WEEKS 8-9 Dynamics of Machinerykisi.deu.edu.tr/hasan.ozturk/makina dinamigi/Makina...WEEKS 8-9...
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WEEKS 8-9 Dynamics of Machinery
• References Theory of Machines and Mechanisms, J.J.Uicker,
G.R.Pennock ve J.E. Shigley, 2011 Mechanical Vibrations, Singiresu S. Rao, 2010 Mechanical Vibrations: Theory and Applications, S. Graham Kelly, 2012
Prof.Dr.Hasan ÖZTÜRK 1
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Prof.Dr.Hasan ÖZTÜRK
-Vibrations are oscillations of a mechanical or structural system about an equilibrium position.
Vibration Analysis
-Any motion that exactly repeats itself after a certain interval of time is a periodic motion and is called a vibration.
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we encounter a variety of vibrations in our daily life. For the most part, vibrations have been considered unnecessary.
washing machine 3
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Condition monitoring
Linear vibration sieve is also called linear sieve, which is one of the most widely used vibrating creening equipment. Linear vibration sieve can easily finish all kinds of material removing impurity, grading, screening.
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Free Vibration of Single Degree of Freedom Systems
Figure A. The space needle (structure)
The structure shown in Figure A can be considered a cantilever beam that is fixed at the ground. For the study of transverse vibration, the top mass can be considered a point mass and the supporting structure (beam) can be approximated as a spring to obtain the single-degree-of-freedom model shown in Figure B.
Figure B. Modeling of tall structure as spring-mass system
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The figure shows an idealized vibrating system having a mass m guided to move only in the x direction. The mass is connected to a fixed frame through the spring k and the dashpot c. The assumptions used are as follows:
1. The spring and the dashpot are massless. 2. The mass is absolutely rigid. 3. All damping is concentrated in the dashpot.
Consider next the idealized torsional vibrating system of the below figure. Here a disk having a mass moment of inertia I is mounted upon the end of a weightless shaft having a torsional spring constant k, defined by
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where T is the torque necessary to produce an angular deflection θ of the shaft. In a similar manner, the torsional viscous damping coefficient is defined by
Next, designating an external torque forcing function by T = f (t), we find that the differential equation for the torsional system is
EXAMPLE:
The below figure illustrates a vibrating system in which a time-dependent displacement y= y(t) excites a spring-mass system through a viscous dashpot. Write the differential equation of this system.
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VERTICAL MODEL: the external forces zero
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FREE VIBRATION WITHOUT VISCOUS DAMPING Free vibrations are oscillations about a system’s equilibrium position that occur in the absence of an external excitation.
Consider the configuration of the spring-mass system shown in the Figure
D Alembert s Principle.
The constants A and B can be determined from the initial conditions of the system.
the system s natural frequency of vibration
9
00
n
vA , B x= =ω
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The ordinate of the graph of the above Figure is the displacement x, and the abscissa can be considered as the time axis or as the angular displacement ωnt of the phasors for a given time after the motion has commenced. The phasors x0 and ν0 /ωn are shown in their initial positions, and as time passes, these rotate counterclockwise with an angular velocity of on and generate the displacement curves shown. The figure illustrates that the phasor ν0 /ωn starts from a maximum positive displacement and the phasor x0 starts from a zero displacement. These, therefore, are very special, and the most general form is that given by , in which motion begins at some intermediate point.
the system s natural frequency of vibration
the period of a free vibration is
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The above equation is harmonic function of time. The motion is symmetric about the equilibrium position of the mass m.
the solution can be written as
Equation can also be expressed as
Harmonic Motion
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where X0 and φ are the constants of integration whose values depend upon the initial conditions.
1 0
0
nxtanv
− ωψ =
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velocity
acceleration
displacement
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There is a difference of 90 degrees between the equations
Phase relationship of displacement, velocity, and acceleration
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Example:
(a) System of the example . A mass is dropped onto a fixed-free beam.
(b) The system is modeled as a mass hanging from a spring of equivalent stiffness. Since x is measured from the equilibrium position of the system, the initial displacement is the negative of the static deflection of the beam.
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φ0
φ0
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Example:
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Simple Pendulum (Approximate Solution)
19 - 18
• Results obtained for the spring-mass system can be applied whenever the resultant force on a particle is proportional to the displacement and directed towards the equilibrium position.
for small angles,
( )
gl
tlg
nn
nm
πωπτ
φωθθ
θθ
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sin
0
==
+=
=+
:tt maF =∑
• Consider tangential components of acceleration and force for a simple pendulum,
0sin
sin
=+
=−
θθ
θθ
lg
mlW
Example:
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The equivalent spring constant of a parallel spring arrangement (common displacement) is the sum of the individual constants.
The equivalent spring constant of a series spring arrangement (common force) is the inverse of the sum of the reciprocals of the individual constants.
Combination of springs
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a compacting machine 20
STEP INPUT FORCING
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Let us assume that this force is constant and acting in the positive x direction. we consider the damping to be zero.
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PHASE-PLANE REPRESENTATION
We have already observed that a free undamped vibrating system has an equation of motion, which can be expressed in the form
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PHASE-PLANE ANALYSIS
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TRANSlENT DISTURBANCES
Any action that destroys the static equilibrium of a vibrating system may be called a disturbance to that system. A transient disturbance is any action that endures for only a relatively short period of time.
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Construction of the phase-plane and displacement diagrams for a four-step forcing function.
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EXAMPLE
(1) Prof.Dr.Hasan ÖZTÜRK
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Phase-Plane Graphical Method.
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Free Vibration with Viscous Damping
we assume a solution in the form
where A and s are undetermined constants. The first and second time derivatives of are
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Inserting this function into Equation leads to the characteristic equation
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Thus the general solution
Critical Damping Constant and the Damping Ratio.
Thus the general solution
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1 2C A, C B= =
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and hence the solution becomes
The constants of integration are determined by applying the initial conditions
is called the frequency of damped vibration
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0 0x v=
And the solution can be written
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Thus, the general solution is
Free vibrations with ζ= 1 are called critically damped because the damping force is just sufficient to dissipate the energy within one cycle of motion. The system never executes a full cycle; it approaches equilibrium with exponentially decaying displacement. A system with critical damping returns to equilibrium the fastest without oscillation. A system that is overdamped has a larger damping coefficient and offers more resistance to the motion.
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Thus, the general solution is
The response of a system that is overdamped is similar to a critically damped system. An overdamped system has more resistance to the motion than critically damped systems. Therefore, it takes longer to reach a maximum than a critically damped system, but the maximum is smaller. An overdamped system also takes longer than a critically damped system to return to equilibrium.
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Prof.Dr.Hasan ÖZTÜRK Logarithmic Decrement:
The logarithmic decrement represents the rate at which the amplitude of a free-damped vibration decreases. It is defined as the natural logarithm of the ratio of any two successive amplitudes. Let t1 and t2 denote the times corresponding to two consecutive amplitudes (displacements), measured one cycle apart for an underdamped system
we can form the ratio
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0
0
0
0
0
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The logarithmic decrement δ can be obtained as:
If we take any response curve, such as that of the below figure, and measure the amplitude of the nth and also of the (n+N)th cycle, the logarithmic decrement δ is defined as the natural logarithm of the ratio of these two amplitudes and is
Example:
12
3
2
1 22 1
N
xlnx
=
πζδ = =
− ζ
2
1 21
n
n N
xlnN x +
πζδ = =
− ζ
N: is the number of cycles of motion between the amplitude measurements.
Example: 32
3 2
2
1 22 1( )
N
xlnx +
=
πζδ = = − ζ
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Measurements of many damping ratios indicate that a value of under 20% can be expected for most machine systems, with a value of 10% or less being the most probable. For this range of values the radical in the below equation can be taken as approximately unity, giving
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EXAMPLE
Prof.Dr.Hasan ÖZTÜRK 43
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Prof.Dr.Hasan ÖZTÜRK
EXAMPLE
2
1 2ln1
nn d
n N
x TN x +
πξδ = = ξω =
− ξ
2 22 , 1 , , 1 2 ,2d n R n n
c kTmkm
π= = ω = ω −ξ ξ = ω = ω − ξ ω =ω
,n
r ω=λ =
ω2
2 2 2
1 (2 )
(1 ) (2 )
rXY r r
+ ξ=
− + ξ
46
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Prof.Dr.Hasan ÖZTÜRK 47
RESPONSE TO PERIODIC FORCING
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Prof.Dr.Hasan ÖZTÜRK 48
The first term on the right-hand side of the above equation is called the starting transient. Note that this is a vibration at the natural frequency ωn, not at the forcing frequency ω. The usual physical system will contain a certain amount of friction, which, as we shall see in the sections to follow, will cause this term to die out after a certain period of time. The second and third terms on the right represent the steady-stale solution and these contain another component of the vibration at the forcing frequency ω.
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Prof.Dr.Hasan ÖZTÜRK 49
Computer solution of the above equation for ωn = 3 ω; the amplitude scales for x and y are equal.
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Prof.Dr.Hasan ÖZTÜRK 50
Examination of the Equation indicates that, for ω/ωn =0, the solution becomes
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Harmonically Excited Vibration
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h p
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Prof.Dr.Hasan ÖZTÜRK 55
Because the forcing is harmonic, the particular part is obtained by assuming a solution in the form
Steady-State Solution
Steady-State Solution
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Prof.Dr.Hasan ÖZTÜRK 56
lagging the direction of the positive cosine by a phase angle of
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only in the steady-state term
and find the successive derivatives to be
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These equations can be simplified by introducing the expressions
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Relative displacement of a damped forced system as a function of the damping and frequency ratios.
Relationship of the phase angle to the damping and ffequency ratios
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FORCING CAUSED BY UNBALANCE
Doç.Dr.Hasan ÖZTÜRK
If the angular position of the masses is measured from a horizontal position, the total vertical component of the excitation is always given by
62 Prof.Dr.Hasan ÖZTÜRK
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A plot of magnification factor versus frequency ratio
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Relative Motion
Automobiles have the input vibratory motion from the ground and hence it comes under Support motion
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Response of a Damped System Under the Harmonic Motion of the Base
Sometimes the base or support of a spring-mass-damper system undergoes harmonic motion, as shown in Fig. 1(a). Let y(t) denote the displacement of the base and x(t) the displacement of the mass from its static equilibrium position at time t. Then the net elongation of the spring is x-y and the relative velocity between the two ends of the damper is From the free-body diagram shown in Fig. 1(b), we obtain the equation of motion:
Fig.1
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Using trigonometric identities, the above equation can be rewritten in a more convenient form as
Prof.Dr.Hasan ÖZTÜRK 67
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Z
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Force Transmitted and ISOLATION
Prof.Dr.Hasan ÖZTÜRK
The steady-state solution
The transmissibility T is a nondimensional ratio that defines the percentage of the exciting force transmitted to the frame.
70
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This is a plot of force transmissibility versus frequency ratio for a system in which a steady-state periodic forcing function is applied directly to the mass. The transmissibility is the percentage of the exciting force that is transmitted to the frame.
rotating unbalance
Prof.Dr.Hasan ÖZTÜRK
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Prof.Dr.Hasan ÖZTÜRK 72
A plot of acceleration transmissibility versus frequency ratio. In a system in which the exciting force is produced by a rotating unbalanced mass, this plot gives the percentage of this force transmitted to the frame of the machine. mu
mu
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Prof.Dr.Hasan ÖZTÜRK 73
We shall choose the complex-operator method for the solution of the system of the figure. We begin by defining the forcing function as
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EXAMPLE
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EXAMPLE
78
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EXAMPLE
79
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EXAMPLE
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EXAMPLE
81
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TORSIONAL SYSTEMS We wish to study the possibility of free vibration of the system when it rotates at constant angular velocity. To investigate the motion of each mass, it is necessary to picture a reference system fixed to the shaft and rotating with the shaft at the same angular velocity. Then we can measure the angular displacement of either mass by finding the instantaneous angular location of a mark on the mass relative to one of the rotating axes. Thus, we define θ1 and θ2 as the angular displacements of mass 1 and mass 2, respectively, with respect to the rotating axes.
I1 I2
kt
I1 I2
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Therefore, the masses rotate together without any relative displacement and there is no vibration.