A noise control package for vibrating screens1),

1) The findings and conclusions in this report are those of the

author(s) and do not necessarily represent the views of the

National Institute for Occupational Safety and Health.

Mention of any company name, product, or software

does not constitute endorsement by NIOSH.

2)

2) This is the fourth paper published in NCEJ on the special

topic of Mining Noise.

M. Jenae Lowea),

a) NIOSH, 626 Cochrans Mill Rd, P.O. Box 18070,

Pittsburgh, PA 15236 USA; email: MLowe@cdc.gov.

David S. Yantekb),

b) NIOSH, 626 Cochrans Mill Rd, P.O. Box 18070,

Pittsburgh, PA 15236 USA; email: DYantek@cdc.gov.

Junyi Yangc),

c) NIOSH, 626 Cochrans Mill Rd, P.O. Box 18070,

Pittsburgh, PA 15236 USA; email: JYang4@cdc.gov.

Kevin C. Schusterd)

d) United States Steel Corporation, 5706 Ollie St, Pittsburgh,

PA 15207 USA; email: KevinSchuster100@gmail.com

Work reported here was undertaken while at CDC.

and Jessie J. Mechlinge)

e)	 NIOSH, 626 Cochrans Mill Rd, P.O. Box 18070,

Pittsburgh, PA 15236 USA; email: JMechling@cdc.gov

Hearing loss was the second-most common illness reported to the Mine Safety

and Health Administration (MSHA) in 2009. Furthermore, between 2000 and

2010, 30% of all noise-related injury complaints reported to MSHA were for

coal preparation plant employees. Previous National Institute for Occupational

Safety and Health (NIOSH) studies have shown that vibrating screens are key

noise sources to address in order to reduce coal preparation plant noise. In

response, NIOSH researchers have developed a suite of noise controls for vibrating screens consisting of constrained layer damping (CLD) treatments, a tuned

mechanism suspension, an acoustic enclosure, and spring inserts. Laboratory

testing demonstrates that this noise control suite reduces the A-weighted sound

power level of the vibrating screen by 6 dB. To provide a comparison to laboratory

results and prove durability, field testing of two noise controls was performed on a

vibrating screen in a working coal preparation plant. The spring inserts and CLD

treatments were selected due to their ease of installation and practicability. Field

testing of these controls yielded reductions that were comparable to laboratory

results.

1 INTRODUCTION

Of all occupational illnesses reported to the U.S.

Bureau of Labor Statistics in 2004 and 2005, 11%

were due to noise-induced hearing loss (NIHL)1

,

making it one of the most common occupational illnesses in the United States2

. As such, hearing loss

prevention is one of twenty-one Priority Research

Areas listed in the National Institute for Occupational

Safety and Health (NIOSH) National Occupational

Research Agenda3

. Of all industries, mining has the

highest prevalence of hazardous noise exposure, with

76% of employees reporting hazardous noise exposure

on the job4

. Despite leading the nation in reported

hearing protector use4

, miners have the second-highest

incidence of reported hearing difficulty of all occupations—in a National Health Interview Survey that

spanned 1997 to 2003, nearly one-fourth of all miners

reported a hearing difficulty5

. Because hearing conservation programs alone do not solve the problem of

NIHL, it is important to use engineering noise controls

as the first line of defense6

, thus reducing harmful

noise exposure for miners. The Mine Safety and Health

Administration reinforced this message in 1999 by

revising the Health Standard for Occupational Noise

Exposure (30 CFR Part 62), which no longer gives

credit for hearing protection in determining a worker’s

noise dose7

. As a result, the regulation has reemphasized the use of noise controls to reduce worker noise

exposure.

The Office of Mine Safety and Health Research

(OMSHR) Hearing Loss Prevention Branch (HLPB)

develops such noise controls for miners. One area where

the OMSHR HLPB has focused its research is within

coal preparation plants. According to MSHA data,

30% of all noise-related injury complaints since the year

2000 have involved preparation plant employees. In

2009, there were 37,405 employees at mine facilities

with preparation plants and 8,343 workers who were
categorized as preparation plant employees at these

facilities8

.

A recent NIOSH study shows that 43.5% of

employees within preparation plants are overexposed

to noise. Furthermore, the study found that not only

were vibrating screens one of the loudest pieces of

equipment at the preparation plants, they were also

the most numerous, thus making vibrating screens a

key noise source to address9

. Vibrating screens are a

major noise problem in coal preparation plants because

screens are used extensively in the plants, are usually

located in high traffic areas, and can generate high

sound levels10. Coal preparation plants use horizontal

vibrating screens more frequently because their smaller

height allows for shorter floors and less overall building volume in a preparation plant. Typically the sound

levels around these machines range from 90 to 95 dB

during clean bituminous coal processing, and from 95

to 100 dB during refuse and anthracite processing11.

To reduce noise exposure within the preparation plants,

NIOSH developed engineering noise controls for

vibrating screens in a partnership with Conn-Weld

Industries, Inc.

A horizontal vibrating screen (Fig. 1) is a large

machine used to process coal.

a 

b 

suspension 

springs 

electric 

motor 

rib 

feedbox 

H-beam 

round 

cross-tube 

A-frame 

A-frame rail 

FEED 

END 

electric 

motor mechanism 

housings 

belt guard 

screening 

deck 

H-beam 

A-frame 

A-frame rail 

DISCHARGE 

END 

left side 

screen body 

Fig. 1—A horizontal vibrating screen used to process coal viewed from (a) feed end and

(b) discharge end.

The screen body has four

sides made of steel plates with a bottom screening surface—also known as a screen deck—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. 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 by an electric

motor.

The screen is designed in such a way that it vibrates

on roughly a 45-degree angle from horizontal. In
operation, coal flows into the feed end of the screen

from a delivery chute. As the screen vibrates, the material moves along the deck and under a water spray that

rinses 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 of the discharge end of the

screen to continue with further processing.

In general, there are two primary mechanisms of

noise generation on a vibrating screen: material flow

noise generated from either impacts between the material and the screen surfaces or between individual

pieces of the material itself, and noise generated by vibration of the screen through its operation. As Ungar et

al. notes, “In general, coal impact noise predominates

in screens handling coarse coal, and drive mechanism

noise predominates in screens handling finer coal12,”

and the same relationship between material feed size

and noise sources in screens has been found for the

screens used in the aggregate industry13. The solution

to material flow noise is well understood, as the majority of prior efforts to reduce vibrating screen noise have

focused on reducing material flow noise through the use

of rubber screen decks11,12,14–25, various screen deck

isolation or attachment techniques26,27, and chute

liners11,12,14–18,28. However, less work has been done

to reduce the noise from the screen itself. The most

common solution to this problem has been to isolate

the screen from its support structure in the plant

through the use of rubber isolation mounts or air bags

instead of springs11,20,23,24,29.

Previous NIOSH studies of a horizontal vibrating

screen used to process clean coal at a preparation plant

indicated that noise due to vibration of the screen itself

was the dominant noise source, whereas noise from material flow was less significant30. This is due to the

smaller material feed size for these particular screens.

Further research in the lab, using the same make and

model of screen found in the aforementioned preparation plant, showed that most of the noise due to screen

vibration is radiated by the screen body and the mechanism housings31,32. One of these studies showed that

the suspension springs were also one of the three main

noise sources of the screen31. Therefore, it was not possible to use the traditional noise controls for reducing

material flow noise, although spring alternatives were

a possibility. Replacing the springs with rubber isolators was attempted, but the results were unfavorable

due to an increase in building vibration30.

As a consequence, a unique noise control solution

was necessary for the screens under study in this project.

Prior noise source identification studies in the lab

showed that screen body noise is the main noise source

below about 1 kHz, while mechanism housing noise is

the primary source above 1 kHz. The sound energy at

frequencies below 1 kHz accounts for about 80% of

the overall A-weighted sound power level. In addition,

operating deflection shape analysis revealed significant

response on the screen sides and feedbox33. As mentioned previously, the suspension springs were also

found to be one of the three main noise sources31. As

a result, noise control efforts were split between developing noise controls to reduce noise above 1 kHz from

the mechanism housings, below 1 kHz from the screen

body, and from spring chatter. NIOSH developed a final

suite of controls which included spring inserts, constrained layer damping treatments34, a tuned mechanism

suspension, and an acoustic enclosure34,35.

Other researchers have attempted some novel solutions for controlling screen noise for the sources listed

above. Several researchers have recommended treatment of the screen body panels and cross beams with

free-layer or constrained-layer damping11,16,19–24,36.

This has been attempted in previous NIOSH studies30

but would be cumbersome to manufacture. The use of

constrained layer damping on the mechanism housings

proved successful34, as it did for Wu et al.22. Stiffening

the body structure by adding a center divider plate, such

as in Ref. 24, is also an effective method to reduce

screen body noise.

To address the excitation noise of the screen, which

includes the motor, eccentric mechanisms, and drive

belt, several different designs have been developed.

Jakobs37 invented a system of multiple tuned excitation

rods connected to the screen surface which produced

quieter operation, but there was no data on the noise

reduction achieved. Others have created a resilient

connection between the mechanism gears and the

eccentric weights38, or developed multi-part flexible

gears to reduce noise20,39,40. Tighter gear and bearing

tolerances have also been employed11,25, and a number

of researchers have installed an isolator ring around

the bearings24,41,42. Hollyfield and Jackson developed

a tuned suspension system for exciting the screen deck

coupled with motor isolation43, and motor isolation

alone was employed by Wang25. Two other research

teams have isolated the mechanism housings from their

attachment plates through the use of various elastomeric

devices21,25,44. As an evolution of this idea, the authors

modeled45 and further developed a tuned mechanism

suspension system which also eliminates high-motion

transients during startup and shutdown. Complete

mechanism redesigns have been accomplished to lower

sound levels20,24,36 as well.

Enclosures have also been commonly recommended

to reduce screen excitation noise. Most of the enclosures

employed have been for the entire screen11,14–17,19,28, or

operator booths have been recommended16,17,19. While
the idea of an enclosure for the mechanism housings has

been presented in the past11, experimental results for

this idea have only been presented by Zhu et al.20,39

and the authors34,35. The work by Zhu et al. in Refs. 20

and 39 provides little information on the enclosure

design, nor does he address the problem of cooling

the enclosure as is presented in this work.

As mentioned above, the common solution to reducing spring noise is to use rubber isolators. Alternate

solutions have been to redesign the spring supports to

use multiple lower-stiffness springs24, and to isolate

the spring from its mounting pad with Bakelite plates25

or resilient pads11. The spring insert design to separate

the contacting spring coils which is presented herein is

singular in the literature.

This work is unique in that the noise controls’ effectiveness has been determined by measuring sound

power in the lab per ISO 3743-246. The great majority

of the prior work measured sound pressure without

specifying the microphone locations or measurement

method. In addition, the authors present a unique design

of a noise control for spring chatter, the development of

an improved isolation system for the mechanism

housings, and the development of a mechanism housing

enclosure that also provides adequate cooling performance. This paper discusses the influence of each noise

control in the final package of spring inserts, constrained layer damping treatments34, a tuned mechanism

suspension, and an acoustic enclosure34,35. Finally, confirmation of the laboratory results for two of the noise

controls in a production environment is presented.

2 EXPERIMENTAL SETUP

2.1 Sound Measurements

All measurements were taken on a Conn-Weld

2.44-m x 4.88-m (8-ft x 16-ft) horizontal vibrating

screen with dual G-Master 1000 vibration mechanisms,

as shown in Fig. 1. The screen was placed in the center

of the OMSHR National Voluntary Laboratory Accreditation Program (NVLAP)-accredited reverberation

chamber47 and rested on the floor with wooden wedges

driven under the frame rails to prevent the screen from

rocking during operation. The belt guard was removed

to avoid measurement inconsistencies from rattling

observed in previous testing34. The screen was run

without coal or water feed, since previous studies

had shown that coal and water flow noise was not a

significant contributor to sound levels in the plant environment30. Sound power tests were performed per ISO

3743-246. Data were simultaneously recorded from 15

microphones arranged in a random distribution47. Three

30-second measurements were taken for each test configuration and logarithmically averaged.

2.2 Temperature Measurements

Temperature measurements were made for those

noise controls installed on the vibrating screen that had

an effect on mechanism temperatures. To avoid mechanical failure of the bearings or other components, the

screen manufacturer, Conn-Weld, recommends that the

mechanisms run at a maximum temperature of 82.2°C

(180°F) for standard oil and an absolute maximum temperature of 90.6°C (195°F) for high-temperature oil

when measured at the surface of the housings. Avariety

of cooling methods were attempted to keep the operating temperature within acceptable limits after the noise

controls were installed.

Five type J thermocouples were connected to the

mechanism housings with thermally conductive cement

(Fig. 2).

a 

b 

Fig. 2—Thermocouple locations on the

vibrating screen mechanism housings.

(a) Left mechanism locations (b) Right

mechanism locations.

A sixth thermocouple measured ambient air

temperature. The thermocouples were connected to a

DATAQ DI-715B data acquisition device, which

recorded the data to a laptop. Temperatures were

recorded until steady state was reached, until it was
apparent that a cooling method was unsuccessful, or

until the temperature reached 82.2–90.6°C (180–

195°F).

3 BASELINE MEASUREMENTS

Baseline measurements of the vibrating screen reveal

a bimodal distribution of sound energy with an overall

A-weighted sound power level of 100.5 dB, as shown

in Fig. 3.

 

100

125

160

200

250

315

400

500

630

800

1,000

2,000

4,000

5,000

8,000

10,000

Overall 

1,250

1,600

2,500

3,150 

6,300 

110 

Baseline w/o CLD on Mechanism Housings, w / Spring Chatter 100.5

A-weighted Sound Power Level (dB) 100 

90 

80 

70 

60 

50 

1/3-Octave-Band Center Frequency (Hz) 

Fig. 3—1/3-Octave-band plot of A-weighted sound power levels of the unmodified Conn-Weld

8′ x 16′ horizontal vibrating screen.

Two humps appear in the data with the first

centered at the 315 Hz 1/3-octave band, and the second

centered at the 2 kHz 1/3-octave band. The distributions

from the humps meet at 1 kHz. As mentioned previously, prior studies showed that the frequencies below

1 kHz are attributable to screen body noise, while the

frequencies above 1 kHz are due to mechanism housing

noise31,32. A package of engineering noise controls was

developed to address these noise sources, as well as

spring chatter noise.

4 NOISE CONTROL DESIGN

A suite of noise controls was designed to address the

different noise sources discovered on the vibrating

screen. The spring inserts were designed to eliminate

spring chatter as a noise source and to ensure testing repeatability because the level of spring chatter was

highly variable. The constrained layer damping treatments and acoustic enclosure were applied to the mechanism housings to reduce their contribution to the

frequencies above 1 kHz. The tuned mechanism suspension was designed to reduce screen body noise below 1 kHz by decoupling the mechanical energy of

the mechanisms from the noise-radiating structure of

the screen body.

4.1 Spring Insert Evolution

Spring inserts were designed to eliminate spring

chatter noise by preventing the last full coils and the

cut and ground coils on each of the spring ends from

coming into contact. An earlier prototype of the inserts,

tested in the lab, was a curved slab of 70 durometer

natural rubber with tapered grooves to seat the spring

coils. The final inserts developed for field testing were

composed of 70 durometer natural rubber bonded to a

6.4-mm-thick (1/4-inch-thick) curved steel shell along

the outer perimeter. The steel shell was intended to provide a striking surface for hammering the spring inserts

into place between the spring coils. The rubber portions

of the inserts were tapered to match the contours of the

springs for better contact, and had a lip to prevent them

from sliding out of place during screen operation, as

shown in Fig. 4. Spring inserts were designed for both

the inner and outer springs of the vibrating screen, so

that all springs on the screen were prevented from

chattering.

4.2 Constrained Layer Damping

Constrained layer damping (CLD) is commonly used

to treat a vibrating surface that is generating noise. It

consists of a layer of flexible damping material applied

to the vibrating surface, with the other side of the damping layer constrained by a more rigid material such as

steel. Vibration energy from the base layer is
transformed into a shear deformation within the damping layer, which reduces the noise radiated by the system while generating a small amount of heat.

Fig. 4—Examples of small (left) and large

(right) spring inserts. Note that the

metal backing plate has been removed

from the small spring insert.

For this application, a thin sheet of 80 durometer,

0.64-mm-thick (0.025-inch-thick) elastomeric damping

material was bonded on one side to the flat faces on the

front, top, and back of each mechanism housing using

epoxy. These layers of damping material were then

constrained by being bonded to 6.4-mm-thick (1/4­

inch-thick) steel plates. Figure 5 shows constrained

layer damping treatments on the top and front face of

the mechanism housings.

Front face treatments 

Top face treatments 

Fig. 5—Constrained layer damping treatments

applied to the top and front faces of

the mechanism housings.

To ensure a good bond between the housings, damping material, and constraining

plates, the paint from the housings was removed using a

grinder and the constraining plates were sandblasted and

cleaned with isopropyl alcohol prior to applying the

treatments.

4.3 Tuned Mechanism Suspension

A tuned mechanism suspension (TUMS) was

developed to selectively transmit forces at the mechanism rotation speed of 900 RPM, while attenuating

forces at higher frequencies that resulted from the gears

and bearings. The mechanism suspension is “tuned” by

adjusting the spring rates of the mechanism suspension

to shift the rigid body resonant frequencies of the screen

with the added mechanism suspension to either a higher

or a lower frequency. In a previous publication45, a

model of the screen and TUMS was examined within

ANSYS to determine the approximate spring rates that

would be needed. The goal was to have the rigid body

frequencies fall above the operating speed of the screen

and below a target value for isolation at 100 Hz45.

The TUMS consisted of a mounting plate, or “raft,”

for the mechanisms, a modified H-beam, a series of

gussets to support the added material to the H-beam,

and a series of two-piece vibration isolators between

the raft and the H-beam, as shown in Fig. 6. In combination, these elements decouple the mechanism housings

and the H-beam. The vibration mechanisms were

mounted on top of a large steel plate, or raft, which

was in turn mounted to the modified H-beam through

the vibration isolators. The goal of adding a mechanism

suspension is to separate the primary source of mechanical energy from the noise-radiating structure, but without degrading the performance of the screen.

The H-beam was modified by adding a “cap,” as

illustrated in Fig. 7, over top of the H-beam to create

a wider platform for mounting the tuned mechanism

suspension in order to avoid problems with the pitching modes involving the mechanism housings being

close to the operating frequency of the screen45. The

resulting modified H-beam was 7.3 cm (2–7/8 inch)

thick and 64.1 cm (25.25 inch) wide. Eight 1.9-cmthick (3/4-inch-thick) gussets supported the H-beam

cap and attached it to the H-beam. Twenty-four natural

rubber two-piece vibration isolators, as shown in Fig. 8,

were used to separate the raft from the modified

H-beam. The isolators used an external steel structure

to prevent the rubber from being cut by the H-beam cap

or mounting hardware once the mount was compressed.

4.4 Acoustic Enclosure

An acoustic enclosure was designed that surrounds

both mechanism housings and is attached to the

H-beam34,35. The motor and drive belts were not

enclosed due to space constraints in coal preparation

plants. Additionally, they were not found to be
significant contributors to noise32. We created a modular

panel-on-frame design with components that weigh less

than 222 N (50 lb.) and can be hand-carried and

installed on any Conn-Weld 2.44-m x 4.88-m (8-ft

16-ft) horizontal vibrating screen. The modular nature

x

 

of this design accommodates different numbers and

spacing of vibration mechanism housings and the design can be modified to adapt to different sizes and types

of vibrating screens.

Mechanisms 

Vibration 

Isolators 

Raft 

Modified H-Beam 

Gussets 

Motor 

a b 

Fig. 6—Model of TUMS system (a) and picture of TUMS system installed on the vibrating

screen (b).

As-built H-Beam 

H-Beam cap 

Raft 

Width increase 

Fig. 7—Illustration showing placement of the H-beam cap, vibration isolators, and raft.

As shown in Fig. 9, the enclosure consists of a steel

frame, which provides a stiff structure, and panels that

enclose the noise source. The frame is made up of a

series of sub-frames that can be bolted together to

make a larger or smaller enclosure as needed. The

panels act as a barrier to contain the noise, while a

layer of acoustic foam on the inside of the enclosure

prevents the reverberant build-up of noise. The bolton panels can be easily removed to reach a fill port,
drain plug, or bearing cover, as shown in Fig. 9.

Inner steel sleeve 

Bonded outer steel sleeve 

Fig. 8—Illustration of Tuned Mechanism

Suspension vibration isolators.

 

 

a b 

2” X 5/8” U-channel, 

1/4” thick 

2” X 2” angle, 

3/16” thick 

Bolts 

Weld nuts 

Fig. 9—Steel frame for the second enclosure showing (a) one frame section and (b) the entire

frame assembly with most panels installed on a vibrating screen.

An

entire frame section with the panels attached can be

removed to change a mechanism. Cooling ducts bolt

onto the side panels separately and can be reconfigured as necessary, as shown in Fig. 10.

Enclosure construction involved a variety of materials and screen modifications. The frame was composed

of steel angle stock and U-channel. It was isolated from

the H-beam using strips of 57 durometer natural rubber.

Panels were fabricated with 3.2-mm-thick (1/8-inchthick) DynalamTM damped steel (a CLD steel). The

insides of the panels were lined with 25.4-mm-thick

(1-inch-thick) PolydampW acoustic foam. A boss made

of 57 durometer natural rubber isolated the right panel

from the bearing cover plate.

Because the enclosure consists of damped steel

panels and is lined with acoustic foam, it makes a good

thermal insulator. As a consequence, temperatures at

the mechanism housings increase with the enclosure installed. A variety of methods to cool the enclosure were

explored, including cooling the entire enclosure with

compressed air, adding an 80.2-L/s (170-CFM) push–

pull fan system at the inlet and exhaust ducts of the enclosure, exhausting the enclosure air with a 330.3-L/s

(700-CFM) fan attached to a modified enclosure duct,

and running this same fan at 75% airflow.

5EXPERIMENTAL RESULTS AND

DISCUSSION

5.1 Spring Inserts

Early in the project, we observed spring chatter both

in the lab and in the field. Later, through a beamforming

study, it was shown that the suspension springs were

one of the three main noise sources of the screen31. It

was also shown that the spring chatter resulted from

impacts between the cut and ground coils on the spring

ends and the coils that they contact48. When separating

the contact between these coils with wooden wedges or

pieces of closed-cell foam, the spring chatter was eliminated and noise was reduced in the 1/3-octave-bands

from 3150 Hz and above49. Because spring chatter is

intermittent and variable, the maximum potential reduction from the spring inserts is not known. However,

spring inserts have been used consistently throughout

all testing to eliminate a known noise source, which

also eliminates a source of variability when testing the

other noise controls.
Panels 

Duct 

Fig. 10—Enclosure installed on vibrating

screen.

5.2 CLD and Spring Inserts

After eliminating the spring chatter with the spring

inserts, the next controls installed were the constrained

layer damping treatments. As shown in Fig. 11, the

greatest reductions were achieved above 1 kHz, which

is the frequency range that the spring inserts and CLD

plates were designed to address, and where the mechanism housings are the most significant noise source.

Overall, the combination of both noise controls reduced

the sound power level by 2.7 dB.

 

 

100

125

160

200

250

315

400

500

630

800

1,000

2,000

4,000

5,000

8,000

10,000

Overall 

1,250

1,600

2,500

3,150 

6,300 

100.5 

97.8 Baseline w/o CLD on Mechanism Housings, w/ Spring Chatter 

Constrained Layer Damping + Spring Inserts Installed 

50 

60 

70 

80 

90 

100 

110 

A-weighted Sound Power Level (dB) 

1/3-Octave-Band Center Frequency (Hz) 

Fig. 11—1/3-Octave-band plot of A-weighted sound power levels of the baseline screen vs. the

screen with constrained layer damping plates and spring inserts installed.

5.3 Tuned Mechanism Suspension

The TUMS system was installed on the screen next,

in addition to the spring inserts and CLD plates. With

the original 67 durometer vibration isolators, the bolt

heads impacted against the modified H-beam, causing

the noise to increase during operation. The sound power

level increase was primarily due to a lack of clearance

between the mechanism bolt heads and the modified

H-beam.

In addition, the first set of vibration isolators was not

stiff enough to hold all of the components together

during operation. Examining the operation of the

screen with a stroboscope revealed that the tops of

the isolators were losing contact with the top of the

modified H-beam. When the two-piece mounts are installed, the mounts are pre-compressed so that both the

top and bottom of the mounts undergo static deflection.

Because the top half of the isolator is slightly stiffer than

the bottom half of the isolator, the static deflection of the

top half of the isolator is slightly less than the static deflection of the bottom half of the isolator. The relative

motion between the raft and the modified H-beam

exceeded the static deflection of the top half of the isolator. This caused the top half of the isolator to become

unloaded and allowed separation of the isolator from

the top of the H-beam.

Several changes were made to address these problems. The height of the isolator tops was increased to

3.2 cm (1.25 inch) to reduce the spring rate of the top

half of the isolator relative to the bottom half. This ensured that the top half would have enough static deflection to prevent the components from separating under

operation. The durometer of the mounts was also increased to increase the mount stiffness. Finally, various

spacers were used to increase the preload on the
vibration isolators and the clearance of the bolt heads

from the modified H-beam, as shown in Fig. 12.

Spacers 

Fig. 12—Spacers installed above and below the top portion of the vibration isolators.

These

changes ultimately resulted in the “tuning” of the system that was desired.

Sound power level reductions were achieved with

these design changes to increase preload, clearance,

and mount stiffness. Table 1 is a summary of the successful iterations of TUMS development.

Table 1—A-weighted sound power levels for

successful tuned mechanism suspension

(TUMS) iterations. The baseline configuration included constrained layer

damping on the mechanism housings

and spring inserts.

Durometer Top

Washer

Bottom

Washer

A-weighted

Sound Power

Level (dB)

N/A

(Baseline)

N/A

(Baseline)

N/A

(Baseline)

97.8

80 1/4 in. 1/8 in. 97.0

80 1/4 in. None 96.4

90 1/8 in. 1/8 in. 96.8

90 1/8 in. None 96.4

The greatest

reductions were accomplished by both the 80 durometer mounts with a 6.4-mm (1/4-inch) spacer on top

and none on the bottom and the 90 durometer mounts

with a 3.2-mm (1/8-inch) spacer on top and none on

the bottom—each of which yielded a 1.4-dB improvement versus baseline.

Performance for individual 1/3-octave bands varied

with durometer and preload, as shown in Fig. 13. The

80 durometer mounts had better performance than the

90 durometer version in the 125–315 Hz frequency

range. All mounts increased the sound power level

compared to the baseline in the 2–10 kHz range, but

the combination of 80 durometer mounts with the

6.4-mm (1/4-inch) top spacers and the 3.2-mm (1/8­

inch) bottom spacers showed the highest increase in

this range.

Another concern in mount selection was heating of

the mounts under operation, which caused the sound

power level to increase over time. Previous measurements of failed mount configurations showed a trend

of increasing sound power as the testing progressed.

Thermocouples were installed on one of the earlier

mount prototypes to observe how they were heating

up. The results showed that the mount temperature

increased as the vibrating screen continued to run. When

the mount temperature reached 93.3°C (200°F), the test

was aborted per the isolator manufacturer’s recommendation that the mount temperature should not exceed

93.3°C (200°F). Due to internal damping, the isolators

heated up during operation as the rubber stretched and

compressed. As the isolators increased in temperature,

their stiffness was reduced. The design of the TUMS is

such that the vibration mode characterized by relative

motion between the raft and the modified H-beam

occurs at a frequency much higher than the 15-Hz operating frequency of the screen. As the mount temperatures increase and the mount spring rates decrease, the

frequency of this vibration mode shifts closer to the

operating speed of the screen. This results in an increase

in motion of the mechanisms, which in turn causes the

impact forces within the gears and bearings to increase.

The temperature increase is a probable cause for the

increase in noise from the early prototypes.

While the 80 durometer mounts with spacers on top

and bottom showed the greatest stability in terms of

sound power levels over time, their overall poor performance in sound power level reductions eliminated them

as a candidate. Removing the bottom spacers from the

80 durometer mounts caused the mounts to increase

in temperature and exhibit the same trend of increasing

sound power levels over time. However, the 90
durometer mounts showed a trend of decreasing sound

power levels with time, suggesting that the ideal mount

stiffness is between 80 and 90 durometer if the stiffness

could remain constant with temperature. For durability

concerns, the 90 durometer mount configuration with

minimum preload was selected. This configuration

yields a sound power level reduction from the CLD

and spring insert combination of 1.4 dB (Table 1).

100

125

160

200

250

315

400

500

630

800

1,000

2,000

4,000

5,000

8,000

10,000 

1,250

1,600

2,500

3,150 

6,300 

60 

65 

70 

75 

80 

85 

90 

95 

CLD with Spring Inserts 

TUMS 80 Durometer with 1/4" top spacers, 1/8" 

bottom spacers 

TUMS 80 Duro no 1/8" bottom spacers 

TUMS 90 Durometer, 1/8" top and bottom washers 

TUMS 90 Durometer, 1/8" top washers only 

A-weighted Sound Power (dB)

Frequency (Hz) 

Fig. 13—A-weighted sound power for baseline vs. four iterations of the tuned mechanism

suspension (TUMS).

 

 

 

 

 

100

125

160

200

250

315

400

500

630

800

1,000

2,000

4,000

5,000

8,000

10,000 

1,250

1,600

2,500

3,150 

6,300 

60 

65 

70 

75 

80 

85 

90 

95 

Baseline w/o CLD on Mechanism Housings, w/ Spring 

Chatter, 100.5 dB(A) 

Constrained Layer Damping + Spring Inserts Installed, 

97.8 dB(A) 

CLD, Spring Inserts, and Tuned Mechanism Suspension, 

96.4 dB(A) 

CLD, Spring Inserts, TUMS, and Enclosure, 94.9 dB(A) 

A-weighted Sound Power Level (dB) 

1/3-Octave-Band Center Frequency (Hz) 

Fig. 14—A-weighted sound power comparison for all configurations tested.

5.4 Acoustic Enclosure

Finally, an acoustic enclosure was added to the noise

control package. Originally, the enclosure had been installed directly onto the H-beam in previous testing34,35,

but due to installation of the TUMS the enclosure frame

had to be modified so it could be welded directly onto

the TUMS “raft.” Figure 14 shows that the enclosure
reduces overall A-weighted sound power levels by

1.5 dB from the configuration with CLD, spring inserts,

and the TUMS. The reductions are seen from the 315 Hz

1/3-octave-band and higher, with some degradation in

the 1/3-octave-bands below 315 Hz. This decrease in

performance at these frequencies was not observed

when the enclosure was originally tested and compared

to baseline34,35, and Fig. 14 shows that the enclosure is

an improvement over baseline even in this lower frequency range. The extra surface area due to the addition

of the enclosure could radiate more noise than the

TUMS configuration alone, and may be an attribute to

investigate further.

As shown in Fig. 14, the entire noise control package

reduced the overall A-weighted sound power level by

5.6 dB, which is nearly a 75% reduction in sound

power. All 1/3-octave-band levels are reduced with the

exception of a small increase in the 160 Hz 1/3-octave

band. The greatest reductions are seen at the 250 Hz

1/3-octave band and above.

As a consequence of adding the acoustic enclosure,

temperature data were taken as discussed previously to

determine if the enclosure would cause premature damage to the bearings due to excessive heat. For all tests

the thermocouple located at position 4 (see Fig. 2)

recorded the highest temperature. Since maximum

temperature is the main concern, all plots show the

temperature data recorded from position 4. Initial

measurements indicated that the system reached the

critical temperature of 90.6°C (195°F) in less than

five hours with the enclosure installed. This result

compares unfavorably to the baseline, where a maximum temperature of 79.4°C (175°F) was reached after

12 hours.

Next, several cooling methods for the enclosure were

evaluated. Because the free convection provided by the

enclosure ducts was not sufficient for cooling, the enclosure was cooled using shop air, because that may

be more convenient for the end user. An air line was

run to the enclosure duct and an adapter was made for

the duct-to-air line connection. However, the cooling

performance was only slightly better than free convection, so this option was not pursued further. To create

a push–pull system of airflow, 80.2-L/s (170-CFM) fans

were then installed at the inlet and exhaust ducts. This

showed moderate success (10 degrees cooler than the

enclosure-only configuration after 4 hours), but because

it was clear that the cooling was not sufficient, the test

was aborted.

The final cooling configuration utilized a 330.3-L/s

(700-CFM) exhaust fan. Additionally, the enclosure

duct was modified to connect the top of the duct to a remotely located exhaust fan. As shown in Fig. 15, after

10.5 hours, this configuration reached 84.4°C (184°F)

running at full flow with an initial ambient temperature

of 23.3°C (74°F), which is sufficient cooling for a production day in a prep plant.

80 

90 

20 

30 

40 

50 

60 

70 

0 100 200 300 400 500 600 700 

Temperature (°C) 

Time elapsed (minutes) 

Baseline 

Enclosure 

Enclosure, 330.3 L/s exhaust fan, modified duct 

Enclosure, 330.3 L/s exhaust fan, 75% flow, modified duct 

Fig. 15—Comparison of 330.30-L/s (700-CFM) exhaust fan configurations with baseline and

enclosure only configurations.

Running at 75% of flow, the

critical temperature of 90.6°C (195°F) is reached in 10

hours and 50 minutes with an initial ambient temperature of 26.7°C (80°F).

The 330.3-L/s (700-CFM) exhaust fan for the enclosure provides adequate cooling. However, further
efficiencies could be gained by optimizing the cooling

system to improve interior air flow. This may result in

reduced airflow requirements for the fan. For example,

the current duct configuration could be modified to reduce turbulence from bends and sharp corners. This

has the potential to eliminate airflow losses in the duct.

Further, during the testing described above for the

80.2-L/s (170-CFM) fans, the coolest mechanism

housing temperature location at point 3 (Fig. 2) was

predicted to reach the critical temperature after 9 hours,

while the hottest housing location (point 4) was predicted to reach the critical temperature after 5.5 hours.

It may be possible to reduce airflow requirements of

the fan by better mixing of the interior air in the enclosure, thus improving cooling efficiency.

6 FIELD TESTING

In order to prove the effectiveness of the noise

control concepts that NIOSH had developed and the

durability of the materials in the designs, NIOSH performed field testing at a nearby coal preparation plant

in northern WV. If the noise controls in this testing

were proven to be successful in a harsh production

environment, then our industry partner would be able

to incorporate these concepts into future designs. The

preparation plant leadership expressed interest in trying two of the four laboratory-tested noise controls:

the constrained layer damping (CLD) treatments and

the spring inserts. These controls were selected for

their ease of installation and practicability. The other

noise controls require significant modification to the

screen in order to be installed.

6.1 Experimental Setup

6.1.1 Plant setup and methodology

NIOSH researchers selected screen #169 (Fig. 16) on

the fourth floor of the preparation plant because it was

the same size and model of screen used in the Pittsburgh

Office of Mine Safety and Health Research (OMSHR)

lab for developmental research.

Microphone 

locations 

Fig. 16—Plan view of microphone positions and barrier-absorber curtain location surrounding

screen #169.

This screen is one in a

bank of eight 2.44-m x 4.88-m (8-ft x 16-ft) horizontal

vibrating screens used to rinse the coal and remove the

liquid and fine coal particles as the coal moves along

the screen. Ideally, when evaluating noise controls, the

machine under study would be measured in isolation

so that only the noise of the machine in question

would be captured. Otherwise, noise from surrounding

equipment could affect the sound level that was measured. Without already knowing the sound levels that

one machine generates, it would not be possible to determine the true sound level of the machine under test.

However, it was not possible to study a machine in

isolation at the preparation plant because of the control

system in the plant. The plant is set up such that the

bank of eight screens is split into two groups of four

screens each, called Side 1 and Side 2. Side 1 or Side

2 may have the coal feed to their screens shut off, one

screen in either Side 1 or Side 2 may be entirely shut

down while the others continue to vibrate and process

coal, and one side at a time may be entirely shut down.
However, due to the required downtime involved and

production needs, shutting down an entire side of the

plant was not feasible. Thus, it was not possible to run

only one screen with coal while shutting down all of

the other screens.

Given the above constraint, in order to determine the

effect of the noise controls, the influence from the

surrounding machines had to be minimized. To accomplish this, the screen under test was isolated from its

surroundings through the use of quilted acoustic

barrier-absorber material to block and absorb a portion

of the surrounding noise. To determine the effectiveness

of the barrier-absorber material at reducing background

noise, measurements needed to be taken before and after

installation to determine the background noise reduction. Background noise was considered to be the noise

due to all other machines in the plant except for screen

#169. If background noise were reduced enough to minimize its effect on the overall levels, sound level measurements could then be made with the package of

noise controls in place to determine their effect on the

vibrating screen.

6.1.2 Data acquisition setup

Six microphones were placed around the vibrating

screen along the feed end and the left side to match

previous testing conducted at the preparation plant30.

The microphones were located 0.61 m (2 ft) from a

reference box surrounding the test screen, which is the

metal perimeter placed around the screen and set into

the plant floor. All microphones were positioned at a

height of 1.52 m (5 ft) to approximate the ear height of

an average employee. Figures 16 and 17 illustrate the

microphone placement at the preparation plant.

Microphones 

Fig. 17—Microphone positions 3–6 on the left

side of the screen.

Fig. 18—Example of seal around pipes

through the barrier wall.

Three recordings of 30 seconds each were taken for

each test configuration. The data were stored by a data

acquisition system for post-processing. The same

microphone locations and data acquisition system position were used for all measurements.

6.1.3 Barrier installation

Quilted barrier-absorber material was installed

around the perimeter of the screen as shown in Fig. 16.

The sheets of barrier material, which were edged with

VelcroW strips to allow for a robust seal between sheets,

formed a continuous perimeter around the vibrating

screen without gaps between the sheets. Care was taken

to ensure a good seal around pipes and other protrusions through the barrier, as shown in Fig. 18, to minimize noise leakage from outside of the barrier. The

barrier extended from the ceiling to the floor, as shown

in Fig. 19.

6.2 Results and Discussion

For each test, the average A-weighted 1/3-octaveband sound levels were calculated. Figure 20 shows a

comparison with and without the barrier material with

all of the screens running except for screen #169. The

results show that the barrier material reduced the background noise by 9 dB. With the reduction in background

noise, the net result is that the noise floor is 7 dB below

the level of the screen and the background noise combined, yielding a background noise correction factor of

0.8 dB.

When the noise control package was installed, spring

chatter noise had not been observed for the baseline.

However, the spring inserts still needed to be evaluated

for durability and ease of installation. Although the final

prototypes of the spring inserts were originally designed

to have an outer steel shell to aid in installation, the steel

shell was found to be both problematic and unnecessary.
Fig. 19—Barrier wall surrounding the screen.

100

125

160

200

250

315

400

500

630

800

1,000

2,000

4,000

5,000

8,000

10,000

Overall 

1,250

1,600

2,500

3,150 

6,300 

95 

Baseline 

Barrier Installed 

94 

85

Sound Pressure Level, dB(A) 

90 

85 

80 

75 

70 

65 

60 

55 

50 

1/3-Octave-Band Center Frequency (Hz) 

Fig. 20—1/3-Octave-band sound pressure level comparison with screen #169 shut down, with and

without the barrier-absorber material installed.

First of all, prior to the field testing, it was determined

that the prototype inserts for the inner spring did not

fit within the outer spring, so all of the inner spring

inserts were modified by removing the outer steel shell

(Fig. 4). Later, during installation at the preparation

plant, we found that hammering on the steel shell to install the spring inserts was not as effective as prying the

end coils apart and placing the inserts in the proper position. Figure 21 shows the spring inserts in place with

the springs installed on the vibrating screen.

To determine the effect of the noise controls on overall sound levels, it is necessary to compare the data

from all of the screens running with the barrier material

in place versus the same conditions with the noise controls installed. This makes it possible to determine the

reduction in sound level due to the noise controls. A

comparison of these conditions in Fig. 22 shows that

the overall sound level was reduced by 1 dB, with

reductions across most of the frequency range.

A prior study conducted in the lab showed that the

CLD plates alone reduced the overall sound power

levels by 1 dB34. We were expecting at least a 1-dB

reduction with some additional improvement from

the spring inserts. However, given that no spring chatter

was observed on screen #169 before the installation of

the noise controls, the spring inserts are not likely to

have affected the noise generated on this particular

screen. Therefore, the results achieved in the plant are

in line with expectations when the test conditions and

past laboratory results are considered.

7 CONCLUSIONS

NIOSH researchers at the Office of Mine Safety and

Health Research developed four noise controls for a

horizontal vibrating screen used in coal preparation

plants. Spring inserts combined with constrained layer

damping treatments on the mechanism housings reduced overall A-weighted sound power levels by

2.7 dB. The addition of a tuned mechanism suspension
reduced the overall A-weighted sound power levels by

an additional 1.4 dB. Finally, adding an enclosure to

the noise control package improved the results by

another 1.5 dB, which resulted in an overall reduction

in A-weighted sound power level of nearly 6 dB—or a

75% reduction in A-weighted sound power—as a result

of the entire noise control package. If implemented on

all vibrating screens within a preparation plant, the

entire noise control package has great potential to reduce overall levels in many areas of the facility.

Fig. 21—Spring inserts in place and springs

installed on the vibrating screen.

100

125

160

200

250

315

400

500

630

800

1,000

2,000

4,000

5,000

8,000

10,000

Overall 

1,250

1,600

2,500

3,150 

6,300 

95 

Sound Pressure Level, dB(A) 

90 

85 

80 

75 

70 

65 

60 

55 

93 92 

Baseline (with Barrier) 

Spring Inserts + CLD Plates 

50 

1/3-Octave-Band Center Frequency (Hz) 

Fig. 22—1/3-Octave-band sound pressure level comparison with all equipment running with

barrier-absorber material installed, with and without noise controls.

To ensure that an acoustic enclosure would be

practical to implement on a vibrating screen and that

the mechanism temperatures would remain within the

manufacturer’s limits, several cooling options were

explored. A 330.3-L/s (700-CFM) exhaust fan for the

enclosure running at 100% flow was found to provide

sufficient cooling. A potential area for future work

would be to design a more efficient cooling system by

using lower-speed fans, redesigning the ductwork, and

improving the air circulation within the enclosure.

Better circulation of the air within the enclosure

would be achieved by mounting fan blades on the

spinning coupler ring between the two mechanism

housings so that the air within the enclosure is mixed

as the mechanisms turn. Better air mixing could potentially allow a shift-long run time with only the 80.2-L/s

(170-CFM) fans. However, the current cooling system

with the 330.3 L/s (700 CFM) fan does make it possible

to utilize an acoustic enclosure in a production

environment.

The two controls evaluated at the preparation plant—

the CLD plates and spring inserts—met the dual goals

of reducing sound levels and demonstrating sufficient

durability in a production environment. However, for

this particular trial, the spring chatter that the spring

inserts are designed to address was not observed in

the field prior to installation. When spring chatter was

observed in the lab, the spring inserts and CLD plates

in combination reduced the A-weighted sound power

levels by 2.7 dB. The combination of spring inserts

and CLD plates in the field reduced sound levels by
1 dB. Given the lack of pre-existing spring chatter in

this case, it is unlikely that the spring inserts had any

effect on the sound levels produced, and the reduction

observed is likely due to the CLD plates alone. In this

context, the field results are comparable to the laboratory results since prior lab testing showed a 1-dB reduction in the overall A-weighted sound power levels due to

the CLD plates alone34.

Notably, this 1-dB reduction achieved from the CLD

treatments is just one example of what can be obtained

from the final package of noise controls. As mentioned

above, previous lab results showed a 1-dB reduction in

overall A-weighted sound power levels from the CLD

plates alone34, and the results presented herein show a

6-dB reduction in overall A-weighted sound power

levels from the complete package of noise controls

including the enclosure, tuned mechanism suspension,

CLD plates, and spring inserts. If the noise controls

are incorporated directly into the final product design,

even greater reductions could be achieved. Some noise

control concepts, such as the CLD plates, are limited

in their success by being add-on features rather than

existing as two cast shells for the mechanism housings

with damping material in between. The optimum noise

control solution would be to integrate the control principles defined here into the development of a new design

for a horizontal screen. Such a design would combine

reduced noise with practicality for maintenance and operation, and possibly reduced overall cost compared to

add-on treatments.

8ACKNOWLEDGEMENTS

The authors would like to thank the employees at our

partner preparation plant for their assistance with our

field study, and NIOSH OMSHR employees: Pat

McElhinney and Jeff Yonkey for their assistance with

the field testing, Vincent Conteen and Andy Lawczniak

for preparing the CLD plates for installation, Vincent

Conteen for modifying the spring inserts, and Lynn

Alcorn for help with the TUMS installations.

9REFERENCES

1.	 Bureau of Labor Statistics, “Workplace injuries and illnesses in

2005”, United States Department of Labor, (2006).

2.	 John R. Franks, Mark R. Stephenson and Carol J. Merry,

“Preventing occupational hearing loss—a practical guide”,

Technical Report No. 96–110, National Institute for Occupational Safety and Health, (1996).

3.	 National Institute for Occupational Safety and Health, National

Occupational Research Agenda — Update 2003, DHHS

(NIOSH) Publication No. 2003-148, (2003).

4.	 SangWoo Tak, Rickie R. Davis and Geoffrey M. Calvert,

“Exposure to hazardous workplace noise and use of hearing

protection devices among US workers—NHANES, 1999–

2004”, Am. J. Industrial Medicine, 52, 358–371, (2009).

5.	 SangWoo Tak and Geoffrey M. Calvert, “Hearing difficulty

attributable to employment by industry and occupation: an

analysis of the National Health Interview Survey—United

States, 1997 to 2003”, J. Occupational and Environmental

Medicine, 50, 46–56, (2008).

6.	 National Institute for Occupational Safety and Health, “Preventing Occupational Hearing Loss: A Practical Guide”,

DHHS (NIOSH) Publication No. 96-110, 19–25, (1996).

7.	 Federal Register, “Health Standards for Occupational Noise

Exposure; Final Rule”, Department of Labor, Mine Safety

and Health Administration, 30 CFR Part 62, 64(176), 49630–

49634, (1999).

8.	 PEIR, “Mine Data Retrieval System”, United States Department of Labor Mine Safety and Health Administration,

<http://www.msha.gov/OpenGovernmentData/OGIMSHA.

asp> downloaded December 9, 2010, with the exception of the

MSHA Accident data file, which was downloaded December

27, 2010.

9.	 J.S. Vipperman, E.R. Bauer and D.R. Babich, D.R., “Survey of

noise in coal preparation plants”, J. Acoust. Soc. of Am., 121,

197–205, (2007).

10. M.N. Rubin, A.R. Thompson, R.K. Cleworth and R.F. Olson,

“Noise Control Techniques for the Design of Coal Preparation

Plants”, BuMines OFR 42–84, (1982).

11. E.E. Ungar, G.E. Fax, W.N.	 Patterson and H.L. Fox, “Coal

Cleaning Plant Noise and Its Control”, BuMines OFR 44–74,

(1974).

12. Eric E. Ungar, Clive L. Dym and Matthew H. Rubin, “Practical

reduction of noise from chutes and screens in coal cleaning

plants”, BBN Report No. 3337, (1976).

13. M. Jenae Lowe and David S. Yantek, “Noise survey of aggregate industry vibrating screens”, NoiseCon11, (2011).

14. Matthew N. Rubin, “Demonstrating the noise control of a coal

preparation plant — Volume I. Initial installation and treatment

evaluation”, BBN Report No. 3631, (1977).

15. Matthew N. Rubin, “Demonstrating the noise control of a coal

preparation plant — Volume II. Long-term treatment evaluation”, BBN Report No. 4922, (1982).

16. R. J. Pokora, T. G. Bobick and T. L. Muldoon, “Retrofit noise

control modifications for crushing and screening equipment in

the nonmetallic mining industry, an applications manual”,

Information Circular No. 8975, U.S. Department of the Interior,

Bureau of Mines, (1984).

17. Harsha Vardhan, Y.V. Rao and N.C. Karmakar, “Noise analysis

of heavy earth moving machinery deployed in opencast mines

and development of suitable maintenance guidelines for its

attenuation — Part 1”, Noise & Vibration Worldwide, 35(8),

11-24, (2004).

18. Kevin S. Hennings and Kenneth N.	 Allan, “Noise reduction

from vibrating screens”, 1977 Coal Convention of the American Mining Congress Session Papers, 1–12, (1977).

19. Robert J.	 Pokora and Terry L. Muldoon, “Demonstration of

noise control techniques for the crushing and screening of nonmetallic minerals”, BM-8001, Foster-Miller Associates, Inc.,

(1982).

20.	 Hua Zhu, Shirong Ge and Yu Xiaoyan, “A novel low-noise

vibrating screen”, Coal Preparation, 24(1–2), 85–96, (2004).

21. K. Hennings, “Noise abatement of vibrating screens using nonmetallic decks and vibration treatments”, OFR 120-82, AllisChalmers Corporation, (1980).

22. Wu Miao et al.,	 “Dynamic characteristics analysis of 2065A

two-axle vibrating screen and its noise control”, Journal of

China University of Mining & Technology, 22(2), 8–12

(1993), (in Chinese).

23. Yueming Zhu,	 “Noise control of vibrating screen”, Journal

of Liaoning Technical University, 18(5), 494-496, (1999),

(in Chinese).

24. Wei An, Hua Zhu and Xin Lu, “Development of low noise twochannel linear vibration screen”, China Mechanical Engineering, 15(4), 300–302, (2004), (in Chinese).
25. Xuedong	 Wang, “Technical innovation of high-frequency

vibrating screen”, Journal of Hebei Energy Institute of Vocation

and Technology, 3(20), 78–80, (2006), (in Chinese).

26. Jean-Guy St-Pierre and André Godbout, “Screening apparatus

for efficiently separating coarse material from finer material”,

U.S. Patent 5,232,098, (1993).

27. Louis W. Johnson and Bruce G. Johnson, “Shaker screen”, U.S.

Patent 4,529,510, (1985).

28. Jun Luo and Yimin Wang, “Noise control of vibrating screen”,

Environmental Engineering, 24(3), 57–59, (2006), (in Chinese).

29. Daiguang Song, Bu Xiangquan and Yanjun Zhao, “Using analysis of YA1340 vibrating screen”, Coal Mine Machinery, 28(7),

156–157, (2007), (in Chinese).

30. D.S. Yantek, P. Jurovcik and E.R. Bauer, “Noise and vibration

reduction of a vibrating screen”, Transactions of the Soc. for

Mining, Metallurgy, and Exploration, 318, 201–213, (2005).

31. D.S. Yantek, H.E. Camargo and R.J. Matetic, “Application of a

microphone phased array to identify noise sources on a horizontal vibrating screen”, 2008 National Conf. on Noise Control

Engineering, (2008).

32. H.E. Camargo, P.A. Ravetta, R.A. Burdisso and D.S. Yantek,

“Noise source identification on a horizontal vibrating screen”,

Mining Engineering, 61(7), 47–51, (2009).

33. D.S. Yantek and H.E. Camargo, “Structural vibration as a noise

source on vibrating screens”, ASME 2009 International

Mechanical Engineering Congress & Exposition, (2009).

34. M.J. Lowe, D.S. Yantek, H.E. Camargo, L.A. Alcorn and M.

Shields, “Noise controls for vibrating screen mechanisms”,

Transactions of the Soc. for Mining, Metallurgy, and Exploration, 330, 446–451, (2011).

35. M. Jenae Lowe, David S. Yantek, Hugo E. Camargo and Lynn

A. Alcorn, “Acoustic enclosure to reduce noise from vibrating

screen mechanism housings”, U.S. Department of Health and

Human Services, Public Health Service, Centers for Disease

Control and Prevention, National Institute for Occupational

Safety and Health, DHHS (NIOSH), Technology News, 538,

(2009).

36. Renzhi Wei et	 al, “Research and design of new type low

noise vibration screen”, Coal Science and Technology, 23(12),

1–4, (1995), (in Chinese).

37. Hans	 Jakobs, “Screen structure”, U.S. Patent 3,024,912,

(1962).

38. P.O. Edmond Booth, “Vibrating screen mechanism with resilient connection between gear and counterweight”, U.S. Patent

2766,629, (1956).

39. Hua Zhu, Gu Yuhua, Wu Miao and Renzhi Wei,	 “Development of low noise two-axle vibrating screen”, Chinese

Journal of Mechanical Engineering, 34(4), 96–101, (1998),

(in Chinese).

40. Renzhi Wei and Wu Miao, “The experimental study on flexible

web gear”, MPT 91 JSME, International Conference on Motion and Power Transmissions, (1991).

41. Jie Li and Hu Ping, “Reducing the noise of vibrating screen

using double-body structure”, Mechanics in Engineering, 20

(5), 27–29, (1998), (in Chinese).

42. Weiyuan Song, Lei Wang and Zhihui Zhang, “Vibrating Isolation Ring Applied in Noise Control in Vibrating Screen”,

Journal of Liaoning Technical University, 22(6), 782–784,

(2003), (in Chinese).

43. Clifford G. Hollyfield and Allen S. Jackson, “Vibrating screen

separator”, U.S. Patent 5199,574, (1993).

44. James E. Shahan, “Vibratory screen having noise level reduction by isolation”, U.S. Patent 4, 180,458, (1979).

45. David S. Yantek and M. Jenae Lowe, “Analysis of a mechanism

suspension to reduce noise from horizontal vibrating screens”,

Noise Control Engr. J., 59(6), 568–580, (2011).

46.	 Acoustics — Determination of sound power levels and sound

energy levels of noise sources using sound pressure — Engineering methods for small movable sources in reverberant

fields Part 1: Comparison method for a hard-walled test room,

International Standard ISO 3743-2:1994, International Organization for Standardization, (1994).

47. J.	 Shawn Peterson, David Yantek and Adam K. Smith,

“Acoustic testing facilities at the Office of Mine Safety and

Health Research”, Noise Control Engr. J., 60(1), 85–96,

(2012).

48. David S. Yantek and M. Jenae Lowe,	 “Horizontal Vibrating

Screens: Noise in Coal Preparation Plants, Dominant Noise

Sources, and Noise Controls”, International Coal Prep

2012—Annual Coal Processing Exhibition and Conference,

(2012).

49. David S. Yantek, presented at the Office of Mine Safety and

Health Research mid-year project reviews, April 30, 2009,

(unpublished).