Seat Suspension Thesis

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.

! ! Fork!Lift!Seat!Suspension!System!Design!!!

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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.

! ! Fork!Lift!Seat!Suspension!System!Design!!!!

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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.

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Contents

ABSTRACT!...............................................................................................................................................!I!

ACKNOWLEDGEMENTS!.....................................................................................................................!II!

CONTENTS!............................................................................................................................................!III!

NOMENCLATURE!................................................................................................................................!VI!

LIST!OF!FIGURES!..............................................................................................................................!VII!

1.!INTRODUCTION!...............................................................................................................................!1!1.1!BACKGROUND!...................................................................................................................................................!1!1.2!DESIGN!BRIEF!..................................................................................................................................................!1!1.3!AIM!AND!OBJECTIVES!.....................................................................................................................................!1!1.4!OVERVIEW!OF!THESIS!....................................................................................................................................!2!

2.!LITERATURE!REVIEW!....................................................................................................................!3!2.1!INTRODUCTION!................................................................................................................................................!3!2.2!CURRENT!STATE!OF!THE!ART!SUSPENSION!CLASSIFICATIONS!................................................................!4!2.2.1&Overview&of&existing&LowSProfile&Suspension&Systems&Designs&..........................................&5!

2.3!RESONANT!FREQUENCIES!OF!THE!HUMAN!BODY!.......................................................................................!5!2.4!EFFECTS!OF!PROLONGED!EXPOSURE!TO!VIBRATION!.................................................................................!6!2.5!DESIGN!CONSIDERATIONS!FOR!REDUCING!WHOLE!BODY!VIBRATIONS!..................................................!8!2.6!DISCUSSION!......................................................................................................................................................!9!

3.!OVERVIEW!OF!STANDARDS!INVESTIGATED!.......................................................................!10!3.1!INTRODUCTION!.............................................................................................................................................!10!3.2!ISO!3411! “EARTH!MOVING!MACHINERYM!PHYSICAL!DIMENSIONS!OF!OPERATORS!AND!

MINIMUM!OPERATOR!SPACE!ENVELOPE”!........................................................................................................!10!3.3!ISO!10326M1!“MECHANICAL!VIBRATION!–!LABORATORY!METHOD!FOR!EVALUATING!VEHICLE!

SEAT!VIBRATION”!.................................................................................................................................................!11!3.4!ISO!7096!“EARTH!MOVING!MACHINERY!–!LABORATORY!EVALUATION!OF!OPERATOR!SEAT!

VIBRATION”!...........................................................................................................................................................!11!3.4.1&Test&conditions&......................................................................................................................................&12!3.4.2&Testing&......................................................................................................................................................&12!

3.5!ISO!11112!“EARTH!MOVING!MACHINERY!–!OPERATOR’S!SEATM!DIMENSIONS!AND!

REQUIREMENTS”!..................................................................................................................................................!13!

4.!DESIGN!METHODOLOGY!............................................................................................................!14!4.1!INTRODUCTION!.............................................................................................................................................!14!4.2!PRODUCT!DESIGN!SPECIFICATION!(P.D.S.)!............................................................................................!14!4.2.1&Performance:&.........................................................................................................................................&15!4.2.2&Economy:&.................................................................................................................................................&15!4.2.3&Quantity:&..................................................................................................................................................&15!4.2.4&Manufacturing&facilities:&..................................................................................................................&15!4.2.5&Environment:&.........................................................................................................................................&15!4.2.6&Size:&............................................................................................................................................................&15!4.2.7&Maintenance:&.........................................................................................................................................&16!4.2.8&Materials:&................................................................................................................................................&16!4.2.9&Ergonomics:&...........................................................................................................................................&16!4.2.10&Appearance:&........................................................................................................................................&16!4.2.11&Finish:&.....................................................................................................................................................&16!4.2.12&Industry&standards:&..........................................................................................................................&16!4.2.13&Testing:&..................................................................................................................................................&17!4.2.14&Safety:&.....................................................................................................................................................&17!4.2.15&Product&and&social&factors:&...........................................................................................................&17!


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4.3.3$Force$Transmissibility$.......................................................................................................................$20!4.4!CONCEPT!DEVELOPMENT!............................................................................................................................!22!4.4.1$Concept$size$envelope$........................................................................................................................$22!4.4.2$Concept$One$...........................................................................................................................................$23!4.4.3$Concept$two$............................................................................................................................................$24!4.4.4$Concept$three$........................................................................................................................................$25!

5.!CHOSEN!CONCEPT!REFINEMENT!AND!COMPONENT!DESCRIPTION!...........................!26!5.1!INTRODUCTION!.............................................................................................................................................!26!5.2!CONCEPT!REFINEMENT!...............................................................................................................................!26!5.3!3MD!MODEL!OF!SUSPENSION!SYSTEM!........................................................................................................!28!5.4!3MD!MODEL!OVERVIEW!OF!INDIVIDUAL!COMPONENTS!..........................................................................!30!5.4.1$Suspension$base$....................................................................................................................................$30!5.4.2$Seat$mounting$plate.$..........................................................................................................................$33!5.4.3.$Rear$vertical$swing$arm$..................................................................................................................$33!5.4.4$Front$vertical$swing$arm$..................................................................................................................$34!5.4.5$Elastomer$Holder$.................................................................................................................................$35!

5.5!ELASTOMER!SELECTION!..............................................................................................................................!36!5.5.1$Introduction$...........................................................................................................................................$36!5.5.2$Choosing$Elastomer$material$.........................................................................................................$40!

6.!VIRTUAL!TESTING!.......................................................................................................................!42!6.1!INTRODUCTION!.............................................................................................................................................!42!6.2!SEAT!FACTOR!...............................................................................................................................................!42!6.3!DAMPING!TEST!.............................................................................................................................................!43!6.3.1$Frequency$analysis$..............................................................................................................................$43!6.3.2$Determination$of$damping$performance$..................................................................................$45!

7!RESULTS!FROM!VIRTUAL!TESTING!.........................................................................................!46!7.1!INTRODUCTION!.............................................................................................................................................!46!7.2!SEAT!FACTOR!...............................................................................................................................................!46!7.2.1$Graphical$results$..................................................................................................................................$46!7.2.2$SEAT$factor$calculation$....................................................................................................................$47!

7.3!DAMPING!PERFORMANCE!...........................................................................................................................!49!

8!3PD!PRINTED!MODEL!...................................................................................................................!50!

9!DISCUSSION!.....................................................................................................................................!52!9.1!INTRODUCTION!.............................................................................................................................................!52!9.2!OVERVIEW!OF!DESIGN!.................................................................................................................................!52!9.3!SIMULATION!RESULTS!.................................................................................................................................!53!9.3.1$SEAT$factor$.............................................................................................................................................$53!9.3.2$Damping$performance$......................................................................................................................$53!

9.4!RECOMMENDATIONS!FOR!FUTURE!WORK!................................................................................................!54!9.4.1$Elastomer$selection$.............................................................................................................................$54!9.4.2$Concept$manufacture$........................................................................................................................$54!9.4.3$Cost$effective$examination$...............................................................................................................$55!

10!CONCLUSIONS!..............................................................................................................................!56!!! !


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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.


! 11!

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.


! !12!

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.


! 13!

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.


! !14!

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.


! 15!

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.


! !16!

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

seat vibration”.


! 17!

• ISO-11112 – “Earth moving machinery – Operator’s seat – Dimensions and

requirements”.

4.2.13 Testing:

• During this project testing was carried out through simulated versions of the ISO

standard, ISO-7096 – “earth-moving machinery - Laboratory evaluation of operator

seat vibration”.

4.2.14 Safety:

• The concepts design will incorporate a protective skirt to stop the operator

accidentally being injured in the mechanism, the concepts design should reduce the

occurrence of bottoming and topping of the seat reducing the risk of injury to the

operator while operating the forklift truck.

4.2.15 Product and social factors:

• The social factors regarding this concept would be the inclusion of all weights of

operators ISO-7096 – “earth-moving machinery - Laboratory evaluation of operator

seat vibration”, outlines the test weights of operators from the 5th percentile to the

95th percentile. Therefore any potential claims that the suspension system was not

designed for inclusion of all sizes of people are avoided.

4.2.16 Design time:

• The project design time started on the 11 of September 2013 and continued until the

project's completion on the 21 of March 2014.


! !18!

4.3 Theoretical calculations Before any concepts were drawn up theoretical calculations based on ISO-7096

– “earth-moving machinery - Laboratory evaluation of operator seat vibration” were

carried out. The spring and damping coefficients required to adequately reduce the

displacement transmissibility and force transmissibility felt by the operator when

operating the forklift truck for a range of operator masses and operating frequencies.

For the initial calculations the problem can be seen as a single degree of freedom

forced vibration problem with the mass of the operator and seat seen as the one entity

with a spring and dashpot connecting the freely vibrating mass to the base that is

harmonically excitated, Figure 4.1 shows a simplified diagram describing the

relationship between the base and the freely vibrating mass.

!!!!!!!!!!!!

FIGURE'4.1'ONE'DEGREE'OF'FREEDOM'SYSTEM'WITH'BASE'EXCITATION.'

!4.3.1 Spring and damping coefficient

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


! 19!

!! = !! (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)


! !20!

!

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!


! 21!

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.


! !22!

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

international standards.

4.4.1 Concept size envelope

From investigating ISO-11112 – “Earth moving machinery – Operator’s seat –

Dimensions and requirements” and ISO-3411- “Earth moving machinery – Physical

dimensions of operators and minimum operator space envelope” a sketch of the

minimum size of seat foot print can be generated, figure 4.4 outlines the minimum seat

foot print allowed to conform with the standards.

Each concept suspension system brainstormed has a footprint smaller than these

minimum dimensions in order for the suspension system to be compatible with all types

of standard conforming seats. Having the suspension system smaller than the seat will

not create protrusions leading to potential health and safety issues with creating

catching points.

300mm!

420mm!!

430mm!FIGURE'4.4'MINIMAL'SEAT'FOOTPRINT'DIMENSIONS.


! 23!

4.4.2 Concept One

This concepts inspiration is based on a double wishbone type car suspension;

figure 4.5 illustrates a simplified sketch of the proposed idea. The suspension system

utilises a common mechanical spring damper system. The parallelogram design, more

clearly shown in figure 4.5 (b) of the linkages ensures that the seat when bolted to the

top of the suspension system remains level throughout a compression cycle of the

system.

!FIGURE'4.5CONCEPT'ONE'SKETCHES,'(A)'3D'SKETCH'ILLUSTRATING'THE'DESIGN'IDEA'(B)'2D'SIDE'VIEW'ILLUSTRATING'THE'

DIRECTION'OF'SEAT'TRAVEL'

This concept uses tried and tested technologies that are similar to products on

the market today, with the spring damper components for sale on the market have

documentation on maintenance schedules and safe working conditions. Therefore these

concept do not fall under the criteria of a new novel suspension system, however the

motion of the seat travel is not common. Most of the products on the market dissipate

the input vibration energy by only allowing movement in the vertical plane. The

diagonal range of motion shown in 4.5 (b) increases the overall length of stroke and

may reduce the occurrence of bottoming and topping due to shock loads.

Direction!of!seat!travel!


! !24!

4.4.3 Concept two

This concepts inspiration is based on the Thudbuster bicycle seat post

suspension system; figure 4.6 illustrates a simplified sketch of the proposed idea. The

suspension system elastomeric materials that act as a combined spring damper system.

Using an elastomeric material, as a combined spring and damper would reduce

production costs compared to the more traditional mechanical spring and dampers.

!FIGURE'4.6'CONCEPT'TWO'SKETCH'ILLUSTRATING'THE'USE'OF'ELASTOMERS'AS'A'COMBINED'SPRING'DAMPER.

This solution also takes inspiration from concept one utilising a modification of

the parallelogram design. The design ensures that when the seat is attached it remains

level throughout the full compression and relaxation cycle of the system. This concept

also employs the same diagonal range of motion seen in concept one allowing for a

greater overall seat displacement over the same vertical distance as a traditional

suspension system.

The elastomeric material used in the inspiration for this concept is a

polyurethane polymer with a shore hardness of between 40a -50a. Analysis and testing

would have to be conducted on different elastomer compounds to obtain the correct

elastomer that would have the mechanical properties needed to obtain the desired

displacement transmissibility.


! 25!

4.4.4 Concept three

The illustration, shown in figure 4.7 investigates the idea of a back mounted

suspension system, this system incorporates rollers mounted to slide rails that allow for

displacement in the vertical direction, the sketch omits a mirrored suspension design for

the left hand side of the seat. In addition to the rear seat mount for the suspension

system a base mounted spring damper would have to be installed the help reduce the

bending moment that would be applied to the vertical mounts when the operator is

seated in the seat.

!FIGURE'4.7'CONCEPT'THREE'ILLUSTRATING'IDEA'OF'A'BACK'MOUNTED'SEAT'SUSPENSION'SYSTEM

This system, although novel, limits the performance of the seat’s operation. A

rear-mounted system would limit the operator’s adjustment options by restricting the

seat back to a vertical position. From studying the market on seat options available for

purchase to retrofit to suspension systems, all seat options available had no option to

mount the suspension system to its back. Initial hand calculations to find the bending

moment and shear stress on the vertical component (A) in figure 4.7 proved that, for a

rear mounted system. Extensive bracing would have to be installed in order for the

structure to be stiff enough to withstand bending.

With all the concepts, additional safety measure has to be put in place to avoid

injury to the user as a result of pinching from the spring or catching on moving

components.

Rear!seat!mounts!

Vertical!component!(A)!Omitted!from!illustration!vertical!component!in!front!of!rollers

unts!

Base!mounted!spring!damper!

Spring!damper!

Rollers!


! !26!

5. Chosen concept refinement and component description

5.1 Introduction Having analysed the different concepts outlined in section 3.4. The chosen

concept takes inspiration from concept two outlined in section 3.4.3. This idea was

selected as it carried the greatest potential to meet the design brief and it has meet the

requirements outlined in the standards relating to seat design. The use of an elastomeric

material as the spring and damping component along with the suspension vector

displacement makes this concept novel and self-contained. Investigating to find an

elastomeric material that would operate like a traditional spring damper system and

designing a frame that will be stiff enough to withstand the repeated vibration forces

was the focus of this project.

Section 5.2 highlights just some of the design considerations and decisions made

when finalising the concept to be modelled, 5.3 introduces the 3-D modelled assembly

and 5.4 outlines the main individual components modelled.

5.2 Concept refinement

Following on from section 4.4.3’s concept, more in depth design detailing was

required. This section highlights some areas where two design ideas could have been

used to refine the design of the concept to be tested.

Below figure 5.1(a) and (b) show and end view sketches of two design concepts

for the placement of the elastomers. Figure 5.1 (a) Illustrates the use of one elastomer

extending the length of the design. Figure 5.1 (b) Illustrates an idea of using two

separate elastomers, one connected to each vertical swing arm.

In order to help understand the geometry of the two sketches better a simple

wooden model was created, figure 5.2 (a.1),(a.2), (b.1) and (b.2) show how the

supporting parallel structure performs with both elastomer setups.

FIGURE'5.1' (A)'ONE'ELASTOMER' (BLUE)'CONNECTING'BOTH'SWING'ARMS,' (B)'TWO'ELASTOMERS'CONNECTING'SWING'ARMS'TO'THE'

BASE


! 27!

FIGURE'5.2'(A.1)'SINGLE'ELASTOMER'AT'REST,'' ' (B.1)'TWO'ELASTOMERS'AT'REST,'

As both sketches and images show the main structure supporting the seat is a

parallelogram, for the structure to compress when load is applied both elastomers in

figure 5.2 (b) would have to compress equally. From examining the both types of

elastomer positions, it was felt that having two elastomers, figure 5.2 (b) did not

improve the potential performance of the suspension system, but only increased the

amount of parts required in the assembly. Therefore the concept, figure 5.2 (a) with only

one elastomer is used was chosen.

Once the elastomer design was chosen the next design problem was how to fix

the elastomer holder to the suspension system that allowed the elastomers to rotate

relative to the swing arms as the suspension system operated. Two main ideas were

considered illustrated in figure 5.3 (a) and (b), method (a) focuses on the idea of having

two separate components for the vertical swing arms with a solid machined elastomer

holder held in place between the two arms with bushings. Method (b) focuses on the

idea of one folded sheet metal part forming the swing arms. It can be described as

having the same function as method (a’s), but more ridged.

'(A.2)'SINGLE'ELASTOMER'FULLY'COMPRESSED,' ' (B.2)'TWO'ELASTOMERS'FULLY'COMPRESSED.

FIGURE'5.3,'(A)'TWO'COMPONENT'SWING'ARM,'' ' (B)'ONE'FOLDED'SWING'ARM.


! !28!

Due to the nature of the suspensions operation, a ridged design is vital to

maintain high reliability. Therefore method (B) was selected for the design of the swing

arms. The elastomer holder described in 5.4.5, is mounted to the swing arm by means of

bushings. Method (b)’s design is described in more detail in section’s 5.4.3 and 5.4.4.

5.3 3-D model of suspension system

In order to perform the dynamic vibration simulations a 3-D model was produced the

following sections and sub-sections outline the suspension system as a whole as well as

each individual component along with static stress evaluations on them. Figure 5.4 and

table 5.1 illustrates an exploded view with component list of the suspensions system

assembly. The system comprises mainly of components that are pressed and folded

from 4mm cold rolled steel. Each component has a full set of working drawings along

with a bill of materials in appendix A outlining material type and dimensions.

FIGURE 5.4 EXPLODED VIEW OF SEAT SUSPENSION CONCEPT TO BE TESTED

!!!!!!!!


! 29!

!!

TABLE''5.1'BILL'OF'MATERIALS'FOR'FIGURE'5.4'

!!!!

Figure 5.5 (a) and (b) illustrates the length of stroke and the range of motion that

this concept delivers. With this concept, for the seat pan’s vertical displacement of

50mm there is a greater vector displacement of 60mm. Therefore reducing the

occurrence of bottoming and topping when compared to a seat suspension system that

has the same vertical displacement stroke.

!

!

!

!FIGURE'5.5'(A)'END'VIEW'OF'SUSPENSION'FULLY'AT'REST,'' (B)'END'VIEW'OF'SUSPENSION'FULLY'COMPRESSED.'

Number Component name Quantity

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


! !30!

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.


! 31!

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.


! !32!

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!!


! 33!

5.4.2 Seat mounting plate.

FIGURE 5.10 VIEW OF SEAT MOUNTING PLATE


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


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.


! !34!


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


! 35!


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.


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.


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


! !36!

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.


! 37!

TABLE'5.1'SPRING'STIFFNESS'(K)'CALCULATED'FOR'SELECTION'OF'FREQUENCIES'AND'OPERATOR'MASSES'

As detailed in section 2.3, it is known that the resonant frequency of the human

Lumbar region is between 4-9Hz. Table 5.1 shows that for an operator mass of 55kg,

operating between 4-9 Hz the suspension system with spring stiffness of between

34.7KN/m and 175.8KN/m may begin to resonate with the lumbar region of the

operator. As determining the spring stiffness is dependent on the operator mass, the

same calculations were carried out to determine the spring stiffness with an operator

mass of 103kg that may induce resonance with the lumbar region. Once the spring

stiffness that may induce resonance with the operator are known they can be eliminated

from the investigation. Therefore the elastomeric material chosen to act as the spring

will have spring stiffness of between 335.8KN/m and 1658.1KN/m.

Damping constants for masses ranging from 55kg to 105kg and frequencies

above 9Hz were calculated using equation 4.4 presented in section 4.3.1. From studying

displacement transmissibility against frequency ratio graphs, similar to figure 4.2

generated for operator masses of 55kg, 65kg, 75kg, 85kg, 95kg and 103kg it was

determined that the optimum damping ratio for the system would be 0.4. The

displacement transmissibility graph figure 5.17, illustrates how the suspension system

performs when the damping ratio is set at 0.4 for each operator mass with an operational

frequency of 4Hz. It shows that at resonance the maximum seat displacement is 1.6

times the displacement of the base. Past resonance there the displacement of the seat

reduces to between 0.2 and 0.4 times the displacement of the base, giving the operator a

smoother ride then the forklift truck is experiencing. A damping ratio of 0.4 gives an

Natural Frequency Operator 55kg Operator 75kg Operator 103kg

Hz k=KN/m K=KN/m K=KN/m 1 4.8 6.6 9.3 4 34.7 47.3 66.3 6 91.7 125.1 149.2

9 175.8 239.8 335.8 10 217.1 296.1 414.5 15 488.5 666.2 932.7 20 868.5 1184.3 1658.1


! !38!

adequate compromise of displacement excitation at resonance to displacement reduction

after the resonant frequency.

!FIGURE'5.16'DISPLACEMENT'TRANSMISSIBILITY'AGAINST'FREQUENCY'RATIO,'WITH'DAMPING'CONSTANT'OF'0.4

From table 5.1 and figure 5.16 values for calculating the damping coefficient for

each spring stiffness between335.8KN/m and 1658.1KN/m can be calculated using

equation 4.4 in section 4.3.1 Table 5.2 outlines presents the spring stiffness with their

corresponding damping coefficient to create a suspension system with a damping ratio

of 0.4. TABLE'5.1'SPRING'STIFFNESS'AND'DAMPING'COEFFICIENT'FOR'DAMPING'RATIO'OF'0.4'

Hooke’s Law states that for a Linear material, like a coil spring of a constant

diameter, pitch and thickness, the force with which a spring pushes back with is linearly

proportional to the distance from its equilibrium length. Therefore for a common coil

spring, adjustment for the operators weight is needed in order to gain the optimum

performance of the suspension system. One of the goals trying to be achieved

throughout this thesis is to design a user-friendly suspension system with minimal

operator adjustment required to gain optimum performance. Determining an elastomeric

0!0.2!0.4!0.6!0.8!1!

1.2!1.4!1.6!1.8!

0! 1! 2! 3! 4! 5! 6!Displacem

ent!Transmissability!!

R!

55kg!

65kg!

75kg!

85kg!

95kg!

105kg!

Natural Frequency Spring stiffness Damping coefficient

Hz KN/m KNs/m 9 335.8 4.7 10 414.5 5.2 15 932.7 7.8 20 1658.1 10.5


! 39!

material with non-linear spring stiffness eliminates the need for the operator to adjust

the spring preload for their mass.

Inspiration for the chosen concept, originally outlined in section 4.4.3 of this

report, the “Thudbuster” bicycle seat post suspension system utilises an elastomer for

the vibration absorption. From contacting Ryan McFarland the Founder of RJ Concepts,

Inc. the company that engineered and sells the “Thudbuster”. It was discovered that the

elastomers used are of a cylindrical shape manufactured from a Polyurethane material

and has a durometer range of between Shore 40A to 50A. RJ Concepts were unwilling

to give more detail regarding the mechanical properties and non-linearity of the

elastomer stiffness during compression, however it was felt by the company that the

required performance parameters could be achieved with the dimension outlined for the

elastomer. The dimensions of the elastomer design chosen for this concept are;

• Internal diameter = 0.15m

• Outer diameter = 0.45m

• Length = 0.2m

From these dimensions and determined spring stiffness a Young’s Modulus can be

calculated for the elastomeric material in question with the use of equation 5.1.

! = !"! (5.1)

E = Young’s Modulus (N/m2)

K= Spring stiffness (N/m)

L= Length of material (m)

A= Cross sectional area (m2)

For each range of spring stiffness’s the Young’s modulus changes the table 5.3 below

details that as the spring stiffness increases the Young’s modulus of the material also

increases. TABLE'5.2'COMPARING'SPRING'STIFFNESS'WITH'YOUNG'S'MODULUS'

Natural Frequency Spring stiffness Young’s Modulus

Hz KN/m MN/m2 9 335.8 47 10 414.5 58.6 15 932.7 131.9 20 1658.1 192


! !40!

5.5.2 Choosing Elastomer material

Mathematical models exist to quantify the visco-elastic behaviour of an

elastomer from deflection experiments. Two such models include the Maxwell model

that defines the elastomeric material as a spring and damper in series and the Kelvin

model that defines the elastomeric material as a spring and damper in parallel. Both

models use a spring to define the instantaneous deflection of the material and the

dashpot to represent deflection over time. Figure 5.17 displays the mathematical

model’s graphical representation.

!'

'

FIGURE'5.17'(A)'MAXWELL'MODEL,'' ' ' ' (B)'KELIVN'MODEL'

The Maxwell model’s governing equation is presented below as equation 5.2 the

data for stress, strain and time are generated from experimental measurements for

specific material composition. (Crawford R.J., 2005)

!"!!" = !!!!"!" +

!! (5.2)

T = relaxation time constant (s)

t = Time (s)

! = Stress (N/m2)

! = Strain

! = Damper

! = Spring Modulus

!=!Spring!Modulus!!

!=!Damper!!

!=!Stress!applied!!


! 41!

The Kelvin model’s governing equation is presented below as equation 5.3 the

data for stress, strain and time are generated from experimental measurements for

specific material composition. (Crawford R.J., 2005)

! = !" + ! !"!" (5.3)

t = Time (s)

! = Stress (N/m2)

! = Strain

! = Damper

! = Spring Modulus

Neither model on their own can accurately predict the deformation behaviour of

the material being tested, therefore when testing the elastomer in question a

combination of the two models must be carried out, where measurements can be

recorded defining the materials non-linear compression response the loading applied to

it.

Contact was made with Silicone Engineering Ltd in England enquiring if sample

elastomeric material could be manufactured and supplied for purpose of testing to find

an elastomer capable of meeting the requirements of the design. Unfortunately time

restrictions of the project did not allow for samples to be manufactured and tested in

time. More detail on testing that could be conducted is mentioned in chapter 9 section

9.4 as part of future work for continuation of this project.


! !42!

6. Virtual testing

6.1 Introduction Investigation into full prototype testing is outlined in ISO- 7096:2000, this

standard was used as the base line to run simulations. Static analysis and dynamic

simulations run on FE software. As discussed in section 5.4, static analysis of

component stiffness was carried out using Abaqus CAE 6.12. These static analysis’s

shows that theoretically each component is stiff enough to withstand the loads applied

to each component individually. The FE package used to run the dynamic simulations

was Solidworks Simulation Premium 2012. This FE package was chosen to run the

simulations as it is an industrially recognised package, Engineers in industry use FE for

dynamic testing of concept components. The analysis's preformed on the assembly's

were frequency analysis, motion analysis and harmonic loading analysis.

6.2 SEAT factor As previously stated in section 3.4.2, determination of the SEAT factor or Seat

Effective Amplitude Transmissibility factor is carried out on a full scale prototype

manufactured from materials specified for the final design. The test is to be performed

with an operator in position two tests have to be performed. One with a light operator

from between 52kg-55kg and one with a heavy operator 98-103kg. Simulations carried

out in the FE software simulate the operator masses as a distributed load across the

suspension platform. The spectral class of the suspension system defines vibration

input; the category for this concept is Earth Moving 9 (EM 9). The input vibration input

used is in accordance with ISO 10326-1, 8.1. The SEAT factor is determined by

equation 6.1

!"#$ = ! !!!!"!!!!" (6.1)

awS12 = Weighted Root mean Squared (RMS) value of the measures vertical

acceleration at the connection between the seat and the suspension system.

awP12 = Weighted Root mean Squared (RMS) value of the measures vertical

acceleration at the suspension base.

The criteria for passing this test is for EM 9 the SEAT factor is to be less than 0.9.


! 43!

The input vibration to the system should be generated from a time history of an

actual and representative signal (ISO 10326-1:1992, 8.1). However as the tester had no

means of gathering this data, the testing input vibration for the damping test was

substituted for the input vibration to determine the SEAT factor as outlined in section

3.4.2. This input vibration was chosen as this input vibration simulates the worst case

operation conditions that the seat is being designed for. The input vibration is a

sinusoidal vibration at intervals of 1Hz from 1-20Hz and at the resonance frequency of

the system as determined in section 6.3.1, with a base displacement of 40% of the total

suspension travel. Results for this test can be found in chapter 7 section 7.2.

6.3 Damping Test

The damping test specified in ISO 7096:2000 determines the damping

performance at the suspension systems resonant frequency.

6.3.1 Frequency analysis

A frequency analysis was conducted to ensure that the resonant frequency of the

assembly did not coincide with either the operational frequencies of the forklift truck or

the natural frequency of an operator, because of the vibration loading applied to the

base of the seat is harmonic it is vital that the natural frequencies of the suspension and

the operational frequencies of the forklift truck do not coincide, if they equal resonance

may occur which may lead to excessive loading on the system ultimately resulting in

catastrophic failure.

When performing the Damping test section 5.4.2 in ISO-7096:2000, the

resonant frequency of the system must be determined before carrying out the test. The

method outlined in the standard determines the resonance frequency by conducting an

linear frequency sweep in maximum steps of 0.05Hz, from 0.5 times the expected

resonant frequency to 2 times it. For each adjustment in the frequency sweep the test

duration must be a minimum of 80 seconds. However this method of determining the

natural frequency of the system cannot be applied to CAD model, the natural frequency

of the suspension system was determined by conducting a frequency analysis.

Solidworks Simulation performs the frequency analysis by utilising the Eigen-value

approach to determine the resonant frequency of the system. Due to the fact that the

resonant frequencies of a component are proportional to the material that the component


! !44!

is manufactured out of, it is important to select a material that closely resembles the

material properties that the prototype will eventually be manufactured out off. Table 6.1

below illustrates the material used and its properties.

TABLE'6.1'PROPERTIES'OF'COLD'ROLLED'STEEL'

Property Value

Young’s Modulus Mass Density

Poisson's Ratio

210x109 N/m2

7900 kg/m3

0.29

Once the material was applied to the components, boundary conditions were

applied to simulate points anchoring of the suspension system to the forklift truck. A

fixed boundary condition was applied to the base of the system, this allowed the

software to simulate bolt mounting points and contact between the suspension base with

the fork lift truck bed. To increase the job processing time virtual connecting pins that

allow joint rotation replaced modelled pins in the assembly. A job was created and

submitted to the solver to determine the first four natural frequencies and mode shapes.

Figure 6.1 (a) illustrates the boundary conditions, virtual connecting pins and virtual

springs applied to the system, figure 6.1 (b) outlines the mode shape of the suspension

system at its first natural frequency.

• Blue represents no movement within the model,

• Red represents maximum movement within the model.

These colour changes represent an exaggerated illustration of the potential

deformation within the model when the model is resonating.

!FIGURE'6.1'(A)'BOUNDARY'CONDITIONS,'PIN'CONNECTIONS'AND'VIRTUAL'SPRINGS,''(B)'MODE'SHAPE'1'AT'FIRST'NATURAL'

FREQUENCY'

!


! 45!

Table 6.2 lists the first four natural frequencies of the full suspension system the

frequencies are displayed in Hertz, It can be seen that there is not a constant relationship

between the natural frequencies making them difficult to predict.

TABLE'6.2.''FIRST'FOUR'NATURAL'FREQUENCIES'

Natural Frequency Number 3-D Model (Hz) 1 212.94 2 293.33

3 305.50

4 369.67

6.3.2 Determination of damping performance

From determining the first resonant frequency, section 5.4.2 of ISO 7096:2000

specifies that a load of 75kg is applied to the seat plate and the base has to be excited by

a sinusoidal vibration. "The damping test and the calculation of the transmissibility

H(fr) at resonance shall be performed according to ISO 10326-1:1992, 9.2. In all cases,

the damping test itself at the resonance frequency shall be carried out with a peak to

peak displacement of the platform of 40% of the total suspension travel even if the 40%

value exceeds 50mm.Only one measurement needs to be carried out at the resonance

frequency of the seat suspension." (ISO 7096 “earth moving machinery – Laboratory

evaluation of operator seat vibration"). The total vertical peak-to-peak displacement of

the modeled concept is 50mm; therefore the sinusoidal vibration will have a peak-to-

peak displacement of 20mm at a frequency of 212.94 Hz. Results from theses

simulations can be found in chapter 7 section 7.3.


! !46!

7 Results from virtual testing !7.1 Introduction

! The results obtained for the dynamic simulations described in Chapter 6 with the

aim of passing a simulated version of ISO 7096:2000 are analysed in this chapter. Static

simulations to determine the stiffness of each component can be found in chapter 5,

section 5.4. Modal analysis to determine the theoretical natural frequency of the

suspension system are illustrated in chapter 6 section 6.3.1. A graphical representation

of the displacements and accelerations felt by the seat plate and suspension base as

defined in ISO 7096, figure 7.1 details the seat plate and suspension base. Results in this

section will be discussed in chapter 9.

!FIGURE 7.1 OUTLINING THE SEAT PLATE AND SUSPENSION BASE

!7.2 SEAT factor

7.2.1 Graphical results

Simulations were run for a period of 180 seconds as defined in ISO-7096, the

aws12 and awp12 were determined by generating vertical linear acceleration plots on the

seat plate and the suspension base. Figure 7.2 illustrates the vertical linear acceleration

of the seat plate when a forcing function is applied to the system. The graph, figure 7.2

is only plotted for the first five seconds of the motion simulation. When examining the

plot for 180 seconds it was found that the acceleration felt on the plate were cycling

with the same plot trace as the first five seconds, therefore only a section of it needs to

be included in this section.

Seat plate

Suspension Base


! 47!

FIGURE 7.2 LINEAR ACCELERATION OF SEAT PLATE INPUT VIBRATION 16HZ

Figure 7.3 below, illustrates the vertical linear acceleration plot of the

suspension base when the forcing function is close to the resonant frequency of the

system for the first five seconds of the 180 second test, as with figure 7.2 the

acceleration felt on the base continued to generate a cyclic response.

!FIGURE 7.3 LINEAR ACCELERATION OF SUSPENSION BASE INPUT VIBRATION 16HZ

7.2.2 SEAT factor calculation

! Utilising figure 7.2 and 7.3 the root mean square of the seat plate and suspension

base were calculated. Substituting the results into equation 6.1 the SEAT factor was

determined for a suspensions systems with a variety of spring stiffness, damping

coefficients and working frequencies. Table 7.1 displays the calculated SEAT factor for

a spring stiffness of 110KN/m and damping coefficient of 450 N/(m/s) over a range of

working frequencies.

M2.5000E+02!

M2.0000E+02!

M1.5000E+02!

M1.0000E+02!

M5.0000E+01!

0.0000E+00!

5.0000E+01!

1.0000E+02!

1.5000E+02!

0.000! 0.500! 1.000! 1.500! 2.000! 2.500! 3.000! 3.500! 4.000! 4.500! 5.000!A

ccel

erat

ion

(m/s

^2)

Time!(sec)!

M1.5000E+02!

M1.0000E+02!

M5.0000E+01!

0.0000E+00!

5.0000E+01!

1.0000E+02!

1.5000E+02!

0.000! 0.500! 1.000! 1.500! 2.000! 2.500! 3.000! 3.500! 4.000! 4.500! 5.000!

Acc

eler

atio

n (m

/s^2

)

Time!(sec)!


! !48!

!!!!!!!!!!!!!!!!!!!!!!!!!!

The passing criteria for an EM 9 seat suspension system the SEAT factor must be less

than 0.9 for its working frequency. From table 7.1 it can be seen that the suspension

system passes when the working frequency of the seat is greater than 13Hz.

!

FIGURE'7.1'CALCULATED'SEAT'FACTOR'

Operator mass

kg

Operating frequency

HZ SEAT factor

103 1 0.29 103 2 0.964 103 3 1.029 103 4 1.75 103 5 1.76 103 6 1.2 103 7 1.1 103 8 0.89 103 9 1.5 103 10 1.29 103 11 1.21 103 12 1.07 103 13 0.91 103 14 0.83 103 15 0.76 103 16 0.86 103 17 0.77 103 18 0.75 103 19 0.76 103 20 0.77

103 Resonance 0.92


! 49!

7.3 Damping performance As determined through modal analysis the natural frequency of the suspension

system was determined to be 212.94Hz. Figure 7.4 and 7.5 illustrate the graphical

response of the linear vertical acceleration of the damping performance tests. It can be

seen that there is a regular amplitude build up and reduction over time this type of

vibration response is called beating, and will be discussed in chapter 9.

!FIGURE 7.4 SEAT PLATE ACCELERATION RESPONSE TO FORCING FUNCTION ALMOST EQUAL TO RESONANCE

!

!!

FIGURE 7.5 SUSPENSION BASE ACCELERATION RESPONSE TO FORCING FUNCTION ALMOST EQUAL TO RESONANCE

!

According to ISO 7096 the damping test is only to be carried out once, at the

resonant frequency. The Root Mean Square (RMS) of the acceleration response to the

input vibration were calculated and the damping performance was carried out using

equation 3.1 and worked out to be;

Damping transmissibility = 0.97

The passing criteria for the damping transmissibility for an EM 9 classification

of suspension system is less than "2" therefore the suspension passes the damping

performance test.!!!

M2.0000E+04!

M1.0000E+04!

0.0000E+00!

1.0000E+04!

2.0000E+04!

0.000! 0.500! 1.000! 1.500! 2.000! 2.500! 3.000! 3.500! 4.000! 4.500! 5.000!

Acc

eler

atio

n (m

/s^2

)

Time!(sec)!

M1.5000E+02!M1.0000E+02!M5.0000E+01!0.0000E+00!5.0000E+01!1.0000E+02!1.5000E+02!

0.000! 0.500! 1.000! 1.500! 2.000! 2.500! 3.000! 3.500! 4.000! 4.500! 5.000!

Acc

eler

atio

n (m

/s^2

)

Time!(sec)!


! !50!

8 3-D Printed Model Following the design and simulation running of the 3-D model, it was decided

that a half scale model of the tested concept was to be carried out. The half scale model

was 3-D printed on a Selective Laser Sintering (SLS) printer in the Engineering

workshop in the "A" block of the main building. The material used is called “Vero

White” this material was selected for its aesthetic visual appeal and crisp sharp lines

that can be produced from printing. However this material is not stiff enough to be used

as a substitute for the material specified in the working drawings for the components.

The model produced as a visual and physical aid as demonstrating 3-D printing can

effectively convey a design idea by the rapid production of scaled concepts.

Comparison images are produced between the 3-D model and the scaled printed

prototype. Figure 8.1 displays an isometric image of the modelled suspension system on

the left and the 3-D printed on the right.

!' FIGURE'8.1'(A)'3CD'COMPUTER'MODEL,' ' ' ' '(B)'PRINTED'MODEL

It can be seen that the computer model differs from the 3-D printed version on

the seat platform and base. In order to generate a more cost effective printed model

sections of the top and bottom plate were removed. Removal of these sections material,

thus reduced the cost of the model sufficiently.

Figure 8.2 shows a left view of the model fully extended with the corresponding

printed image on the right.


! 51!

!'''''''FIGURE''8.2'(A)'FULLY'EXTENDED'COMPUTER'MODEL,' ' '(B)'FULLY'EXTENDED'PRINTED'MODEL

Figure 8.3 shows a left view of the model fully compressed with the

corresponding printed image on the right.

!''''FIGURE''8.3'(A)'FULLY'COMPRESSED'COMPUTER'MODEL,'' ' '''(B)'FULLY'COMPRESSED'PRINTED'MODEL

Figure 8.4 presents on the left the modelled seat suspension with a modelled seat

attached to it, the image on the right has the same seat model projected onto the printed

suspension system in order to gain a visualisation on how big the suspension system is

compared to a generic seat size.

!'''''''''FIGURE''8.4'(A)'COMPUTER'MODEL'WITH'SEAT,'' ' ''''''''(B)'PRINTED'MODEL'WITH'SEAT'PROJECTED'ON


! !52!

9 Discussion

9.1 Introduction The final concept designed throughout this project had the aim of meeting the

objectives set out in section 1.2 and 1.3 of this thesis. For the concept to be meet the

criteria the design had to be a fully contained suspension system that operates outside

the natural frequencies of the human body and that passes the ISO standards pertaining

to vibration transmissibility and size constraints. The following sections in this chapter

outline the work done throughout this project, discuss the results obtained from

simulated testing and outline recommendations for future work to be carried out if the

project is to be perused further.

9.2 Overview of design

Research relating to injuries caused by whole body vibration directed this thesis

to try and design a suspension system that reduced the occurrence of such injuries such

as lumbar syndrome as discussed in chapter 2. The concept chosen to model and run

simulations on was picked because it was felt that due to the diagonal motion of the

suspension system during the compression stroke would reduce the occurrence of

bottoming and topping. If the system was to bottom out the resulting shock load felt by

the operator would be spread over the operators buttocks and back, rather than transmit

the shock vertically up through the operators spine compressing the vertebra. figure 9.1

(a) illustrates a suspension system that has vertical travel only and the direction the

resulting shock load due to bottoming. (b) illustrates the prototypes motion and

direction of resulting shock load due to bottoming.

!!!!

!!!!!

Load!absorption!area!

FIGURE'9.1'(A)'VERTICAL'TRAVEL'SUSPENSION'SYSTEM'WITH'REACTION'LOAD'ABSORBTION'AREA,''

' (B)'DIAGIONAL'TRAVEL'WITH'REACTION'LOAD'ABSORBTION'AREA'


! 53!

The elastomeric material design for the vibration-absorbing component of the

suspension system was chosen as the researcher found that there was a gap in the

literature for suspension systems that incorporated elastomer only as a combined spring

and damper system. However due to time limitations manufacturing and testing of

elastomer that have non-linear spring stiffness, to determine an elastomer that performs

as required for the design was not carried out. Research into elastomer testing methods

was conducted; these testing methods are outlined in section 9.4.1 in recommendations

for future work.

9.3 Simulation results Static simulation results were discussed in section 5.4 for each of the suspension

components, it was found that through FE testing that the design of each component

was stiff enough to withstand the forces exerted on the system. However theses

simulations have not been validated yet in section 9.4.2, it is recommended that a full-

scale prototype be manufactured for the proposes of testing to validate results gained for

both static and dynamic simulations. The results obtained in this thesis have as yet not

been verified by real life laboratory experimentation; therefore the results presented are

theoretical predictions of how the suspension system should operate.

!!!!9.3.1 SEAT factor

As stated in section 7.2, there was no method of gathering vibration data

generated from travelling over different terrain, therefore the decision was made to

modify the input forcing function required for the damping simulation. SEAT factor

simulation were carried out with a constant vibration amplitude of 40% the total travel

of the suspension system but with different input frequencies ranging from 1-20Hz.

Table 7.1, illustrates the SEAT factor calculated for each of the frequencies tested. It

can be seen that for a frequency above 13Hz the suspension system passes the criteria

for ISO 7096. However, below this frequency the vibration transmitted through the

system is amplified rather than reduced. Due to the complexity of the input vibration

and the model dynamics it is felt by the designer that in order to fully understand the

response of the system a full scaly prototype must be manufactured and tested.!!

!

9.3.2 Damping performance

! Stated in section 7.3, the suspension system simulation passed the criteria for

damping transmissibility. It was found that at resonance the damping transmissibility


! !54!

was 0.97, which is below the maximum permissible damping transmissibility of 2. The

damping test was carried out when the base of the suspension system was harmonically

excited at the systems natural frequency. Figure 7.4 confirms that the input frequency of

the forcing function was at the resonant frequency of the system, as the response of the

seat plate displayed a beating pattern. Rao explains that when the forcing function is

close to but not exactly equal to the natural frequency of the system beating will occur.

(Rao, 2004) Beating is a kind of vibration that allows the amplitude to build up and

reduce in regular patterns.

!9.4 Recommendations for future work

Having completed the initial design of the suspension system and generated

simulations defining the theoretical performance of the suspension system, the

following recommendations are suggested to further develop this concept. These

suggestions are made with the intention of them to be carried out as part of the post-

graduate, Masters in Mechanical Engineering.

9.4.1 Elastomer selection

A variety of elastomeric materials should be manufactured with similar

performance characteristics outlined in section 5.5; these materials should then be

subjected to dynamic testing on a Q800 DMA system. Compression tests should be

carried out on this machine. The material is to be placed between two plates. The first a

fixed flat surface, mimicking the bottom elastomer holder in the design and the second a

sinusoidal oscillating plate applying a harmonic compression stress to the material. This

test will determine the compression on the elastomer for a given force and spring rate.

The Q800 DMA system can perform these tests under a variety of testing temperatures

to predict the materials resistance to hot and cold temperature that it may experience

when put into production.

9.4.2 Concept manufacture

The next stage in testing is to manufacture a full-scale model to the

specifications outlined in sections 5.3 and 5.4. Dimensions of each component can be

found in Appendix B. Testing can be carried out in accordance with ISO 7096; results

from real world testing can then be compared to the simulated analysis, to verify that

the simulation results were accurate. Once testing of the prototype is completed and


! 55!

proves that the predicted simulation results are accurate, design changes can be made to

the suspensions design to alter its performance. Simulations on the design changes can

then be carried out with the knowledge that the results obtained are representative of

real world results.

9.4.3 Cost effective examination

Once the elastomeric material is chosen and the design of the suspension system is

finalised. Production costs, service life and replacement costs can be calculated and

compared with existing designs already on the market. Design changes, manufacturing

methods can then be revaluated to ensure that the final design can be produced and sold

cheaper than existing suspension systems on the market.


! !56!

10 Conclusions The aim of this thesis was to develop a novel forklift seat suspension prototype

that was capable of passing international standards for vibration transmissibility.

Specifications were generated, which acted as building blocks for all the concepts

developed throughout the designing of the final prototype. a novel suspension system

was developed utilising elastomers as vibration absorbing combined spring and damper

component.

3-D FE analysis was conducted, validating the components stiffness and

ensuring that the system pivots about its hinges without any interference. Dynamic

analysis was carried out in accordance with a modified version of ISO 7096 allowing

the prototype to be tested virtually. Results show that the model passes the criteria for

ISO 7096, however when determining the SEAT factor a representative time history of

an actual representative signal for the input vibration was not used, resulting in the

simulation only passing the criteria when the input frequency is above 13Hz. Not using

a representative input vibration means that the simulation did not produce results

according to the standard, therefore it is recommended in section 9.4.2, that the model

be manufactured and physical simulations be conducted.

A 3-D printed 50% scale model was then printed this model displays the motion

of the suspension system allowing faster communication of the concept amongst peers.

as well as outlining issues with assembling components together that were not made

obvious when viewing a virtual model.

With Simulations carried out the initial design and motion of the suspension

system passed the criteria outlined at the start of this thesis, therefore this concept has

been successful and as a result of this success has the potential to be developed further

as outlined in section 9.4.


! 57!

References !Crawford, R.J., Plastics Engineering, Butterworth-Heinemann, 3rd Ed., 2005.

Donati, P. 2002. Survey of technical preventative measures to reduce whole-body

vibration effects when designing mobile machinery. Journal of sound and vibration, 253

pp. 169-183.

Hill, T. E., Desmoulin, G. T. and Hunter, C. J. 2009. Is vibration truly an injurious

stimulus in the human spine. Journal of Biomedichanics, 42 pp. 2631-2635.

ISO 3411 “earth moving machinery- Physical dimensions of operators and

minimum operator space envelope”

ISO 10326-1 “mechanical vibration – Laboratory method for evaluating vehicle seat

vibration”

ISO 7096 “earth moving machinery – Laboratory evaluation of operator seat vibration”

Hostens, I., Deprez, K. and Ramon, H. 2003. An improved design of air suspension for

seats of mobile agricultural machines. Journal of sound and vibration, 276 pp. 141-156.

Maciejewski, I., Meyer, L. and Krzyzynski, T. 2008. Modelling and multi-criteria

optimisation of passive seat suspension vibro-isolating properties. Journal of sound and

Vibration, 324 pp. 520-538.


! !58!

Schwarze, S., Notbohm, G., Dupuis, H. and Hartung, E. 1998. Does-Response

relationship between whole-body vibration and lumbar disk disease- A field study on

388 drivers of different vehicles. Journal of sound and vibration, 215 pp. 613-628.

Schwarze, S., Notbohm, G., Dupuis, H. and Hartung, E. 1998. Does-Response

relationship between whole-body vibration and lumbar disk disease- A field study on

388 drivers of different vehicles. Journal of sound and vibration, 215 pp. 618.

Silsoe research institute. 2013. Ride vibration: Reduction of shocks arising from over

travel of seat suspensions. [report] Sheffield: HSE books, pp. 73-119.

Rao, S. S. 2004. Mechanical vibrations. Upper Saddle River, N.J.: Pearson Prentice Hall

pp. 661.

Rao, S. S. 2004. Mechanical vibrations. Upper Saddle River, N.J.: Pearson Prentice Hall

pp. 226.


! A!

Appendix !Appendix A: DVD of;

Part files

Assembly files

Simulations, Dynamic simulations are contained in the "Testing assembly" file

Video of dynamic analysis.


! !B!

Appendix B:

Working!drawings!of!assembled!concept,!Exploded!view,!working!drawings!of!each!individual!component.!

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Seat Suspension Thesis

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