IRJET- Development of Semi-Active Suspension System-Variable Orifice based Damper Prototype

International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056
Volume: 05 Issue: 10 | Oct 2018 www.irjet.net p-ISSN: 2395-0072
© 2018, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 980
Development of Semi-Active Suspension System- Variable Orifice based
Damper Prototype
Rupesh Vetal1, Dr. S S Bhavikatti2
1M.Tech, Automotive Technology, College of Engineering Pune
2Professor, Dept. of Mechanical Engineering, College of Engineering Pune
------------------------------------------------------------------------***-------------------------------------------------------------------------
Abstract – The semi-active suspension in automotive
employs the damper unit with varying performance based on
condition of roads, bumps, potholes when vehicle passes over
them. The variable orificedampersusedfortuningof dampers,
are once tuned gives particular kind of dynamicresponse. This
aspect of existing variable orifice damper is examined for
developing the semi active suspension system in this research
work. The mathematical modelling for damping resistance is
developed and is compared for variousorificesizes, thisdatais
then used for developing the proposed CAD model dimensions
of damper. The final objective is to calibrate the developed
prototype to make it work best suitable as mechanically
adjustable semi-active suspension system, which is presented
in another research paper. The simulation work is done with
the software MATLAB, and CAD modelling is done with the
help of software Creo 4.0
Key Words: Mechanically adjustable orifice,
Mathematical modelling, MATLAB, Semi-active
suspension, MR dampers vs mechanically adjustable
dampers
1. INTRODUCTION
Handling and ridecomfortareveryimportantcharacteristics
that influence the qualityofthevehicle.Thesecharacteristics
depend on the suspension system of the vehicle. A semi-
active damping system is basically a dissipative element in
which the dissipation law can be actively modulated. This
system due to its cost effectiveness, light weight and low
energy consumption, is expected to be used in passenger
vehicles in the near future. With sensible control law, the
semi-active systems provide an intermediate performance
between fully active systems and passive systems. An active
suspension requires significant external power to function
and that there is also in considerable penalty in complexity,
reliability, cost and weight. With a view to reducing
complexity and cost while improving ride, handling, and
performance, the concept of a semi-active suspension has
emerged. In this kind of system,theconventional suspension
spring is usually retained, while the damping force in the
shock absorber can be modulated in accordance with
operating conditions. With passive damping these
requirements are conflicting, since a hard damper results in
better stability but reduced ride comfort, whereas a soft
damper results in the converse effect. Semi-active devices
can be broken into two distinct categories: those that use
smart materials, and those that do not. Both are computer
controlled intelligent systems, yet their principles of
operation differ. Smart material devices include electro-
rheological and magneto-rheological (MR) fluids, Shape
Memory Alloys, Piezoceramics, and any other material
whose physical properties change in response to magnetic,
electrical, or thermal stimulus. Semi-active devices that do
not use smart materials instead use more conventional
actuators to modify their output force. Common actuators
include voice coils and solenoid valves, pneumatic or
hydraulic cylinders, as well as stepper motors and servos.
The resistance or dampening characteristics of passive
suspension damper depends mainly on the resistance
offered by the fluid flowing into compression and extension
chamber through orifices or valves.
1.1 Adjustable valve damper
The geometrical featureof theseflowvalves(ororifice)in
the damper, orifice area is responsible for the resistance
offered by the damper. The effect of this orifice area change
on the force output i.e. the resistance of the damper, is
studied in this project.
The variation in the orifice area can be achieved with
opening and closing the valves as in usual hydraulic circuits.
This technique to vary the valve area is discussed in the
thesis in later chapter.
Finally, complete content and organizational editing
before formatting. Please take note of the following items
when proofreading spelling and grammar:
1.2 MR materials
A magnetorheological (MR) material is one for which the
rheological properties, such as yield stress and viscosity,
depend upon the magnetic field. MR liquids are sometimes
described as ‘low voltage’ in contrast to ERliquids,butthisis
misleading. They are not subject to any voltage, only to a
magnetic field, but the field is generated by a current of
order one amp at low voltage in a field coil external to the
liquid.
The MR liquid is formed by suspending numerous small
solid particles, typically, a few micrometers in diameter, in a
low-viscosity mineral or silicone carrier oil. The average
diameter is about 8 mm with a normal range of 3–10 mm.
The solid particles are ferromagnetic, basicallyjustsoftiron.
Fibrous carbon may be added, and also a surfactant to
minimize settling out. The result is a very dense ‘dirty’ grey
to black oil. The shear strength achievable with magnetic
field on is typically 50–100 kPa at fields of 150–250 kA/m.

The magnetic activation is not sensitive to electrical
conductivity, so temperature has less effect than for ER
devices. [3]
Gordaninejad and Breese examined the effects of
temperature on the force characteristics of three separate
MR dampers [9]. Additionally,heatgenerationovertimewas
evaluated experimentally. A theoretical model was
developed to represent the temperature increase in a MR
damper due to a constant sinusoidal inputandcurrenttothe
electromagnet.
Fig -1: A typical MR damper design[3]
The semi-active suspension system developed so far are
using MR (magneto-rheological) dampers. The working of
the MR dampers depends on the ferromagnetic fine powder
suspended or mixed in base damper oil. The Dixon J []
discusses the MR damper and working of the MR damper.
Fig -2: Section view: Typical mono tube MR damper
A significant amount of research has been conducted on
electro-mechanically adjustable dampers, although much of
it has been proprietary. Kitching, et al., developed a solenoid
actuated spool valve hydraulic damper [12]. The non-
linearity of the valve was modeled and evaluated
experimentally through hardware in the loop testing. The
semi-active damper showed the most influence in the low
frequency bounce (heave) mode, reducing RMS body
accelerations by up to 12.3 %, dependingonroadconditions.
The semi-active damper was also shown to reduce peak
absolute body acceleration due to transient bump inputs,
without compromising settling time. The damperdeveloped
in [12] is similar to many other semi-active hydraulic
dampers previously studied. ZF Sachs GmbHutilizea similar
solenoid actuated proportioning valve in their Continuous
Damping Control® semi-active hydraulic damping system.
Usman and Parker developed a low cost electro-hydraulic
proportioning valve suitable for semi-active suspensions.
Their goal was to reduce the unit costs associated with high
precision machining, while providing a valve that is
insensitive to fluid contamination and periods of inactivity
[21]. A floating double disc configurationwasdeveloped and
modeled mathematically. Magnetic and fluid forces versus
displacement were calculated and used to relate differential
coil current to disc position. Sun andParker expandedon the
previous work by integrating the double disc valve
configuration into a semi-active damper [20].
MR dampers utilize magnetorheological fluids, instead of
hydraulic oils, as the working fluid. Theapparentviscosityof
the MR fluid is varied by controlling the current through an
electromagnet. This electromagnet creates magnetic flux
path between the piston and damper body which locally
energizes the fluid flowing through an annular gap between
the outside of the piston and the damper body. A schematic
of the flux path, piston, and damper body layout is shown in
Figure 2, as well as a plot of the pressure drop and flow rate
relationship for different magnetic field strengths.[6]
2. MATHEMATICAL MODELLING OF VISCOUS
DAMPING
Conventional hydraulic dampers work by passing fluid
through an orifice, or series of orifices, in response to the
presence of a displacement input. When the working fluid,
typically petroleum based oil, is passed through the damper
valves, a pressure differential is created on opposing sides of
the piston. This pressuredifferentialactsupontheareaofthe
piston to create a force which opposes damper motion. To
better understand the natureof the various forces at work in
a conventional monotubedamper,afreebodydiagram(FBD)
is shown in Figure 2.
Fig -3: Free body diagram of damper internals
In addition to the damping, drag, and gas forces shown in the
FBD, the pressures on the rebound and compression sides
(Preb and Pcomp, respectively) give rise to the dominant force
which reacts Fdamping. A forcebalance in they-directionyields,
Σ𝐹𝑦 = 0
𝐹𝑑𝑎𝑚𝑝𝑖𝑛𝑔 + 𝑃𝑟 𝐴𝑝 − 𝐴𝑟 − 𝑃𝑐 𝐴𝑝 − 𝐹𝑑𝑟𝑎𝑔 − 𝐹𝑔𝑎𝑠 = 𝑚 ∗ 𝑎 𝑦
where Fdamping is the input force, Preb and Pcomp are the static
pressures on the rebound and compression sides of the
piston, Ap is the area of the piston, Ar is the cross-sectional
area of the rod, Fdrag is the coulomb friction force between
piston guideand damper body, and Fgas is the force duetothe
gas reservoir. Fdrag is a function of the materials used for the
wear band, damper body, damper rod, and main bearing,
their surface finishes,and thefitment between them.Forthis
analysis Fdrag is neglected because it is generally very small
when compared to fluid forces. Fgas, which depends on the

pressure in the gas reservoir and the rod area upon which
this pressure acts, is also neglected because it gives rise to a
spring force that is independent of velocity. Thus, equation
reduces to-
𝐹𝑑𝑎𝑚𝑝𝑖 𝑛 𝑔 = 𝑃𝑐 𝑜𝑚𝑝 𝐴𝑝 − 𝑃𝑟𝑒𝑏(𝐴𝑟 − 𝐴𝑟)
The major resistance given by the damper is due to
pressure head loss in oil when oil flowing about orifice. The
formulae for calculating head loss are as follows. Thus by
substituting the selected dimensionsofdamperassemblythe
pressureheadlosseswerecalculatedwithfollowingformulae
Fig 4: Fluid flow sudden contraction
Fig 5: Fluid flow sudden enlargement
2.1 Piston- Disc Valve Geometry
To allow the required area variation, the assembly of the
piston in hydraulic damper can be modified as shown in Fig
5. It shows the section view of the damper; a disc valve is
incorporated within a two-piece piston. The orifices on the
disc can be rotated out of phase with the orifices onthemain
piston, allowing the effective flow area of the valve to be
controlled via external actuation.
Fig 6: The damper section view
The valve design is shown in further detail inFigures3.3and
3.4, each depicting section views of the piston assembly, the
first of which includes the control rod that rotates the
internal disc. The rebound and compression decoupling is
important because desired rebound and compression
damping often differ considerably. This decoupling maybe
achieved by using the thin shims.
Fig 7: Piston Assembly
Fig 8: Piston Assembly without control rod
Fig 9: Piston Assembly front view
The formula derived for above calculation is as foll0ws-
𝐴𝑟𝑒𝑎 𝑂𝑝𝑒𝑛 𝑐𝑜𝑚𝑝𝑟𝑒𝑠𝑠𝑖𝑜𝑛 𝑜𝑟 𝑒𝑥𝑡𝑒𝑛𝑠𝑖𝑜𝑛 = 3 ∗ 2 ∗ 𝑟2
∗ 𝑐𝑜𝑠−1
𝑑
2 ∗ 𝑟
−
𝑑 ∗ 4𝑟2 − 𝑑2
2
𝑑 = 2 ∗ 𝑅 ∗ sin⁡
𝜃
2
Where, d= Distance between the hole on piston and
respective hole on disk,
R=Radius at which the center of hole is situated on piston,
θ= The angle of disk turned with respect to damper piston
r= Radius of hole on piston (or orifice)

3. SIMULATION
The input data for the simulation is as follows
Piston diameter= 40 mm,
Hole diameter on piston= 8 mm and 10 mm
Position of hole diameter from piston center=28 mm,
Angle= 0o to 24o
Table -1: Results
Relative
Position
(Angle)
Hole Radius,
r= 3 mm
Hole Radius,
r= 4 mm
Hole Radius,
r= 8 mm
0 84.8230016469244 150.796447372310 235.619449019234
1 76.0363889860067 139.075281816393 220.964706726709
2 67.3090830285452 127.398890302516 206.346149999856
3 58.7015461609999 115.812512305351 191.800191321587
4 50.2766764436387 104.362345792499 177.363705643527
5 42.1013687206500 93.0960993851025 163.074283955891
6 34.2485950996590 82.0636437788007 148.970515879334
7 26.8004206616365 71.3178185359942 135.092315040855
8 19.8527714905555 60.9154785510896 121.481305386694
9 13.5237784318741 50.9189151162581 108.181293471286
10 7.97050624444608 41.3978814665410 95.2388626600194
11 3.43054101494638 32.4326425714503 82.7041428371896
12 0.386690224332264 24.1188829605926 70.6318387786262
13 0 16.5763207789880 59.0826520479661
14 0 9.96580819250863 48.1253266481196
15 0 4.53067030422575 37.8397367753296
16 0 0.746744015768744 28.3218390834275
17 0 0 19.6922830276123
18 0 0 12.1132029996445
19 0 0 5.82747150048806
20 0 0 1.28943888319652
21 0 0 0
22 0 0 0
23 0 0 0
24 0 0 0
Fig 10: Orifice area variation wrt disc valve position
Fig 11: MATLAB Simulink model for finding damping force
wrt displacement and velocity of piston
The results fordampingforceversusvelocityofpistonare
given below-
The input data for the simulation is-
Density of damper oil = 860 kg/m3,
Height of bump = 100 mm
Length of bump = 800 mm
Static Pressure = 1 atm=1.0132 bar
Disc position= 9o,
Input Frequency = 5 Hertz
Fig 12: Damping force Vs. velocity of piston
The results for damping force versus displacement of
piston are given below-
The input data for the simulation is-
Density of damper oil = 860 kg/m3,
Height of bump = 100 mm
Length of bump = 800 mm
Static Pressure = 1 atm=1.0132 bar
Disc position= 9o,
Input Frequency = 5 Hertz

Fig 13: Damping force Vs. displacement of piston
4. CONCLUSIONS
4.1 Area opening variation with disc position for
various orifice diameter
The change in disc position by rotating disc (changed
relative position) changes the orificevalvearea non-linearly.
The change in the area thus is second degree curve, or the
area is function of the square of the position angle
𝑉𝑎𝑙𝑣𝑒 𝐴𝑟𝑒𝑎, 𝐴 = 𝑓(𝜃)
4.2 Damping force Vs. velocity of the piston
As an example, for one of the case for disc positionθ=10o
In Compression cycle- for maximum velocity of piston
Hole radius r= 3 mm, Maximum damping force= 3963 N
In Extension cycle- for maximum velocity of piston
4.3 Damping force Vs. velocity of the piston
The results obtained for various disc positions with circular
hole were simulated in this study simulation.
The Orifice area geometry along with disk hole geometrical
features can be modified to get more linear valve area
(orifice) variation.
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