Design, Modelling and testing of a Forklift seat suspension system Andrew Mac Guinness 12042854 Bachelor of Engineering in Mechanical Engineering University of Limerick Supervisor: Dr. Conor McCarthy Final Year Project report submitted to the University of Limerick, 21 st March 2014 I declare that this report is my work and that all contributions from other persons have been appropriately identified and acknowledged.
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Design, Modelling and testing
of a Forklift seat suspension system
Andrew Mac Guinness
12042854
Bachelor of Engineering in Mechanical Engineering
University of Limerick
Supervisor: Dr. Conor McCarthy
Final Year Project report submitted to the University of Limerick, 21st March 2014
I declare that this report is my work and that all contributions from other persons have
been appropriately identified and acknowledged.
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Abstract This thesis details the steps undertook when designing a fully contained forklift
seat suspension system. Research was carried out on journal papers and products
currently on the market. From the research it was learned that poorly designed and
maintained suspension systems can increase the risk of operators being exposed to large
whole body vibrations, which over time can lead to operators contracting pain in the
lower back known as Lumbar Syndrome. The objective of this thesis is to design a seat
suspension system that requires no adjustment by the operators to ensure correct spring
and damping coefficients. Investigations were carried out with multiple concept ideas, it
was chosen that an elastomer may be manufactured to deliver the properties required.
Therefore an investigation was begun to develop an elastomeric material that has a non-
linear spring stiffness, which may lead to eliminating the need for operator to make
adjustments to the system. Theoretical analysis was carried out along with static and
dynamic finite element simulations on a 3-D computer model of the chosen concept to
be developed. The simulations carried out conformed to virtual versions of International
Organization for Standardisation standards for vibration transmissibility. The Seat
Effective Amplitude Transmissibility (SEAT) factor and displacement transmissibility
at resonance was determined. Results indicate that the model if manufactured, would
pass these standards. A SEAT factor of less than 0.9 was determined for frequencies
above 13Hz and displacement transmissibility was determined to be 0.97. A 3-D model
of the simulated concept was printed to illustrate the motion of the concept to peers and
determine any design issues with components that may occur when assembling the
prototype, which may not be noticed when assembling in the virtual space.
Recommendations are made for continuing designing this concept further, such
recommendations include: produce a full scale prototype and carry out physical
simulations in accordance with ISO 7096, carry out physical experiments of a variety of
elastomeric materials to define a material to be used as a combined spring damper
component.
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Acknowledgements I would like to thank the staff of the University of Limerick who have helped me
throughout the year with my project whose knowledge and guidance has been
invaluable namely;
My supervisor, Dr. Conor McCarthy for his guidance and help throughout this project.
Dr. Joseph Leen for organising and facilitating the printing of the 3-D scale model. Brian Nestor for printing the 3-D scale model.
I would like to thank Gareth Murry for supplying a licensed copy of Solidworks
Premium 2012.
I would also like to thank my family and friends for their support, encouragement and
guidance throughout the project, especially my sister Amanda-Jane Gainford who was
always there for me no matter what time day or night.
Nomenclature !Symbol Description Unit A Cross sectional area m2
a Acceleration m/s2 !!(!!) Unweighted rms value of the measured vertical
acceleration at the seat disk at the resonance frequency
Hz
!!(!!) Unweighted rms value of the measured vertical
acceleration at the platform at the resonance frequency
Hz
C Damping coefficient Ns/m E Young’s Modulus N/m2
F Force N Fd Force transmissibility Ratio k Spring stiffness N/m m Mass kg r Frequency ratio Ratio T Time s Td Displacement transmissibility Ratio X Seat response m Y Base excitation m
! Strain Ratio ! Damping ratio Ratio ! Stress N/m2 ! Frequency Hz ! Frequency Hz
!
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List of Figures Figure! Description! Page!
Number!
Figure 2.1 (a) Rear mounted suspension system 5 Figure 2.1 (b) Base mounted suspension system 5 Figure 2.2 Illustration of human natural frequencies, (source; Rao
2004,P661) 7
Figure 2.3 Prevalence of Lumbar syndrome (Schwarze ,1998 ,P618 7 Figure 4.1 One degree of freedom system with base excitation 18 Figure 4.2 55kg Human, natural frequency of pelvis 4Hz 20 Figure 4.3 55kg, natural frequency of pelvis of 4Hz 20 Figure 4.4 Minimal seat footprint dimension 22 Figure 4.5 Concept one sketches, (a) 3D sketch illustrating the design
idea 23
Figure 4.5 (b) 2D side view illustrating the direction of the seat travel 23 Figure 4.6 Concept two sketch illustrating the use of elastomer as a
combined spring damper 24
Figure 4.7 Concept three illustrating idea of a back mounted seat suspension system
25
Figure 5.1 (a) One elastomer (blue) connecting both swing arms
26
Figure 5.1 (b) Two elastomers connecting swing arms to the base 26 Figure 5.2 (a.1) Single elastomer at rest 27 Figure 5.2 (a.2) Single elastomer fully compressed 27 Figure 5.2 (b.1) Two elastomers at rest 27 Figure 5.2 (b.2) Twe elastomers fully compressed 27 Figure 5.3 (a) Two component swing arm 27 Figure 5.3 (b) One folded swing arm 27 Figure 5.4 Exploded view of seat suspension concept to be tested 28 Figure 5.5 (a) End view of suspension fully at rest 29 Figure 5.5 (b) End view of suspension fully compressed 29 Figure 5.6 Isometric view of suspension base 30 Figure 5.7 Half of suspension base to be analysed 31 Figure 5.8 Mesh convergence graph of suspension base 32 Figure 5.9 (a) Illustrates overall stresses induced in the base 32 Figure 5.9 (b) Illustrates close up of the maximum stress induced 32 Figure 5.10 View of sear mounting plate 33 Figure 5.11 Isometric view of the rear vertical swing arm 33 Figure 5.12 (a) Dissipation of stress throughout the component 34 Figure 5.12 (b) Maximum stress felt by component 34 Figure 5.13 Isometric view of the front vertical swing arm 34 Figure 5.14 Isometric view of the elastomer holder 35 Figure 5.15 Illustration the stress concentration point 36 Figure 5.16 Displacement transmissibility against frequency ratio, with
damping constant of 0.4 38
Figure 5.17 (a) Maxwell model 40 Figure 5.17 (b) Kelvin Model 40
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Figure 6.1 (a) Boundary conditions, pin connections and virtual springs
44
Figure 6.1 (b) Mode shape 1 at first natural frequency 44 Figure 7.1 Outlining the seat plate and suspension base 46 Figure 7.2 Linear acceleration of seat plate input vibration 16 Hz 47 Figure 7.3 Linear acceleration of suspension base input vibration 16
Hz 47
Figure 7.4 Seat plate acceleration response to forcing function almost equal to resonance
49
Figure 8.1 (a) 3-D computer model 50 Figure 8.1 (b) Printed model 50 Figure 8.2 (a) Fully extended computer model 51 Figure 8.2 (b) Fully extended printed model 51 Figure 8.3 (a) Fully compressed computer model 51 Figure 8.3 (b) Fully compressed printed model 51 Figure 8.4 (a) Computer model with seat 51 Figure 8.4 (b) Printed model with seat projected 51 Figure 9.1 (a) Vertical suspension system with reaction load
absorption area 52
Figure 9.1 (b) Diagonal travel with reaction load absorption area
52
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1. Introduction
1.1 Background By design forklift trucks have a rigid chassis for stability when lifting and
moving heavy loads. Therefore they are able to manoeuvre easier while keeping heavy
loads stable. However, having a rigid chassis can result in all vibrations generated from
the truck driving over rough terrain, transferring vibrations to operator as there is no
way of them being absorbed and dissipated.
Exposure to constant whole body vibrations over time can cause pain and
vibrations because the effects of vibrations are not seen instantaneously. The pain
associated with exposure to whole body vibrations is due to the spine weakening as a
result of cumulative trauma, which will be discussed further later in the report.
1.2 Design Brief Design and carry out performance simulations of a new novel forklift truck seat
suspension system. The design should be more cost effective for producing and
maintaining then existing suspension system designs while being fully self contained.
The design should also have adequate vibration transmissibility dampening in order to
potentially reduce operator discomfort and pain during operation of the forklift truck as
well as meet all international standards relating to seat suspension design.
1.3 Aim and Objectives ! The propose of this report is to investigate and design a novel prototype seat for
off road forklift trucks to reduce vibrations transmitted from the forklift trough to the
driver. Due to project time restrictions, development and testing of the prototype will be
restricted to theoretical analysis, computer modelling and testing by means of Finite
Element software (FE). The Finite Element Analysis (FEA) will demonstrate if the
prototypes design will reduce vibration-transmitted through to the operator. The
suspension system should reduce the peak accelerations experienced by the operator, in
turn reducing the effects of being exposed to whole body vibration over time, which is
discussed further in this report. From the Design Brief section 1.2, the seat suspension
concept has to be designed with the following criteria in mind:
• Cost effective production and operation
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• Fully contained, i.e. no external power source required • Pass computer simulated version of ISO-7096 – “earth-moving machinery-
laboratory evaluation of operator seat vibration” • Achieve an natural frequency outside the natural frequencies of the human body
1.4 Overview of Thesis Prototype development was carried out in stages of the investigation. Firstly a
review of the market was carried out, focusing on the current state of the art suspension
systems, in order to gain a better understand of the design problem and review existing
technologies.
Journal papers were reviewed in order to understand the need for further
development of the prototype seat suspension system. The medical implications of
exposure to prolonged periods of whole body vibrations, testing on previous prototypes
and how suspension system motion affects the ergonomics of the seat for the operator
were just a few criteria taken into consideration. A review on the background research
can be found in chapter two.
Having gained an understanding of the current state of the art products currently
on the market and the up-to-the-minute research presented in section 2.2. Brainstorming
and concept development took place. This was followed by preliminary calculations of
each sketched concept detailed in chapter four. The chosen concept utilises the idea of
using an elastomeric material to act as a combined spring and damper. Before modelling
the chosen concept, a basic wooden model was constructed, figure 5.2, to ensure the
linkages would move as intended. The dimensions were based on International
Organization for standardization (ISO) to accommodate all operators in both size and
weight.
Initially 2-D FE models were constructed which facilitated a simple analysis of
the design. From the results obtained in the 2-D analysis a 3-D model was created for
the purposes of simulating, ISO-7096 – “earth-moving machinery - Laboratory
evaluation of operator seat vibration”. Simulations for the 3-D model were carried out
using Solidworks Simulate 2012. A CD is included in appendix A illustrating the 3-D
model, simulation setup of 3-D model and the reactions to the simulated conditions.
Results illustrated in chapter 7 illustrate that the initial simulations pass the criteria for
the standards mentioned previously. This thesis concludes with recommendations that
further work should be carried out to further validate and develop the chosen concept.
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2. Literature Review
2.1 Introduction Currently there are many different styles of low-profile forklift seat available to
the consumer, however each year in Europe many forklift operators still incur injuries to
their lower back. Epidemiological and biomechanical studies have been carried out with
results concluded that these injuries occur due to prolonged exposure to constant
vibration generated in the forklift, which are transmitted though the seat to the operator,
(Silsoe research institute, 2000). Therefore the forklift seats aim of reducing the
vibration transmitted, reducing the frequency of vibration and reducing the shock loads
felt from seat suspension reaching its limit of travel has been and still remains a major
topic of research in this industry. As to be expected, the main driving force in the
continuation of seat design research is to prevent workplace injuries, thus making the
customers business more productive and efficient by reducing the amount of sick days
operators will take due to injuries caused by operating forklift machinery.
Considerable literature is available that outlines International Industry Standards
for many aspects of designing a forklift seat, some of which are outlined in chapter 3.
The literature includes areas such as concept testing, material selection and vibration
transmissibility allowances. There are also many medical journals discussing the
resulting implications to operator’s health that are exposed to Lumbar syndrome. The
main symptom of the syndrome is lower back pain, which is caused from prolonged
periods of constant whole body vibration. There are journals that medically examine the
operators’ health over a period of time, which have been published by Donati,P (2002)
and Schwarze,S & Notbohm,G (1998). The aim of this literature review is not to
summarize and rewrite the work and conclusions of these authors, but to focus on
subject areas most relevant to the objectives of this report, discuss what was learnt from
reading these papers and outline the direction further work in this report will take.
Subject areas that are focused on are;
• Current state of the art suspension classifications
• Natural frequencies of the human body
• Effects of prolonged exposure to whole body vibration
• Design considerations for reducing whole body vibrations
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!2.2 Current State of the art suspension classifications
Currently on the market there are many different suspension systems available, the
seat suspension can be split into three different categories:
1. Conventional
2. Low profile
3. Compact
Conventional
Typically this type of design has a general suspension travel length between
130-150 millimetres (mm). This seat suspension is mainly used in the marine and
Heavy Goods Vehicles (HGV) industries where the operators are travelling at speed
over a wide variety of terrain, the most common type of suspension system incorporates
pneumatic systems to automatically adjust to the operator’s weight.
Low Profile
Typically this type of design has a general suspension travel length between 35-
60mm, which is commonly used where there is rough terrain. However the operator cab
has restricted headroom for the operator and incorporating a suspension travel length
similar to that of the “conventional” designs are not feasible. There are many types of
suspensions systems on the market under the banner of “low profile” suspension
system that use many different methods of damping and absorbing the vibrations. Some
of which include mechanical spring/damper system, pneumatic systems, hydraulic
systems and electrically controlled spring/ damper systems.
Compact
Typically this type of design has a general suspension travel length between 25-
45mm, commonly used where the terrain is relatively smooth. Therefore the only
vibrations transmitted to the operator theoretically, are the vibrations generated from the
engine of the forklift truck. For the “compact” seat suspension system, it is common
practice to utilize mechanical springs and dampers.
This report will focus on designing a prototype seat suspension design to fall
within the “Low profile” category, because of the type of operating conditions that the
concept is being designed for is a forklift truck that is capable of operating off-road as
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well as on smooth terrain. This Hybrid design of forklift truck means that the seat
suspension system needs to adequately dampen the shock loads applied while the
forklift truck is off-road. The Low-profile suspension category should function well
with any degree of shock applied.
2.2.1 Overview of existing Low-Profile Suspension Systems Designs
The main design characteristic of a Low-profile suspension system is that the
length of stroke of the suspension is between 35-65mm. The short travel length allows
the seat to be fitted in machines where height is limited, but the length of stroke is long
enough to adequately dampen shock and vibrations transmitted through the machine.
With such a diverse range of seat designs on the market today there are many different
methods of reducing vibration transmissibility through the seat. Regardless of each
individual design, they all have the same goal. Figure 2.1 below shows the side profile
of two types of seat suspension systems, Figure 2.1(a) incorporates a spring damper
type of set up attached to the rear of the seat, which design allows for a relatively large
vertical seat displacement as the hardware for controlling the spring and damping rate
are positioned behind the seat rather than positioned below the seat as in Figure 2.1(b),
whereby the suspension system itself is obstructing the movement of the seat, and in
turn limiting the max stroke length.
FIGURE 2.1(A). REAR MOUNTED SUSPENSION SYSTEM FIGURE 2.1(B). BASE MOUNTED SUSPENSION SYSTEM!
2.3 Resonant frequencies of the human body
If the frequency of the external force to a system coincides with one of the
natural frequencies of the system itself, a condition known as resonance occurs. This
occurs when the system undergoes dangerously large oscillations (Rao 2004, P.16).
Therefore, it is fundamental that the natural frequency of the human body is understood.
As the human body contains a variety of materials, geometries and masses there is not
one frequency that resonates with the entire human body. Rao determined the natural
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vibration frequencies of the main sections of the human body, by illustrating the natural
frequencies as a system of springs and dampers, shown below in figure 2.2.
FIGURE 2.2ILLSUTRATION OF HUMAN NATURAL FREQUENCIES, (SOURCE; RAO 2004, P. 661)
Further research was carried out to validate Rao’s work focusing on the pelvic
mass, buttocks and lumbar region of the human body. Through mathematical models
and FEA simulations, it was confirmed that the natural frequency of the pelvis, lumbar
region is between 4-9 Hz (Maciejewski and Meyer et al., 2008, pp. 520-538). This data
was supported by research carried out by Hostens, which stated that the natural
frequency of the Lumbar region in the back is on average 5 Hz but a range of 4-6 Hz
must be used in order to take into account the variation in size of operators operation the
machinery (Hostens and Deprez et al., 2003, pp. 141-156). Taryen Hill (2009) reported
from several studies researched from widely accepted journals that the notable values
for the lumbar vertebrae is 4.4 Hz (Hill and Desmoulin et al., 2009, pp. 2631-2635).
2.4 Effects of prolonged exposure to vibration As mentioned in Chapter 1, there are many case studies linking lower back pain
to prolonged exposure of whole body vibrations because of the cumulative exposure to
the vibrations. The pain experienced is usually localized in the lumbar region of the
back along the spine. The constant vibrations extend and compress the intervertebral
discs, which over time wear them down until the vertebra start to rub against one
another damaging nerves, known as Lumbar syndrome. Lumbar Syndrome is the
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degeneration of the vertebrae and the intervertebral discs in the Lumbar Region of the
spine (Silsoe research institute, 2000). As there is no blood flow to the Vertebra
themselves, the condition is non recoverable naturally. However, there are surgical
options to restrict the movement of damaged vertebra giving relief from the pain at the
cost of permanent reduced flexibility and movement.
Multiple studies have been carried out in order to accurately determine the exact
cause of pain. Determining the exact cause of pain has proved to be difficult, therefore a
large number of studies with multiple operators and machines were carried out.
S.Schwarze and G. Notbohm carried out one such experiment, in 1990. They
began to conduct an experiment trying to determine the response relationship between
exposure to whole body vibrations and Lumbar syndrome. The experiment lasted two
years while exposing 388 machine operators to varying degrees of vibration on a day-
to-day basis. X-rays of the lumbar region were taken at the start and end of the two-year
study and then compared. The focus on vibration measurement was not the frequency of
the vibrations, but the accelerations generated from the vibrations Figure 2.3, below
shows the number of participants, type of job and the percentage of operators for each
job experiencing Lumbar Syndrome.
FIGURE 2.3 PREVALENCE OF LUMBAR SYNDROME (SCHWARZE, 1998, PP.618)
Of note, this figure illustrates that out of 159 forklift operators chosen to take
part of the study, over 60% of them contracted some form of Lumbar Syndrome,
Schwarze’s experiment shows that the industrial limit to prolonged whole body
vibration in the vertical direction of 0.8m/s2 when exposed for eight hours a day was too
high and that reducing the acceleration by 0.2 m/s2 to 0.6 m/s2 for an eight hour work
period would greatly reduce the likelihood of contracting Lumbar Syndrome (Schwarze
and Notbohm et al., 1998, pp. 613-628).
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2.5 Design considerations for reducing whole body vibrations
Donati (2001) under took a comprehensive look at methods to reduce whole body
vibration effects when designing mobile machinery; his work states that there are two
main areas to focus on when designing mobile machinery;
• Insert suspension devices between the operator and the source of vibration
• Improve seat profiles, workstation ergonomics, visibility and cab dimensions
He outlines that the seat suspension is the only form of suspension that exists in
forklift trucks therefore a well-designed seat and suspension system is crucial to the
health and safety to the operator. It is stated in his research that documentation
providing technical specifications for seat suspensions from suppliers was non-existent.
Within his work a list was created outlining important parameters when designing a
suspension seat.
• Suspension damping must be sufficient to;
o Avoid amplification when the motion frequency is close to the seat
resonant frequency,
o Minimize suspension bottoming and topping due to transient motion.
• Weight adjustment;
o A suspension system is only effective when the seat is adjusted for a
specific weight of operator, therefore weight adjustment must be simple
and quick to perform.
• End-stop buffers;
o Suspension system should be fitted with top and bottom end-stop buffers
to prevent metal-to-metal contact when a suspension seat tops or bottoms
due to high-magnitude shocks.
Donati (2002) also outlines ISO standards that must be observed when designing
a seat suspension system (Donati, 2002, pp. 169-183).
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2.6 Discussion
Section 2.1 illustrated the issue of operator health and safety as a concern when
designing a prototype forklift seat and suspension system. It is imperative that the seat
suspension system designed meets the criteria mentioned above in order for the operator
to have the best experience.
Section 2.2, 2.3 and 2.4 detailed the current market range of suspension systems
and identified the main type of suspension system that applies to forklift trucks. This
report will focus on designing a “low profile” prototype seat suspension design as the
type of Forklift truck that the seat is being designed for is one that is capable of
operating off-road as well as on smooth terrain. The low-profile suspension category
should function well with any simulated shock loads applied in conjunction with
constant harmonic vibrations. It was also found that it is imperative that the seat
suspension works outside the range of 4-6 Hz, ensuring a maximum acceleration in the
vertical direction of 0.6m/s2 in order for the operator to be at a reduced risk of
contracting Lumbar Syndrome.
Finally section 2.5 outlines design considerations that need to be taken into
account for the design of the suspension to ergonomically meet the needs of the
operator. The design concepts and suspension system that were tested, either in reality
or by simulation, were constructed with generic spring and damper systems. With
today’s technology, mass production of polymers is becoming more prevalent than ever
before. There is a gap in research in the field of utilizing polymers as visco-elastic
vibration absorbers. Therefore, designing and testing will be carried out on a novel
suspension system concept that utilizes elastomers as the vibration absorbing material.
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3. Overview of standards investigated
3.1 Introduction From research carried out during the course of this project there are no
international standards for running FE simulations on concept designs before producing
a prototype. Therefore the standards referenced in this section have been used as the
baseline for the simulations ran in FE software. FE software is an excellent
development aid for designing, testing and developing concepts. It allows for various
design concepts to be base lined against each other to validate and streamline the design
concepts in advance of physical prototypes being built. However physical prototype
testing would be conducted to validate the FE software. Due to budget and time
restrictions manufacturing and testing a prototype will unfortunately not be carried out
as part of this project, however the simulations will create an excellent base point for
further work with upcoming projects.
3.2 ISO 3411 “earth moving machinery- Physical dimensions of operators and
minimum operator space envelope” Data for generating this standard for the operator sizes was generated from the
United States of America (CAESAR data), Europe (ISO 15534-3:2000) and Asia
(China, Japan, Korea and Thailand). The dimensions stated range from the 5th to the 95th
percentile of operator sizes combined from the countries stated above. Male and female
measurements are combined in this standard, measurements stated in the standard are
actual measurements, where specific measurements could not be obtained they were
derived by proportional scaling. Measurements stated in the standard show the operator
in an erect posture. Erect posture is defined in the standard as, standing or sitting upright
without a backrest.
Relevant information in this standard stated the dimensions of the operator in a
seated position; this information is vital in order to gain an understanding on how small
an operator may be. This information then determines the smallest seat depth of the
base, which in turn determines the maximum size of the suspension system, the
suspension system cannot have a bigger footprint then the seat. Figure 2, page 4 within
this standard is a labelled illustration of an operator in a seated position with a
corresponding table outlining a; small, medium and large operator.
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3.3 ISO 10326-1 “mechanical vibration – Laboratory method for evaluating
vehicle seat vibration” The basic laboratory requirements when testing vibration transmissibility from
the machine through a seat suspension system to the operator is defined in this standard.
The equipment needed to record vibration accelerations, is documented, however as this
test is carried out through FE software, the positions of the sensors was referenced to
gain results that can be compared to physical prototype testing. Section 8.1 of the
standard states that the simulated test vibration shall be specified in accordance with the
vehicle groups, defined by the time history of an actual and representative signal. The
application of the standard specifies the number of measured points, frequency,
amplitude spacing and the sampling rate.
When calculating the transmissibility at resonance of the seat for the damping
test outlined in 3.4.2 the following formula 3.1, is used;
! = !!(!!)!!(!!)
(3.1)
T = transmissibility
!!(!!)= Unweighted rms value of the measured vertical acceleration at the seat disk at
the resonance frequency
!!(!!)= Unweighted rms value of the measured vertical acceleration at the platform at
the resonance frequency
3.4 ISO 7096 “earth moving machinery – Laboratory evaluation of operator seat
vibration”
This standard was introduced to aid Engineer’s design and test seat suspension
designs that will be exposed to low frequency vibration of between 0-20Hz. Where the
vibrations are generated by movement of the vehicles over uneven ground. The design
of the seat is said to be a compromise between the requirements of reducing the effect
of vibration and shock on the operator and providing him/her with stable support so that
he/she can control the machine effectively. The standard states that the criteria provided
is what can be achievable using present design practice and that the criteria involved do
not ensure the complete protection of the operator against exposure to vibration and
shock. This standard obtained its input test methods from ISO 10326-1 outlined in
section 3.3 of this report.
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Test methods outlined are for physical prototype testing, where accelerometers
can be fitted to the device and physical results can be produced. As FE is the basics for
testing in this project, the test methods outlined were adapted to run on a finite element
simulation package to estimate how the concept will performance initial prototypes are
produced.
3.4.1 Test conditions
o Two tests were performed, firstly with a light operator mass of 52-55kg
and secondly a heavy operator mass of 98-103kg.
o Input vibrations were in accordance with ISO 10326-1 outlined in
section 3.3.
3.4.2 Testing
• Test one, Seat effective amplitude transmissibility (SEAT) factor
o The test is to last for a minimum of 180 seconds where the suspension
system is run through a range of frequencies from 0-20Hz. The standard
states that for each input spectral class the corresponding graph, figure 2-
10 in the standard, illustrates the target values to be produced at the base
of the seat for the simulated input vibration test.
o The test shall be deemed valid if the test configuration deviation is less
than +- 5% from the arithmetic mean for a minimum of three test runs.
• Test two, Damping Test
o The test seat is to be loaded with a mass of 75kg a sinusoidal base
excitation will then be applied ranging from 0.5 to 2 times the resonant
frequency of the suspension system.
o For the case of this project the resonant frequency will be identified by
means of a modal analysis trough FE software.
o The frequency sweep will be made over the course of 80 seconds with a
constant peak to peak displacement of 40% of the total suspension travel
or 50mm, whichever is smaller.
o Calculating the transmissibility at resonance is to be performed in
accordance with ISO 10326-1 outlined in section 3.3.
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3.5 ISO 11112 “Earth moving machinery – Operator’s seat- Dimensions and
requirements”
In this standard the focus is around the seat and not the suspension system,
however the standard outlines the minimum dimensions allowable when designing a
seat. Dimensions are specified for the width of the base of the seat. The length is not
specified, therefore when developing the minimum size for a seat footprint this standard
has to be combined with ISO 3411 outlined in section 3.2. Minimum and maximum
dimensions are outlined in figure 1, page 2 of this standard by means of a labelled
sketch and a corresponding table.
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4. Design Methodology
4.1 Introduction
Chapter 1, sections 1.2 and 1.3 outline the design brief and the general direction
in which the design follows. Chapter 4 details the direction and fundamental operation
of the design by taking the specific performance parameters, International standards,
ethical implications of designing a product for human interaction, manufacturing
processes, to name but a few into account. The subsections of this chapter outline the
design methodology used to help guide the design of the forklift seat suspension system.
Section 4.2 references the Product Design Specifications (P.D.S.), the P.D.S. is a
listing of design requirements, specifications and critical parameters of the concept
design. It forms the basis for the blueprint for the final product.
Section 4.3 explains the theoretical calculations carried out to find the spring and
damping coefficients required to adequately reduce the displacement transmissibility
and force transmissibility felt by the operator when operating the forklift truck.
Calculations were carried out for a range of operator masses and working frequencies as
defined by ISO-7096 – “earth-moving machinery - Laboratory evaluation of operator
seat vibration”.
Section 4.4 examines the initial novel concepts which were explored and details
the reasoning behind each concepts inspiration. Details of the advantages and
disadvantages of each concept are analysed. In turn each concept design is cross-
compared against each other in order to determine the best concept to further develop.
4.2 Product Design Specification (P.D.S.) The PDS is a document used by engineers to outline a product that is not yet
designed. It details what the product is intended to do. It does not specify the product
itself. It is used to ensure that the product which is to be designed meets with the needs
of the user. It is used as a boundary to ensure that engineers and designers stay within
the scope of the project. It details what the user requirements are and outlines the
functions required from the product. It sets out the limits to be considered during the
design. This ensures that the product, once designed meets with the users expectations
to ensure the sale of the finished product.
! ! Fork!Lift!Seat!Suspension!System!Design!!!
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4.2.1 Performance:
• Seat suspension should minimise the amplitude of vertical displacements of the
forklift truck transmitted through the suspension system to the operator.
• Loading on the seat suspension will be the weight of the seat and the weight of the
operator, operators weight will be obtained from standard ISO-7096 “earth-moving
machinery – laboratory evaluation of operator seat vibration”.
4.2.2 Economy:
• There was a limited budget for this project, therefore manufacturing a full-size
prototype was not preformed, in order to keep costs minimal all testing was carried
out on simulated FE model. A scale model was 3-D printed to aid in the
visualization of the design concept and to demonstrate the overall design
performance.
4.2.3 Quantity:
• One fully developed concept was modelled using Solidworks Premium Simulation
3-D modelling software, with one scale model to be 3-D printed in order to display
the concept for presentations and discussions.
4.2.4 Manufacturing facilities:
• The scaled model was manufactured in the University of Limericks workshop A0-
044.
4.2.5 Environment:
• The product must be designed to work as required in temperatures ranging from -
200C to 300C.
• The product will be exposed to damp, dirty and dusty conditions and should
function properly without any disturbance to operation.
4.2.6 Size:
• The vertical displacement of the seat suspension shall fall into the “Low profile”
category as detailed in section 2.2.
• The footprint of the suspension system will be smaller than the minimum size of
forklift seat allowed as dictated in ISO-11112 – “Earth moving machinery –
Operator’s seat – Dimensions and requirements”. The size envelope is specified in
section 4.4.1.
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4.2.7 Maintenance:
• Weekly visual inspections should be carried out on the elastomer to ensure that the
elastomer doesn’t show any physical signs of wear or damage.
4.2.8 Materials:
• Elastomeric materials were chosen to act as the vibration-dampening component
within the system. This material range was examined as a replacement for
traditional springs and dampers in an effort to reduce costs in manufacturing and
maintenance.
4.2.9 Ergonomics:
• The suspension system must reduce the vertical displacements transmitted from the
forklift trucks body through the seat to the operator.
• The suspension system must minimise the frequency and force by the operator when
the suspension experiences a shock base excitation.
4.2.10 Appearance:
• As the product is being designed for the industrial market aesthetics are not an
important factor, however a rubber safety skirt must be fitted around all moving
components in order to prevent the operator being potentially exposed to pinch
points.
4.2.11 Finish:
• All mild steel structural components are to be powder coated to reduce the exposure
to conditions that may induce corrosion.
• Pins that are exposed to surface ware are to be manufactured out of stainless steel to
ensure adequate corrosion resistance.
4.2.12 Industry standards:
• The product when fully developed must conform to the following standards in
relation to size, operator safety, vibration reduction and operator ergonomics. These
standards include but are not limited to;
• ISO-3411 – “earth-moving machinery - Physical dimensions of operators
and minimum operator space envelope”.
• ISO-7096 – “earth-moving machinery - Laboratory evaluation of operator
In order to find the displacement transmissibility, force transmissibility and
amplitude ratio, first a range of spring stiffness and damping ratios were calculated.
From ISO-7096, the range of masses used were stated for the 5th percentile and the 95th
percentile of forklift operators masses from 55kg to 103kg. Stated in section 2.3
research shows that a human spine and pelvic region has a resonance frequency of
between 4-9Hz therefore calculations were carried out for a range of frequencies from
1-20Hz with emphasis on moving the working frequency of the suspension system
away from the natural frequency of the spine. Knowing the range of frequencies and
masses used Equation 4.1 can be rearranged to find the corresponding spring stiffness
required for each design configuration Equation 4.2.
Freely vibrating combined mass of seat and operator
Base with harmonic excitation
Spring with stiffness K (N/m)
Dashpot with damping coefficient C (Ns/m)
+x
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!! = !! (4.1)
! = (!!!)(!) (4.2) k= Spring stiffness (N/m) m= mass (kg) ω!= Natural frequency (rad/s)
As a result of determining the spring stiffness is found for a range of masses and
frequencies equation 4.3 can be rearranged to find the damping coefficient associated
with each spring stiffness for a range of damping ratios equation4.4.
! = !! !" (4.3)
! = !2 !" (4.4) c= Damping coefficient (Ns/m) k= Spring stiffness (N/m) m= mass (kg) != Damping ratio
4.3.2 Displacement transmissibility
Displacement transmissibility is a ratio of the amplitude of the response of the
freely vibrating masses to the base motion. The displacement transmissibility graph
below, Figure 4.2 shows sample displacement transmissibility graph generated from
using Equation 4.5. The data for an operator mass of 55kg is used to generate the graph
using, a range of damping ratios from 0.1 to 2, giving a graphical representation of how
the suspension system should react theoretically over a range of frequency ratios. The
graph illustrates that for a frequency ratio of one, the resonant frequency of the system
the displacement transmitted through the system is largest. However outside of the
resonant frequency, the displacement transmitted reduces to below one, meaning that if
the operating frequency of the suspension system is greater than the resonant frequency
the displacement transmitted through the system will be less than the displacement of
the base excitation.
!! = [ !!! !" !
!!!"! !! !" !] (4.5) Y= Base excitation (m) X= Seat response (m) c= Damping coefficient (Ns/m) k = Spring stiffness (N/m) m = Mass (kg) ω= Operating frequency (rad/s)
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!
FIGURE 4.2 55KG HUMAN, NATURAL FREQUENCY OF THE PELVIS OF 4HZ
From examining the graph it can be seen that increasing the damping ratio can
reduce the amplitude of vibration transmitted through the system at resonance, however
the higher the damping ratio the less the displacement transmissibility is reduced when
the frequency ratio increases beyond one. Theoretical analysis consisted of generating
graphs for a range of natural frequencies of the pelvic region of the body for each
human test mass from 55kg to 105kg. These graphs are then compared with each other
to find the best overall spring and damper setup that can be used to ensure that the
operator experiences the best operating conditions.
4.3.3 Force Transmissibility
The force transmissibility is a method of calculating the force transmitted from
the base through the suspension system to the operator. Figure 4.3 outlines the force
transmitted from the base motion through the suspension and into the operator.
!FIGURE 4.3 55KG HUMAN, NATURAL FREQUENCY OF THE PELVIS OF 4HZ
M0.5!
0.5!
1.5!
2.5!
3.5!
4.5!
5.5!
0! 1! 2! 3! 4! 5!
Dis
plac
emen
t Tra
nsm
issi
bilit
y T
d
R=w/Wn
0.1!
0.4!
0.8!
1!
1.4!
1.8!
2!
0!
2!
4!
6!
8!
10!
0! 1! 2! 3! 4! 5!
Forc
e Tr
ansm
issi
bilit
y (B
ase
mot
ion)
R=w/Wn
0.1!
0.2!
0.4!
0.6!
0.8!
1!
1.2!
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Figure 4.3, was generated using Equation 4.6 rearranging it to find Ft, the force
transmissibility.
!!!" = !! !! !!" !
(!!!!)!! !!" !] (4.6)
Ft= Force transmissibility Y= Seat response (m) r= Frequency ratio k= Spring stiffness (N/m) != Damping ratio
From examining the graph, shown in Figure 4.3, it can be seen that increasing
the damping ratio can reduce the force transmitted at resonance but above an R-value of
2; where all force transmitted trough the system is equal, the lower the damping ratio
the less force is transmitted after this point. Theoretical analysis consisted of generating
graphs for a range of natural frequencies of the pelvic region of the body for each
human test mass from 55kg to 105kg. These graphs are then compared with each other
to find the best overall spring and damper setup that can be used to ensure that the
operator experiences the best operating conditions.
When comparing Figure 4.2 with Figure 4.3 it can be seen that designing a
suspension system that operates above the natural frequency of the human body is
paramount due to the magnitude of the forces and displacements transmitted when the
pelvic region resonates with the suspension system.
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4.4 Concept development The design process calls for many concepts to be initially sketched outlining
different ideas and variations of ideas to find a possible solution to the problem. The
concepts outlined are inspired from the design brief and current state of the art
technologies. Each concept is individually analysed discussing their function,
movement and validity towards solving the design problem. Before brainstorming
concepts it is important to gain an understanding of the size limitations and design
constrains. Outlined in section 1.2 the design brief states that the concepts must meet all
1 Suspension base 1 2 Elastomer holder base 2 3 Rear vertical swing arm 2 4 Front vertical awing arm 2 5 Elastomer holder top 2 6 L bracket base right 1 7 L bracket base left 1 8 Elastomer guide 2 9 Seat plate 1 10 L bracket top left 1 11 L bracket top right 1 12 Elastomer holding pin 2 13 Washer 20 14 Clevis pin 6 15 Short clevis pin 4 16 Split pin 10 17 Bushing 8
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5.4 3-D model overview of individual components In this section, the individual components of the seat suspension assembly are
described. Each component was designed with the manufacturing processes required to
produce each part in mind. The following sub-sections outline the components function
and manufacturing process. Also contained within the theses sub-sections are static FE
simulations carried out on each component in order to gain an understanding on if the
component in question is stiff enough to withstand the forces exerted on it without
plastically deforming. Complete working drawings of the suspension systems individual
components are shown in appendix B.
5.4.1 Suspension base
FIGURE 5.6 ISOMETRIC VIEW OF THE SUSPENSION BASE
3.4.1,A Component overview
The suspension base, Figure 5.6 is the component to which all other components
are assembled. This part design allows for multiple mounting points to the forklift truck
and a stiff platform for the rest of the components to be attached to. It is to be
manufactured from 4mm thick sheet steel. The flanges protruding up from the base are
pressed and folded during the manufacturing process and are to provide a secure
mounting point for the vertical swing arms outlined in section 5.3.3 and 5.3.4.
5.4.1,B Static FE analysis
Abaqus finite element software was used to conduct static, linear perturbation
analysis to gauge it the component is stiff enough to transfer the loads exerted by the
operator trough the system and into the forklift without excessive deformation.
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The forces exerted on the suspension base's come from the vertical swing arms
that are attached to the vertical flanges by means of pin loads though the mounting
holes, the boundary conditions applied to the component are fixed at the mounting point
holes where the suspension system would be bolted to the forklift truck. Therefore the
flanges are the main area of concern for this analysis. As the base is mirrored through
the Right Plane, analysis of the component can be performed on one half of the
component illustrated in Figure 5.7. This allows for the simulation to run faster as there
are half the elements to calculate.
FIGURE 5.7, HALF OF SUSPENSION BASE TO BE ANALYZED.
As ISO 7096 states that the maximum operator mass for testing should be 103kg
the load applied on each hole should be at least 103kg, as the position of the load cannot
be guaranteed to be distributed evenly over the suspension mounting points when the
suspension system is in service simulations were carried out simulating the worst case
scenario of, the total maximum applied load being exerted on each mounting point.
Therefore, a force of 1500N was applied vertically down to the bottom half of the
mounting holes for the vertical swing arms to simulate the weight of a heavy operator
and the seat weight. Multiple simulations were run, a tetrahedral mesh was applied
evenly over the area under investigation, with a number of mesh densities simulated. A
mesh convergence graph was then generated to ensure that the maximum stress
predicted by each of the simulations was consistent. Figure 5.8 illustrates the mesh
convergence graph, number of nodes is plotted on the x-axis with maximum stress
result felt on the system on the y-axis. The Von Mises stress theory was used. Von
Mises stress is a method of combining the three principle stresses in the x, y and z
planes into equivalent stresses, to determine the maximum stresses experienced by the
component.
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FIGURE 5.8 MESH CONVERGENCE GRAPH OF SUSPENSION BASE
Figure 5.8 is a good example of why a mesh convergence study was undertaken.
It illustrates that with a coarse mesh there may be areas of stress risers that may not
actually be the case. Therefore performing the mesh convergence study is vital; the
graph shows that the simulation predicts that the maximum stress educed in the
component is around 8.6x104N/m2which is below the yield stress of the material. The
material simulated was cold rolled steel and has a yield of 370MN/m2. Including a 50%
safety margin on top of the worst-case scenario the yield cannot be above 185x106N/m2.
Figure 5.9 (a) details an overview of the stresses found in the component while Figure
5.9 (b) illustrates a zoomed in image of how the force is transmitted throughout the area
of concern. The results of this simulation indicate that if manufactured this component
will be able to withstand the forces exerted on it.
FIGURE 5.9 (A) ILLUSTRATES OVERALL STRESSES INDUCED IN THE BASE, (B) ILLUSTRATES CLOSE UP OF THE
MAXIMUM STRESS INDUCED
0!
20000!
40000!
60000!
80000!
100000!
120000!
140000!
0! 5000! 10000! 15000! 20000! 25000!
Stress!N/m
2 !
Number!of!nodes!!
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5.4.2 Seat mounting plate.
FIGURE 5.10 VIEW OF SEAT MOUNTING PLATE
5.4.2,A Component overview
The seat mounting plate, Figure 5.10 mounts the seat to the suspension system,
the geometry and forces exerted on this component are similar to the Suspension base.
Therefore it is not necessary to perform simulations on this component, as the results
obtained should be similar to the suspension base.
5.4.3. Rear vertical swing arm
FIGURE 5.11 ISOMETRIC VIEW OF THE REAR VERTICAL SWING ARM
5.4.3,A Component overview
The rear vertical swing arm, Figure 5.11 connects the seat mounting plate to the
suspension base. This component design allows for the elastomer guide rail to slide
freely through this component without impeding its trajectory when the suspension is
compressed. Like the suspension base, the swing arms are to be manufactured from
4mm cold rolled steel. The front and rear vertical swing arms have the same
dimensions, however the rear swing arm has a slot cut out of it to allow the elastomer
guide to move through therefore FE analysis was carried out on the rear component
only as in theory this is the weaker component.
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5.4.3,B Static FE analysis
The forces exerted on the swing arms come from the seat mounting plate from
above, the boundary condition applied is a fixed pin condition through the bottom hole
on of the component. As with all the components tested, a mesh convergence study was
conducted to ensure a consistent maximum Von Misus stress result. The maximum
stress exerted on the component was predicted to be 1x105 N/m2 which is below the
yield stress of the material applied in the simulation. The material simulated was cold
rolled steel and has a yield of 370MN/m2. Including a 50% safety margin on top of the
worst case scenario the yield cannot be above 185x106N/m2. Figure 5.12 (a) illustrates
the dissipation of stress felt throughout the component with Figure 5.12 (b) showing the
point of maximum stress exerted on the component by the force applied.
FIGURE 5.12 (A) DISSIPATION OF STRESS THROUGHOUT THE COMPONENT, (B) MAXIMUM STRESS FELT BY
COMPONENT
5.4.4 Front vertical swing arm
FIGURE 5.13 ISOMETRIC VIEW OF THE FRONT VERTICAL SWING ARM
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5.4.4,A Component overview
The front vertical swing arm, Figure 5.13 connects the seat mounting plate to the
suspension base. This component design allows for the suspension to move in the
diagonal trajectory that it was designed to. Like the suspension base the swing arms are
to be manufactured from 4mm thick sheet steel. The front and rear vertical swing arms
generally have the same overall dimensions. Therefore as FE analysis was already
conducted on the rear swing arm, there is no need to run the same analysis on this
component.
5.4.5 Elastomer Holder
FIGURE 5.14 ISOMETRIC VIEW OF THE ELASTOMER HOLDER.
5.4.5,A Component overview
The elastomer holder, Figure 5.14 acts as end plates for the elastomer in the
assembly. The round shafts protruding from the centre of the component are press fitted
onto bushings that are free to rotate around the horizontal axis within the assembly, this
degree of freedom allows rotational movement needed to allow the suspension so move
during the compression and relaxation strokes of the suspension system. The hole in the
centre of the component is to allow for an elastomer guide to slide trough ensuring that
the elastomer compresses in one plane only. The elastomer holder is to be machined
from one solid billet of steel material complete working drawings of this component are
outlined in appendix B.
5.4.5,B Static FE analysis
The forces exerted on the component come from the elastomer compressing
against it. Boundary conditions applied to the component are fixed at the ends of the
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protruding cylinders. A mesh convergence study was carried out and it was predicted
that the maximum stresses exerted on the component is 3x108N/m2 which is below the
yield stress of the material applied in the simulation, Figure 5.15 illustrates the
dissipation of stress felt throughout the component.
FIGURE 5.15 ILLUSTRATION THE STRESS CONCENTRATIONS POINT
It can be seen from Figure 5.15 that the maximum stresses are where the
protruding shafts come into contact with the bushings at the point of maximum tension
and compression. This is an area to be looked into, as the stress is concentrated over a
small area. However as the simulated results show the maximum stress is still below the
yield point by a magnitude of 700 times less.
5.5 Elastomer selection
5.5.1 Introduction
The goals of the seat suspension system are to minimise the absolute
acceleration of the seat loaded by the operator and to minimise the displacement of the
seat relative to the forklift truck. These two goals work in opposition with each other
therefore a compromise must be achieved to both maintain the operators health and
allow the operator to retain control of the forklift truck. Calculations were carried out to
determine the spring stiffness and damping coefficients along with the static deflection,
required to reduce the displacement and force transmitted through the system for a
range of operator masses, as detailed in ISO-7096. Using the equations outlined in
section 4.3 a range of values were determined, the frequencies used to calculate the
spring stiffness ranged from 1Hz to 20 Hz, table 5.1 illustrates the spring constants
required for a selected range of operator masses and natural frequencies.