LABORATORY INVESTIGATION OF SEAT SUSPENSION 

PERFORMANCE DURING VIBRATION TESTING

Alan G. Mayton Joseph P. DuCarme 

Christopher C. Jobes Timothy J. Matty 

National Institute for Occupational Safety and Health, Pittsburgh Research Laboratory 

 626 Cochrans Mill Road, Pittsburgh, Pennsylvania 15236 

Telephone: (412) 386-4657; fax: (412) 386-6764; e-mail: AMayton@cdc.gov

ABSTRACT 

Mining injury statistics show that a significant number of 

back, neck, and head injuries are linked to exposure from 

vehicle vibration. Use of a suspension seat is a common way 

to isolate the vehicle operator from the adverse effects of 

vibration exposure. Thus, researchers at the National Institute 

for Occupational Safety and Health1

 – Pittsburgh Research 

Laboratory (NIOSH – PRL) performed laboratory studies on 

four passive and two semi-active seat suspension designs. 

These are typical of seat suspensions commonly found on 

large off-road heavy surface mining, construction and 

agricultural vehicles as either replacement or original 

equipment manufacturer (OEM) systems. One included a 

pneumatic (air bladder) spring mechanism. The fifth and sixth 

suspensions were a NIOSH magnetorheological (MR) semiactive damper design based on the pneumatic (air bladder) and 

one of the coil spring suspensions above. These suspensions 

were modified with a commercially available MR damper 

substituted for the OEM damper. 

These six seat suspension systems were tested and 

analyzed, for vertical vibration only, using the ISO 5007 

Standard [1]. This paper describes the laboratory vibration 

tests using a MTS® shaker table and discusses the results 

obtained for the different suspension designs and highlights 

the rheonetic technology studied. Implications of the seat 

suspension designs relative to their capabilities for isolating 

vehicle operators from vibration exposure are discussed. 

Results for suspensions 1 through 3 showed frequencies 

of isolation from 2.1 to 3.0 Hz using the 40-kg (88-lb) mass 

and from 1.65 Hz to 1.8 Hz using the 80-kg (176-lb) mass. 

Suspension #4, in tests with only the 80-kg (176-lb) mass, 

showed an isolation frequency of 3.7 Hz. With the MR 

 

1

 The findings and conclusions in this report are those of the 

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

National Institute for Occupational Safety and Health. 

Mention of any company or product does not constitute 

endorsement by NIOSH. 

damper added to seat suspension #4, the peak transmissibility 

was lowered from 1.3 to 0.95 and showed a corresponding 

downward shift in frequency from 2.25 Hz to 1.4 Hz. In fact, 

the results for suspension #5 (the MR damper added to seat 

suspension #4), using test #3 conditions of the programmed 

control algorithm, showed isolation (attenuation of transmitted 

vibration) throughout the test frequency range from 1.0 to 

6.0 Hz. 

INTRODUCTION 

NIOSH mining vehicle seat and vibration research is 

dedicated to reducing the risk of injuries to the back, neck, or 

head for vehicle operators through improved mining vehicle 

seat designs. In this regard, injury statistics for mobile mining 

equipment operators from the Mine Safety and Health 

Administration (MSHA) showed incidences of exposure to 

whole-body vibration (WBV) and mechanical shock (vehicle 

jarring or jolting). These injuries can be described as acute 

and chronic musculoskeletal disorders affecting the back, 

neck, and head. For example at surface mining operations, 

Wiehagen et al. [2] reported that vertical and lateral vehicle 

jarring or jolting were the most frequent cause of injury to 

bulldozer operators while operating those vehicles. Moreover, 

the total MSHA reported injuries involving the back, neck, 

and head for 1999 through 2003 resulted in an average of 57% 

per year for injuries related to vehicle jarring or jolting. 

One way to lessen the adverse effects of vehicle vibration 

to the operator is through the use of a seat suspension system. 

Suspension seats are included in most industrial vehicles and 

are passive in nature in that they consist of a damper and some 

form of a spring. Moreover, during operation, these vehicles 

are subjected to significant energy in the 2 to 4 Hz region 

where conventional seats tend to amplify vertical vibration. A 

common method used to reduce vehicle operators’ exposure to 

vehicle vibration is to design the natural frequency of the seat 

suspension so that it is considerably lower than the typical 

operating frequency generated by the vehicle. However, a 

large decrease in seat stiffness (“softer” or lower spring rate) 

is required to achieve this. Considering equation (1): 

1
 (1) 

where, ω is the natural frequency, 

 k is the spring constant, and 

 m is the mass of the seat and occupant. 

A four-fold decrease in seat stiffness is necessary to reduce the 

natural frequency by 50 % [3]. Doing this however, now 

increases the likelihood that the suspension system will 

“bottom out.” 

Moreover, a high vehicle torque requirement also 

contributes to vibration at the operator deck. Passive seat 

suspension design includes selecting an appropriate natural 

frequency and optimization of damping according to a set 

stiffness. The consequence is a trade-off between isolation of 

vibration peak amplitudes at resonance and isolation of higher 

frequency vibrations. The potential benefit of a semi-active 

system is the removal of this trade-off since the system is 

designed to adjust the damping in real time. 

METHOD 

NIOSH-PRL researchers conducted laboratory 

evaluations on different seat suspension designs, including 

two based on a magnetorheological (MR) semi-active design. 

Four of the seat suspensions were replacement or original 

equipment manufacturer (OEM) systems commonly found on 

large heavy vehicles. The different suspensions were 

evaluated to determine how effective they would be if used to 

improve seat performance on large heavy vehicles used at 

mining operations. 

A single-axis, MTS® hydraulic shaker was used to 

conduct vibration experiments. The shaker is capable of 

producing accelerations in excess of 10 g=s at frequencies 

below 1 Hz up to 15 Hz with a maximum travel of 152 mm (6 

in). Seat suspensions were tested using the ISO 5007 

Standard. The test protocol in ISO 5007 Section 10.1 requires 

a sinusoidal vibration of amplitude ±15 mm (0.59 in) and a 

frequency range from 0.5 to 2 Hz at 0.05 Hz intervals. Test 

weights or masses of 40 and 80 kg (88 and 176 lbs) are 

specified by ISO 5007 to simulate the upper and lower ends 

(5th percentile female and 95th percentile male) for the range of 

seated vehicle operators. In addition to the test mass, extra 

mass was added to duplicate seat hardware that was removed, 

such as the seat cushion and backrest. A test frequency range 

from 2 to 8 Hz at 0.25 Hz intervals was added to measure the 

transmissibility characteristics for each seat suspension system 

in the range most sensitive to the human body overall. The 

acceleration of the shaker and the seat pan was measured and 

used to calculate the ratio of the seat pan RMS acceleration to 

the shaker table RMS acceleration (i.e., transmissibility ratio). 

This ratio is commonly used as a measure of how well the seat 

suspension is isolating the occupant from vehicle vibration. 

To aid in the analysis of the data, graphs showing the 

transmissibility of each suspension were created. 

TESTING PROCEDURES 

The test procedures included the following steps: 

1. Each channel of an 8-channel Sony® PC208AX 

digital recorder was calibrated to ensure input. 

2. On a data-acquisition log sheet, recorder channel 

assignments with accelerometer mounting location, 

accelerometer axes, signal conditioners, and recorder 

settings, along with a recording log, were noted. ω = / mk 3. Preliminary shaker testing determined the 8-channel 

recorder voltage range (-5vdc to +5vdc volts) to 

capture desired events without clipping signals or 

being within noise levels (weak signals) of the 

system. 

4. The 8-channel recorder sampling frequency was set 

to the normal setting – 5 kHz per channel. 

5. Accelerometer amplifiers were set for the desired 

accelerometer charge sensitivity, gain, and a low-pass 

filter with a cut-off frequency of 100 Hz to reduce 

unwanted noise. 

6. Two B&K® 4370 accelerometers, with magnetic 

mountings, were attached to the suspension and 

mounting base plate installed on the shaker; the 

small-diameter, accelerometer cables were secured to 

minimize noise induced from unrestrained movement 

during testing. 

a. One accelerometer was placed in the center of 

the rigid portion of the shaker, as close as 

possible to the seat suspension to record the 

shaker input accelerations. 

b. A second accelerometer was placed at the center 

of the seat suspension (at the seat pan location) 

to record vibration transmitted through the seat 

suspension. 

7. To identify each test frequency change during data 

collection and aid in analyzing the data, a 5-Vdc 

pulse was used as a marker on one of the recorder 

channels. 

8. A completed data-acquisition log sheet described the 

logging of recorder channel assignments, scaling, 

amplifier settings, accelerometer charge sensitivity, 

and recorder test logging. 

9. Before testing, the seat cushion and backrest of the 

seat suspension system were removed and weighed in 

order to add their weights to the overall test weight. 

10. With the appropriate test weight attached, the seat 

suspension was adjusted to its mid point of travel. 

11. With the loaded seat on the shaker, a test with a 

duration of more than 40 minutes or >5000 cycles 

was conducted, in the amplitude control mode, with a 

displacement of ±15mm (0.59 in) at 2 Hz. The seat 

test load consisted of 40 kg (88 lbs) plus extra weight 

equal to the removed seat cushion and backrest. This 

procedure constituted the “break-in” period of the 

seat suspension. 

12. After the break-in period, testing was again 

performed with amplitude control at a displacement 

of ±15mm (0.59 in) loaded with 40 kg (88 lbs) plus 

extra weight for the seat cushion and backrest. The 

tests continued from 0.5 Hz to 2 Hz in 0.05 Hz 

intervals and from 2 Hz to 8 Hz in 0.25 Hz intervals; 

each interval was recorded for 15 s. 

13. The same test setup in step number 12 was run by 

reversing the frequency order. 

14. The seat test was conducted in amplitude control 

mode with a displacement of ±15mm (0.59 in) loaded 

with 80 kg (176 lbs) plus extra weight equal to the 

removed seat cushion and backrest. The tests ran 

2
from 0.5 Hz to 2 Hz in 0.05 Hz intervals and 

continuing from 2 Hz to 8 Hz in 0.25 Hz intervals 

recording each interval for 15 seconds. 

15. Testing was again repeated in a fashion similar to 

step 14 but with descending frequency order. 

In contrast to testing of the replacement and OEM 

suspension systems, a modified version of the ISO 5007 

Standard was used to test the MR fluid damped suspension 

systems. Modifications included testing solely with the 80-kg 

(176-lb) weight, using a frequency range of 1 to 6 Hz, and 

using a displacement of 18 mm (0.7 in). Frequencies below 1 

Hz were not tested in that brief testing showed that the 

inherent frictional resistance of the suspension components 

caused suspension system transmissibilities to be essentially 

unity in this region, where no movement of the suspension 

occurred. A displacement of 18 mm (0.7 in) was selected, 

instead of 15 mm (0.59 in) specified by ISO 5007, in order to 

apply enough force to operate the suspensions in the frequency 

range of interest. Frequencies above 6 Hz were not tested due 

to safety concerns from high accelerations when operating the 

shaker at a displacement of 18 mm (0.7 in) in lieu of 15 mm 

(0.59 in). Moreover, a linear displacement transducer was 

used to record displacement data. In turn, this data was used to 

compute accelerations and to verify results computed using 

the acceleration data. 

DESCRIPTION OF RHEONETIC SUSPENSION 

TESTING 

Rheonetic technology offers the ability to adjust 

suspension damper characteristics in real time. The fluid used 

in current semi-active dampers is light oil that contains 

micron-sized iron particles. When exposed to a magnetic field, 

the viscosity of the fluid changes proportionally to the strength 

of the magnetic field. Viscosity changes occur very rapidly, 

on the order of several milliseconds. The magnetic field is 

induced by current flow through a coil installed in the piston 

of the damper. Figure 1 depicts a typical rheonetic damper 

[4]. 

Two commercially available seat suspensions were 

modified using the rheonetic system. These suspensions are 

shown in Figures 2 and 3. Both were low travel and low cost 

designs. The designs were similar, with one using an airbladder type spring and the other using a standard steel coil 

spring. The second design also used rolling elements and was 

subject to much less column friction than the first which used 

sliding elements in its linkage. 

 The “brain” of the rheonetic system is a micro-controller 

that monitors seat position, calculates seat velocity, and 

provides a control signal to the rheonetic damper. A prototype 

controller was developed for this project and the hardware 

assembly is shown in Figure 4. It consists of a Kanda Systems 

STK200 Micro-controller Development Board, low-pass filter, 

voltage-to-current converter, and various input-output 

Fig. 2: Suspension #4 mounted to shaker table.

Fig. 3: Seat suspension #5. 

Fig. 4: Prototype rheonetic controller.

80-kg (176-lb) Mass

Linear 

Displacement 

Transducer 

Accelerometers

Fig 1: Typical rheonetic damper. 

MR Damper Air Bladder

3
terminations. The integrated circuit within the microcontroller was an Atmel AT90LS4434. The micro-computer 

board also contained a communications port to a personal 

computer that was used for software development and system 

tuning. 

Numerous control algorithms for semi-active suspension 

systems have been published in literature. Most of them sense 

or measure the relative travel and velocity between the 

suspension system and the sprung mass. Two of the most 

common algorithms are named Skyhook control and Relative 

control. Referring to Figure 5, they are normally defined as 

follows: 

Skyhook Control 

Set Fdamper = C•X ́ if X ́•(X ́-Y ́) > 0 

Fdamper = 0 if X ́•(X ́-Y ́) < 0 

Relative Control 

Set Fdamper = C•(X ́-Y ́) if (X-Y)•(X ́-Y ́) < 0 

Fdamper = 0 if (X-Y)•(X ́-Y ́) > 0 

To implement Skyhook control, the absolute velocity (X ́) 

of the seat is required as an input to the controller. This is 

difficult to obtain in a vehicle because of the lack of an 

absolute reference frame. For Relative control, only one 

position sensor, mounted between the vehicle floor and the 

seat pan, is required. The relative velocities can be obtained 

by differentiation of the displacement signal. Thus the 

quantities (X ́-Y ́) and (X-Y) can be replaced with the relative 

measurement Z ́ and Z, respectively. 

TEST RESULTS 

Seat Suspension #1 (two linear extension springs) 

Suspension #1 is used in a commercially available seat 

that has a vehicle operator weight adjustment of 50 to 120 kg 

(110 to 265 lbs). The suspension included two linear 

extension springs in the upright position, mounted in parallel 

with a damper, and with a 10o layback and 8.6-cm (3.4-in) 

displacement. Figure 6 shows the seat and suspension; 

whereas, figure 7 shows the suspension alone mounted to the 

shaker table. The results from testing seat suspension #1 

appear in Figure 8. Transmissibility (RMS output 

acceleration/RMS input acceleration) versus frequency is 

plotted for the 40 and 80 kg (88 and 176 lbs) masses. 

Fig. 6: Seat with suspension. 

0.4

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 6 6.5 7 7.5 8

Frequency, Hz

Transmissibility Ratio

40 Kg Average 80 Kg Average 1 to 1 Ratio

Fig. 5: Basic parameters for considering 

control algorithm. 

Fig. 7: Seat suspension #1 mounted to shaker. 

Fig. 8: Seat suspension #1 average results of 

transmissibility for 40-and 80-kg (88-amd 176-lb) tests. 

Accelerometers

80-kg (176-lb) Mass 

Linear 

Displacement 

Transducer 

4
Seat Suspension #2 (two linear extension springs)

The seat from which suspension #2 comes is adjustable 

for vehicle operators weighing 60 to120 kg (130 to 265 lbs). 

The mechanical geometry of suspension #’s 1 and 2 are 

identical. The differences lie in the spring and damping rates. 

These two seat suspensions are popular replacement seating 

systems on heavy vehicles weighing 17,499 kg (7,938 lbs) or 

less. Figure 9 shows the average results for tests with the 40- 

and 80-kg (88- and 176-lb) masses. 

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

1.3

0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 6 6.5 7 7.5 8

Frequency, Hz

Transmissibility Ratio

40 kg Average 80 kg Average 1 to 1 Ratio

Seat Suspension #3 (air spring) 

Suspension #3 uses an air spring with a standard damper. 

The mechanical design of the seat is a cross-buck design with 

the pneumatic spring located in the center with a suspension 

stroke of 8.9 cm (3.5 in). The average results for the 40- and 

80-kg (88- and 176-lb) masses appear in Figure 10. 

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

1.3

1.4

0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 6 6.5 7 7.5 8

Frequency, Hz

Transmissibility Ratio

40 kg Average 80 kg Average 1 to 1 Ratio

Seat Suspension #4 (single coil spring) 

A fourth seat suspension was a cross-buck damper, coil 

spring system. The results of tests with this suspension system 

are presented and discussed in the context of rheonetic design 

tests. 

Rheonetic Design Compared with Commercially 

Available Seat Suspensions 

Our research showed that using rheonetic dampers in a 

semi-active seat suspension system can provide modest 

improvements in transmissibility. Figures 11 and 12 show 

transmissibility curves of the rheonetic, OEM mechanical 

(suspension #4), and air suspensions (suspension #3). 

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

1.3

1.4

0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 6 6.5 7 7.5 8

Frequency, Hz

Transm issibility Ratio

Test 003 Test 004 Test 005 OEM Seat 1 to 1 Ratio

Fig. 9: Seat suspension #2 average results of 

transmissibility for 40-and 80-kg (88-amd 176-lb) tests. Fig. 11: Transmissibility curves for suspension #4 

(OEM mechanical) versus suspension #5 (rheonetic 

damper with suspension #4) using several test variations of 

the programmed control algorithm. 

0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

1

1.1

1.2

1.3

1.4

1.5

0.5 1 1.5 2 2.5 3 3.5 4 4.5 5 5.5 6 6.5 7 7.5 8

Frequency, Hz

Transmissibility Ratio

Rheonetic Orginal Suspension 1 to 1 Ratio

Fig. 12: Transmissibility curves for suspension #3 

(air) versus suspension #6 (rheonetic damper with 

suspension #3).

DISCUSSION 

Fig. 10: Seat suspension #3 average results of 

transmissibility for 40-and 80-kg (88-amd 176-lb) tests. All spring-mass-damper systems exhibit a resonant 

frequency. A seat suspension is amplifying the input vibration 

when the transmissibility is greater than one. For seat 

suspension systems, it is desirable to have the resonant or 

natural frequency as low as possible without adversely 

affecting the performance at higher frequencies. This is due to 

the sensitivity of a human occupant to vibrations. Typically, 

5
vibrations at frequencies less than 2 Hz need to be of 

significant amplitude before becoming objectionable. Because 

of this, higher transmissibility below 2 Hz should not 

necessarily be viewed as poorer performance. As shown in 

the graphs, resonant frequencies (the peak of the graph) for the 

six seat suspensions tested were similar, in that they were at or 

near 1.5 Hz. 

The 40- and 80-kg (88- and 176-lb) mass curves for 

suspension #1 show good correlation up to 5.5 Hz. After 

5.5 Hz there is more variation, particularly with the curve 

showing decreasing Hz. Intuitively speaking, more mass 

should provide better attenuation and thus lower 

transmissibility. This is supported by the 80-kg (176-lb) 

curve, which shows a greater and more rapid reduction in 

transmissibility compared to the 40-kg (88-lb) curve, which 

shows a more gradual reduction in transmissibility. 

 As shown by a transmissibility less than one, isolation for 

suspension #2 occurs at about 1.6 Hz with the 80-kg (176-lb) 

mass. This compares with the similar isolation and lower 

transmissibility at 2.5 Hz with the 40-kg (88 -lb) mass. As 

noted above, the larger mass shows a greater and more rapid 

drop in transmissibility as shown by the steepness of the 

curves for the 80-kg (176-lb) mass in the 1.4 to 2.5 Hz 

frequency range. Reductions in transmissibility for 

suspension #2 parallel those for suspension #1 at 1.4 and 2 Hz, 

respectively, for the 40- and 80-kg (88- and 176-lb) masses. 

As noted by equation (1) above, the natural frequency of a seat 

depends on its mass. Thus, a shift in natural frequency with a 

change in mass is certainly expected. 

 The isolation frequency for suspension #3 and the 40-kg 

(88-lb) weight is lower at 2.1 Hz compared to the same for 

suspension #’s 1 and 2. In general, the results shown for the 

air-ride suspension #3 are similar to suspension #2. 

 Common spring-damper systems, such as suspensions #1 

through #4, are called passive suspension systems. Their 

reactions to vibrations are determined by their design 

parameters. Variable rate springs and velocity proportional 

dampers can be used to improve ride characteristics over fixed 

rate components, but as with any mechanical or pneumatic 

system, the results are a collection of design compromises. 

Active systems dissipate energy from the suspension 

system by forcing extension or retraction in response to 

measured and anticipated motion. In these suspensions, the 

dampers are replaced with hydraulic cylinders that are 

pressurized by a hydraulic pump. The major drawback for this 

type of system is that it is complicated, costly, and needs a 

power supply. Some safety concerns also apply. 

Semi-active systems, such as suspensions #5 and #6, 

change the suspension response using rheonetic technology to 

adjust suspension damper characteristics in real time. 

The OEM mechanical system, suspension #4, has 

significant friction in its design due to sliding components. 

The rheonetic damper (suspension #5) was tested using the 

relative control algorithm programmed into the controller. 

Several parameters (e.g., damping coefficient, velocity, 

displacement) were varied for tests. These variations appeared 

to have little effect on the performance in the frequency range 

tested. As the frequency of the input vibration increased, it 

appears that any advantage the rheonetic technology could 

provide in vibration isolation becomes less. This is 

anticipated outcome since damping effects are greatest near 

resonance. 

Suspension #6 (addition of the rheonetic damper and 

controller to suspension #3 - air), did not provide as significant 

an improvement in vibration isolation as it did in the OEM 

mechanical suspension (#5). This is most likely due to the 

design of the air-ride suspension being a superior design. It 

was a more expensive and robust design and included rolling 

elements to reduce friction. Since it incorporated an air 

spring, the rate could be somewhat adjusted for the load, 

unlike suspension #4, the cross-buck damper, coil-spring 

system with its fixed rate spring. The advantage of the 

rheonetic damper was not as great, although the system was 

only marginally investigated. Nevertheless, it did provide a 

measurable improvement. Moreover, the rheonetic 

technology is available from LORD Corporation 

(http://www.lord.com) as a retrofit kit to several popular seat 

suspension systems. 

Likely applications for any of the above seat suspensions 

are large surface mining vehicles such as haulage trucks, 

bulldozers, pan scrapers, front-end loaders, and graders. 

Using the above suspensions with improved seat padding, 

such as that reported by Mayton et al. for underground coal 

mine haulage vehicles [5], shows the potential for improved 

isolation of vehicle operators from WBV and terrain-induced 

vehicle jars and jolts that can increase the risk of operator 

injuries. 

CONCLUSIONS 

Seat suspensions 1 through 3 showed peak 

transmissibilities, and amplification of transmitted vibration, 

occurred at 1.6 Hz to 2.0 Hz for the 40-kg (88-lb) mass and 

1.3 to 1.4 Hz for the 80-kg (176-lb) mass. Similarly, peak 

transmissibilties for suspension #’s 4 and 5 showed 

amplification of vibration at 2.5 Hz and 1.4 Hz with only the 

80-kg (176-lb) mass. 

For suspension #5 (the MR damper added to seat 

suspension #4), the peak transmissibility was lowered from 

1.3 at 2.25 Hz to 0.95 at 1.4 Hz. Moreover, peak 

transmissibilties for several variations of the programmed 

control algorithm for suspension #6 showed amplification at 

1.75 Hz and attenuation at 1.4 Hz with only the 80-kg (176-lb) 

mass. 

Isolation frequencies for seat suspensions 1 through 3 

occurred from 2.1 Hz to 3.0 Hz using the 40-kg (88-lb) mass 

and from 1.65 Hz to 1.8 Hz using the 80-kg (176-lb) mass. 

Suspension #4, in tests with only the 80-kg (176-lb) mass, 

showed an isolation frequency of 3.7 Hz. 

Furthermore, seat suspension #5 for test #3 setting of the 

programmed control algorithm attenuated the transmitted 

vibration throughout the test frequency range from 1.0 to 

6.0 Hz. These results suggest that the application of rheonetic 

technology, in particular, should provide heavy vehicle 

operators with improved isolation from seat transmitted 

vibration and a subsequent reduction in back and neck strain 

injuries. 

REFERENCES 

[1] ISO, 1990, “ISO 5007 – Agricultural Wheeled 

Tractors – Laboratory Measurement of Transmitted 

6
Vibration. International Organization for Standardization, 

Geneva, Switzerland. 

[2] Wiehagen WJ, Mayton AG, Jaspal JS, Turin FC 

[2001]. An analysis of serious injuries to dozer operators 

in the U.S. mining industry. Pittsburgh, PA: 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) Publication No. 2001-126, IC 9455, 23 

pp. 

[3] Griffin, M. J., 1990, Handbook of Human Vibration. 

Academic Press Ltd., NY, pp 399-401. 

[4] Spencer BF, Dyke SJ, Sain MK, Carlson JD. [1996]. 

Phenomenological Model of a Magnetorheological 

Damper. ASCE Journal of Engineering Mechanics, 23 

pp. 

[5] Mayton AG, Amirouche F, and CC Jobes (2005). 

Comparison of Seat Designs for Underground Mine 

Haulage Vehicles Using the Absorbed Power and ISO 

2631-1(1985)-Based ACGIH Threshold Limit Methods. 

International Journal of Heavy Vehicle Systems, vol. 12, 

no. 3, pp. 225-238. 

7