation along the screen sides and from the feed 

box. Multiple ribs are positioned along the screen sides to 

increase the stiffness of the plates that make up the sides. The 

spacing between the ribs is non-uniform. The area of high noise 

radiation near the discharge end of the screen (see Fig. 4a) 

corresponds to the location just in front of the large H-beam 

that connects the screen sides. The area of high noise radiation 

on the screen sides towards the feed end of the machine (see 

Fig. 4b) corresponds to locations just in front of a large tube 

that connects the screen sides. It is interesting to note that the 

locations near the feed and discharge ends of the screen that 

have the highest noise radiation are in areas with the largest 

spans between the ribs. Figure 4b clearly indicates the back 

and/or the bottom of the feed box are also significant 

contributors to noise radiated in this band. 

For the 400 Hz one-third-octave band, the screen sides 

appear to be the most significant noise sources. Figure 5 shows 

resulting contour map with the array positioned at the discharge 

end (a) and with the array to the right side (b).

(a) 

(b) 

Figure 5: Beamforming results for the 400 Hz 1/3-OB 

with (a) array at discharge end and (b) array to rt. 

side. 

 The calculation 

planes for these images correspond to the plane of the electric 

motor and the screen side, respectively. The image from the 

discharge end of the screen indicates two sources that seem to 

be along the sides of the screen. The contour for the screen side 

clearly shows most of the noise is radiated from the sides. The 

area of highest radiation matches the area between two ribs just 

below where the H-beam connects to the screen sides. 

Figure 6 shows the contour map for the right side of the 

screen with the calculation plane at the screen side. Once again, 

the highest radiation appears to correspond to the area between 

two ribs and below the H-beam attachment point. A secondary 

source appears to match the area just above the screen 

suspension brackets. 

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Figure 6. Beamforming results for the 500 Hz 1/3-OB 

for the rt. side with calculation plane at screen side. 

Figure 7 shows the beamforming image for the right side 

of the screen with the calculation plane at the screen side. The 

image shows the most significant source is at a location 

between two stiffeners that is just behind the H-beam 

attachment location. A secondary source is shown near the top 

of this panel. The image also shows either the belt guard or the 

right vibration mechanism housing is also a significant source. 

Figure 7. Beamforming results for the 630 Hz 1/3-OB 

for the rt. side with calculation plane at screen side. 

The contour map for the 800 Hz one-third-octave band 

with the calculation plane at the eccentric shaft location is 

shown in Fig. 8a. Several sources appear to be significant for 

this band. Both the left and right vibration mechanism housings 

are shown to be sources. In addition, the area between the belt 

guard and the right vibration mechanism housing appears to be 

a source of high noise radiation. Finally, two areas along the 

screen sides appear to be significant. The image for the right 

side confirms the location of the sources along the screen side 

(see Fig. 8b).

(a) 

(b) 

Figure 8. Beamforming results for the 800 Hz 1/3-OB 

with array (a) at feed end and (b) to rt side. 

 Once again this location coincides with a location 

between two stiffeners that is just below the H-beam 

attachment point. 

The results for the 1 kHz one-third-octave band are shown 

in Fig. 9 with the source plane set to the eccentric motor 

location. The image from the discharge end indicates the area 

near the top of the right vibration mechanism and the electric 

motor is the most significant contributor to noise in this band. 

The image from above the screen shows the electric motor is 

the most significant source for this band. 

The beamforming results point to the screen sides and feed 

box as dominant noise sources for each one-third-octave band 

in the 250 Hz through 800 Hz bands whereas the electric motor 

appears to be the most significant source for the 1 kHz band. It 

should be noted that beamforming has a limited range in terms 

of identifying noise sources. In most cases, only the top 6 to 15 

dB of noise sources can be identified with this technique. Since 

structural vibration at lower frequencies is generally of a global 

nature, further analysis was needed to understand the screen’s 

dynamic behavior for the bands below 800 Hz. Therefore, ODS 

data collection and analysis was performed. 

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(a) 

(b) 

Figure 9. Beamforming results for the 1 kHz 1/3-OB 

with the array (a) at the discharge end and (b) above 

the discharge end. 

OPERATING DEFLECTION SHAPE TESTING AND 

ANALYSIS 

Structural vibration is complex in nature. All structures 

exhibit their own dynamic characteristics that depend on the 

mass, stiffness, and damping of the structure. Structures exhibit 

resonant frequencies that yield high vibration and noise levels 

when excited by operating forces. Each resonant frequency has 

its own characteristic shape referred to as a mode shape. The 

response of a structure to a vibratory input force is described by 

[ ] X = [H] [F] (1)

where [X] is the displacement at each location, [H] is the 

transfer function matrix, and [F] is the matrix of applied 

forces [11]. The transfer function matrix is a function of 

frequency; hence it is sometimes called the frequency response 

function (FRF) matrix. The FRF matrix depends on the mass, 

stiffness, and damping of a structure and it contains all the 

natural frequency and mode shape information of a structure. 

Experimental modal analysis is a technique that is used to 

extract the natural frequencies and mode shapes of a structure. 

In the application of this technique, broadband excitation is 

applied to the structure using either an impact hammer or one 

or more electromechanical shakers at discrete locations. The 

applied forces and resulting vibration at numerous locations are 

measured and FRFs are computed between each force and 

response location. This data is then curve fit to determine the 

natural frequencies and mode shapes. This data can be used to 

examine the dynamic characteristics of a structure and for 

purposes of benchmarking finite element models, for example. 

However, the actual behavior of an object subject to 

vibration is the result of both the natural frequencies and mode 

shapes of the structure and the real operating forces. ODS 

testing is a technique that can be applied to examine this 

operating behavior. The operating forces do not have to be 

measured when performing ODS testing. In practice, the 

vibration at numerous locations is measured on a structure by 

roving one or more accelerometers around the structure. In 

addition, one or more reference accelerometers are positioned 

at fixed locations. For frequency-based ODS testing, the 

measured autopower spectra at each location is computed along 

with the cross-spectra between the roving accelerometers and 

the reference accelerometer(s). The autopower spectra are used 

to determine the magnitude of the response at each location and 

the cross-spectra are used to determine the phase relationship 

between the roving accelerometer(s) and the reference(s). The 

resulting data can be animated (or plotted) to examine the 

resulting operating deflection shapes. 

To analyze the operating vibration of the screen, ODS 

measurements were performed (see Fig. 10) using Bruel & 

Kjaer ODS Consultant software. Uniaxial accelerometers were 

used to measure operating vibration at over 400 locations. At 

most of the locations, only the vibration normal to the surface 

was measured, since this is the most useful in terms of noise 

radiation. Three reference accelerometers were used. One 

reference accelerometer was fixed to the left side of the screen 

near the feed end and one was attached to the right side of the 

screen near the discharge end. A third reference accelerometer 

was attached to the H-beam. The data were collected with a 

bandwidth of 1600 Hz. 

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 6 
Figure 10. Researchers roving accelerometers during 

ODS testing. 

Prior to animating the collected ODS data, the sound 

power level radiated by the screen sides was estimated using: 

L = 10 log 2

W 10 v

s ,t

+ 10 log10 S +

 (2) 

10 log10 σ + 146

where L 2 is the sound power level, W v is the surface 

s ,t

averaged mean-squared velocity, Sis the surface area, and σ

is the radiation efficiency. To obtain the surface average 

velocity, the measured acceleration autopower spectra were 

integrated in the frequency domain. In the analysis, each 

portion of the screen was separated into panels between rib 

stiffeners. The perimeter and thickness of the steel panels were 

used to determine the radiation efficiency [12]. 

Figure 11 shows the estimated sound power level from 

each screen side and the feed box. The figure shows the 

surfaces have significant contributions to the sound power level 

in the 100 Hz one-third-octave band through the 800 Hz onethird octave bands. 

60

65

70

75

80

85

90

95

100

80 100 125 160 200 250 315 400 500 630 800 Σ 100 Hz

– 800 Hz

Preferred One-third-octave-band Center Frequency (Hz)

Estimated Lw (dB re 10-12 watts)

LHS

RHS

Feedbox

Figure 11. Estimated sound power level radiated by 

the screen sides and feedbox. 

Previous research identified the 200 Hz 

through 800 Hz frequency range as the dominant frequency 

range for the measured A-weighted sound power level [9]. 

Each of the three surfaces has comparable estimates of radiated 

sound power. In order to significantly reduce the noise in this 

frequency range, all three screen components must be 

addressed. 

A wire frame model of the screen was made in ME’scope 

software (see Fig. 12). 

left side

feedbox

H-beam

Panel 1

Panel 2

Panel 3

Panel 4

Panel 5

Panel 6

Panel 7

Stiffening

tube

Stiffening

tube

Figure 12. ME’scope wireframe model of the screen. 

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 7 

The collected data were imported into 

ME’scope. The data were examined for peaks in the 

acceleration autopower spectra that coincide with frequencies 

of high noise radiation. The operating deflection shapes were 

then examined at these frequencies. To be succinct, only 

selected results will be presented here. For each of these 

figures, the color indicates the amount of motion at each 

location. 

Two of the operating deflection shapes for peaks that are in 

the 100 Hz one-third-octave band are shown in Fig. 13. Each 

shape is dominated by motion along the screen sides. For the 

shape at 92 Hz, the sixth panel along each side exhibits 

dominant motion whereas the shape at 107 Hz is dominated by 

motion at the seventh panel. In each case, the motion of the 

panel is similar to the first mode of vibration of a plate with 

fixed boundary conditions. 
(a) 

(b) 

Figure 13. Operating deflection shapes for (a) 92 Hz 

and (b) 107 Hz. 

Figure 14 shows two operating deflection shapes in the 

200 Hz one-third-octave band. 

(a) 

(b) 

Figure 14. Operating deflection shapes at (a) 182 Hz 

and (b) 204 Hz. 

 The shape at 182 Hz is 

dominated by bending of the back panel of the feedbox. In 

addition, the third and sixth panels on the left side and the sixth 

panel on the right side undergo significant motion. In this case, 

a nodal line exists between the upper and lower portions of the 

panels. The shape at 204 Hz shows the feedbox and the third, 

fifth, sixth, and seventh panels have the dominant motion. The 

motion at the fifth panel is confined to the bottom portion of 

the panel. It is interesting to note that the top of the fifth panel 

is the location where the H-beam connects to the screen sides. 

Two of the operating deflection shapes for frequencies in 

the 250 Hz one-third-octave band are shown in Fig. 15. The 

motion for the shape at 239 Hz is dominated by motion of the 

feedbox. The motion on the back of the feedbox shows five 

regions of high vibration levels. The shape at 253 Hz is 

dominated by motion at the sixth and seventh panels along the 

screen sides. In this case, the sixth and seventh panels exhibit 

vertical nodal lines. The back and inside of the feedbox also 

exhibit significant motion. 

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 8 
(a) 

(b) 

Figure 15. Operating deflection shapes for (a) 239 Hz 

and (b) 253 Hz. 

Figure 16 shows two operating deflection shapes for 

frequencies within the 315 Hz one-third-octave band.

(a) 

(b) 

Figure 16. Operating deflection shapes for (a) 332 Hz 

and (b) 337 Hz. 

 Each of 

these shapes is dominated by motion of the back and sides of 

the feedbox. Each shape also shows significant motion at the 

seventh side panel. However, for these shapes, the motion on 

the screen sides is somewhat more distributed than the other 

shapes presented here. 

After comparing the presented operating deflections 

shapes, it appears that there are some common areas of 

significant motion among the shapes. The back of the feedbox 

exhibits significant motion for all of these shapes except those 

in the 100 Hz one-third-octave band. In addition, some 

combination of the third, fifth, sixth, and seventh panels have 

significant motion for all the presented shapes. Noise controls 

should focus on these key areas in order to produce the largest 

reductions in noise with the fewest changes. 

NOISE CONTROLS FOR THE SCREEN BODY 

The excitation source for the presented operating 

deflections shapes is most likely bearing slap within the 

vibration mechanisms. The bearings for these machines must 

incorporate clearance within the bearings to reduce heat 

generation. This clearance probably results in relative motion 

between the eccentric shafts and the bearing housings of the 

vibration mechanisms. In effect, this provides a source of 

broadband excitation to the structure which excites the 

structural modes of the screen. 

In order to reduce the motion on the screen body, some 

combination of stiffness and damping treatments could be 

applied. Increasing the plate thickness would increase the 

stiffness of the structure and slightly increase the resonant 

frequencies for the structure. This would reduce the amplitude 

of motion for the presented deflection shapes. However, 

increasing the plate thickness substantially would require the 

eccentric force to be increased to maintain the same processing 

capability. This in turn would require more durable bearings. 

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 9 
Application of constrained layer damping to the screen 

sides and feedbox would reduce the amplitude of motion at the 

resonant peaks that are excited during operation. This might 

also make a substantial noise reduction. However, this also may 

be unacceptable to the screen manufacturer due to the 

additional weight for the reasons discussed above. 

Perhaps the best combination of screen modifications 

would be to add stiffeners and damping treatments to selected 

locations based on the operating deflection shapes presented 

above. For example, rib stiffeners could be added to the panels 

on the screen sides and feedbox where the most significant 

motion occurs. Stiffeners could be positioned so that each 

stiffener crosses the areas near the maximum motion for the 

panel in question. In addition, constrained layer damping 

treatments could be applied to these areas to reduce motion. 

It is important to note that adding rib stiffeners can 

increase the radiation efficiency of panels due to the increase in 

the perimeter when a stiffener is added. However, it may be 

possible to increase the modal stiffness so that the reduction in 

vibration outweighs the increase in radiation efficiency. 

CONCLUSIONS 

A 121-microphone phased array was used to perform 

beamforming measurements on a vibrating screen. The 

beamforming results indicate the screen sides and feed box are 

dominant noise sources for the 250 Hz through 800 Hz onethird-octave bands whereas the electric motor appears to be the 

most significant source for the 1 kHz band. Operating 

deflection shape analysis showed several key areas are the 

main locations of high motion for vibration in the 100 Hz, 200 

Hz, 250 Hz, and 315 one-third-octave bands. The back and 

sides of the feedbox and the third, fifth, sixth, and seventh 

panels on the screen sides have significant motion for the 

presented shapes. Noise controls should target these areas in 

order to reduce noise generated by the screen. 

ACKNOWLEDGMENTS 

The authors would like to acknowledge Conn-Weld 

Industries, Inc. for their support of this research. In addition, 

the authors would like to extend their gratitude to Pat 

McElhinney, Lynn Alcorn, and Adam Smith from NIOSH, 

Pittsburgh Research Laboratory for their assistance in 

collecting the data for this work. 

REFERENCES 

[1] National Institute for Occupational Safety and Health, 

“National Occupational Research Agenda,” DHHS (NIOSH) 

Publication No. 96-115 (1996). 

[2] National Institute for Occupational Safety and Health, 

“Preventing occupational hearing loss – a practical guide,” 

Edited by Franks, J, MR Stephenson, and CJ Merry. DHHS 

(NIOSH) Publication 96-110, 91p. (1996) 

[3] JR Franks, Analysis of Audiograms for a Large Cohort of 

Noise-Exposed Miners. Unpublished technical report. NIOSH, 

Cincinnati, OH: (1996) 

[4] Department of Labor, Mine Safety and Health 

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[5] S. Fiscor and J. Lyles, “Prep Plant Population Reflects 

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[8] D. Yantek, P. Jurovcik, and E.R. Bauer. “Noise and 

Vibration Reduction of a Vibrating Screen,” 2005 SME Annual 

Meeting, preprint 05-71, (Society for Mining, Metallurgy, and 

Exploration, Inc., Littleton, CO, 2005). 

[9] D. Yantek, H. Camargo, and R. Matetic, “Application of a 

microphone phased array to identify noise sources on a 

horizontal vibrating screen,” Noise-Con 2008. 

[10] Thomas Mueller (ed.), Aeroacoustic Measurements, 

Springer, 2002. ISBN 3-540-41757-5. 

[11] P. McHargue and M. Richardson. “Operating Deflection 

Shapes from Time Versus Frequency Domain Measurements”, 

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[12] D.A. Bies and C.H. Hansen, Engineering Noise Control 

Theory and Practice. Third Edition. New York: Spon Press, 

2003. 

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unlimited. 

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