Modal Analysis Damped Systems

Transcription

1 Modal Analysis of MDOF Forced Damped Systems

2 Lecture : MODAL ANALYSIS OF MDOF FORCED DAMPED SYSTEMS TABLE OF CONTENTS Page. Introduction What is Mechanical Damping? Internal vs. External Damping Distributed vs. Localized Damping Modeling Structural Damping Matrix Equations ofmotion Equations of Motion Using Undamped Modes Three Ways Out Direct Diagonalization Damping Factor Guessing Diagonalization by Modal Projection Diagonalization by Proportional Damping Remainder oflecture Damped matrix EOM Diagonalization By Energy Balance, aka RQ Disgonalization..... Comparison ofexact Versus Diagonalized Responses......

3 . WHAT IS MECHANICAL DAMPING?.. Introduction The present lecture introduces damping within the context of dynamic modal analysis. After a brief overview of a very complex but fuzzy subject, the lecture focuses on a very specific damping model. Namely linearly viscous and light. Linearly viscous damping is proportional to the velocity. Light damping means a damping factor (also called damping ratio and damping coefficient) that is small compared to unity. In the terminology of Lecture 7, a lightly damped mechanical system is said to be underdamped. Accounting for damping effects brings good and bad news. All real dynamic systems experience damping because energy dissipation mechanisms are like death and taxes:* inevitable. Hence inclusion makes the dynamic model more physically realistic. The bad news is that it can seriously complicate the analysis process. Here is where the assumption of light viscous damping helps: it allows reuse of major parts of the modal analysis techniques introduced in the previous 3 lectures... What is Mechanical Damping? Damping is the (generally irreversible) conversion of mechanical energy into heat as a result of motion. For example, as we scratch a match against a rough surface, its motion generates heat and ignites the sulphur content. When shivering under cold, we rub palms against each other to warm up. Those are two classical examples of the thermodynamic effect of friction. In structural systems, damping is more complex, appearing in several forms. These may be broadly categorized into internal versus external, and distributed versus localized.... Internal vs. External Damping Internal damping is due to the structural material itself. Sources are varied: microstructural defects, crystal grain relative motions, eddy currents in ferromagnetic media, dislocations in metals, and molecular chain movements in polymers. The key macroscopic effect is the production of a hysteresis loop in stress-strain plots. The hysteresis loop area represents the energy dissipated per unit volume of material and per stress cycle. This kind of damping is intimately related to cyclic motions such as vibration. External damping come from boundary effects. An important form is structural damping, which is produced by rubbing friction: stick-and-slip contact or impact. That may happen between structural components such as bolted or riveted joints, or between a structural surface and non-structural solid media such as soil in building foundations. This form is often modeled by Coulomb damping, which describes the energy dissipation of rubbing dry-friction. Another form of external damping is fluid damping. When a structure is immersed in a fluid, such as air or water, and there is relative motion between the two, a drag force appears. This causes energy dissipation in the fluid through mechanisms such as viscosity, convection or turbulence. This effect is collectively known as fluid damping. A well known application is a vehicle shock absorber: a fluid (liquid or air) is forced through small openings by a moving piston. Parachutes operate on the same basic principle. * One exception: tax-exempt vampires. 3

4 Lecture : MODAL ANALYSIS OF MDOF FORCED DAMPED SYSTEMS... Distributed vs. Localized Damping All damping ultimately comes from frictional effects, which may however take place at different scales. If the effects are distributed over volumes or surfaces at macro scales, we speak of distributed damping. But occasionally the engineer uses damping devices designed to produce beneficial effects. For example: shock absorbers, airbags, parachutes, motion mitigators for buildings or bridges in seismic or hurricane zones, and active piezoelectric dampers for space structures. Those devices can be often idealized as lumped objects, modeled as point forces or moments, and said to produce localized damping. The distinction largely appears at the macro modeling level, since all motion-damper devices ultimately work as a result of some kind of internal energy conversion at the molecular level. Localized dampers may be classified into passive or active. The latter respond to motion feedback. But active devices would take us into control systems, which are beyond the scope of this course..3. Modeling Structural Damping The foregoing summary should make clear that damping is a ubiquitous but complicated business. In structures containing joints, for example, Coulomb (dry friction) damping often dominates; this model is partly nonlinear because the damping force depends on the sign of the velocity. Fluid damping tends to be highly nonlinear if the interacting flow is turbulent, since if so the drag is roughly quadratic in the relative velocity between solid and fluid. Another modeling complication is that friction may depend on fabrication or construction details that are not easy to predict; e.g., bolted versus welded connections. Balancing those complications is the fact that damping in most structures, especially metallic ones, is light in the sense that the damping factor ξ introduced in Lecture 7 is much smaller than one. In addition, the presence of damping is usually beneficial to safety in the sense that resonance effects are mitigated. If so, underestinating or ignoring damping is conservative in a safety sense. This gives structural engineers leeway to simplify a dynamic analysis, by proceeding as follows. A simple damping model, such as linear viscous damping, can be often assumed. Mode superposition is applicable because the EOM is linear. Moreover the frequencies and mode shapes for the undamped system can be reused if additional assumptions studied later, such as Rayleigh damping or modal damping, are made. It should be stressed that such simplifications are not recommended if precise modeling of damping effects is important to safety and performance. This occurs in the following scenarios: Damping is crucial to function or operation. Think, for instance, of a shock absorber, a parachute, or a piezoelectric damper. Fortunately such devices can be often modeled using empirical data (e.g., that provided by the manufacturer). Damping may destabilize the system by feeding energy instead of removing it. This can happen in active control systems and aeroelasticity. The last two scenarios are beyond the scope of this course. In this lecture we will focus on linear viscous damping. Moreover, damping levels will usually assumed to be light in the sense that the damping factor ξ<<; typically less than 0%. 4

5 (a).4 MATRIX EQUATIONS OF MOTION (b) Static equilibrium position k c F = k u s. F = c u d u = u (t) u = u (t) Static equilibrium position k Mass m c Mass m s s F = k (u u ) F = m u.. I d F x F d.. F = c (u u ) F = m u.. I Figure.. Two-DOF, forced, damped spring-mass example system: (a) configuration, (b) DFBD..4. Matrix Equations of Motion Consider again the two-dof mass-spring-dashpot example system of Lecture 9. This is reproduced in Figure. for convenience. The physical-coordinate EOM in detailed matrix notation are [ ] [ ][ ] [ ][ ] [ ] m 0 ][ü c + c + c u k + k + k u p =. (.) 0 m ü c c u k k u p In compact notation, M ü + C u + Ku= p, (.) in which M, C and K are the mass, damping and stiffness matrix, respectively, whereas p, u, u and ü are the force, displacement, velocity and acceleration vectors, respectively. The latter four are function of time: u = u(t), etc, but the time argument will be often omitted for brevity. In the sequel it will be assumed that M, C and K are symmetric. Furthermore M is taken as positive definite (PD) whereas K is nonnegative definite (NND)..4.. Equations of Motion Using Undamped Modes This technique attempts to reuse the modal analysis methods introduced in Lectures 9. Suppose that damping is removed whence C = 0. Get the natural frequencies and mode shapes of the unforced, undamped system governed by M ü + Ku = 0, by solving the eigenproblem KU i = 5

6 Lecture : MODAL ANALYSIS OF MDOF FORCED DAMPED SYSTEMS ωi MU i. Normalize the vibration mode shapes U i φ i so that they are orthonormal with respect to the mass matrix: φi T M φ i = δ ij, (.3) in which δ ij denotes the Kronecker delta: δ ij = ifi = j, else zero. Let Φ be the modal matrix constructed with the orthonormalized mode shapes φ i, and denote by η the array of modal amplitudes η i, also called generalized coordinates. As before, assume mode superposition, so that physical DOF are linked to modal amplitudes via u = Φ η. (.4) Following the same scheme as in the previous two Lectures, the transformed EOM in generalized coordinates are Φ T M Φ ü + Φ T C Φ u + Φ T K Φ u = Φ T p(t). (.5) Define the modal (also called generalized) mass, damping, stiffness and forces as M g = Φ T M Φ, C g = Φ T C Φ, K g = Φ T K Φ, f = Φ T p(t). (.6) Of these, the modal mass matrix M g and the modal stiffness matrix K g were introduced in Lecture 0. If Φ is built by stacking mass-orthonormalized vibration modes as columns, it was shown there that M g = I, K g = diag[ωi ]. (.7) That is, M g reduces to the identity matrix whereas K g becomes a diagonal matrix with squared frequencies stacked along its diagonal. The modal forces f(t), also called mgeneralized forces, were introduced in Lecture. The new term in (.6) is the modal damping matrix C g = Φ T C Φ, also called the generalized damping matrix in the literature. Substituting (.6) and (.7) into (.5) we arrive at the modal EOM for the damped system: η(t) + C g η(t) + diag(ωi ) η(t) = f(t). (.8) Here we ran into a major difficulty: matrix C g generally will not be diagonal. If that happens, the modal EOM (.8) will not decouple. We seem to have taken a promising path, but hit a dead end..4.. Three Ways Out There are three ways out of the dead end: Diagonalization. Keep working with (.8), but make C g diagonal through some artifice. Complex Eigensystem. Set up and solve a different (augmented) eigenproblem that diagonalizes two matrices that comprise M, C and K as submatrices. The name comes from the fact that it generally leads to frequencies and mode shapes that are complex. Direct Time Integration or DTI. Integrate directly (.) numerically in time without passing to modal coordinates. In the textbook of Craig-Kurdila the first two approaches listed above are called mode superposition through real modes of the undamped system, and mode superposition through complex modes of the damped system, respectively. See their Section 0..0 for details. Those mouthfuls are abbreviated above to more easily remembered labels. 6

7 .5 DIRECT DIAGONALIZATION Each approach has strengths and weaknesses. (Obviously, else we would mention just one.) Diagonalization allows straightforward reuse of undamped frequencies and mode shapes, which are fairly easy to obtain with standard eigensolution software. The uncoupled modal equations often have physical interpretation and can be validated with experiments. Only real arithmetic is necessary. The down side is that we dont solve the original damped EOM (.) directly, so some form of approximation is likely. This is counteracted by the fact that structural damping is often difficult to quantify since it can come from many sources, as discussed earlier. Thus the approximation may be tolerable in view of modeling uncertainties. This is particularly true if damping is light. There are certain problems, however, in which diagonalization cannot adequately represent damping effects within engineering accuracy. Three of them are: () Structures with localized damper devices (e.g., shock absorbers, piezoelectric dampers) () Problem in which structure-media interaction is important. Examples: building and bridge foundations, tunnels, aeroelasticity, marine structures, surface ships (3) Active control systems In those situations one of the two remaining approaches must be taken. The complex eigensystem method is mathematically irreproachable and can solve (.) without additional approximations. No assumptions as to light versus heavy damping are needed. But it involves a substantial amount of preparatory work since the EOM must be transformed to the socalled state space form. For a large number of DOF, solving complex eigensystems is unwieldy. Physical interpretation of complex frequencies and modes is less immediate and may require substantial expertise in math as well as engineering experience. But in the scenarios listed above it should be preferred to diagonalization as long as the EOM are linear. This precludes its use in certain damping models. Direct time integration (DTI) has the advantage of being completely general. In fact it can handle not only the linear EOM (.), but also nonlinear systems, as well as non-viscous damping mechanisms. No transformations to mode coordinates is required, and no complex arithmetic emerges. The main disadvantage is that it requires substantial expertise in computational handling of ODE, which is a hairy topic onto itself. Since DTI can only handle fully numerically-specified models, the approach is not particularly useful during preliminary design stages when free parameters float around. Because the last two approaches lie outside the scope of an introductory course (they are usually taught at the graduate level), our choice is easy: diagonalization it is. This choice can be implemented in several ways, three of which are discussed in the following sections:. Direct diagonalizaton: assume uncoupled modal equations and a damping ratio for each mode. Matrix C need not be explicitly constructed. See.5.. Proportional damping: pick a form for C that diagonalizes on passing to modal equations. See.7. Diagonalize by modal projection: form C and project it on the modal coordinates, retianing only the diagonal. See.6 This form is studied in more advanced courses. 7

8 Lecture : MODAL ANALYSIS OF MDOF FORCED DAMPED SYSTEMS.5. Direct Diagonalization In this approach the generalized damping matrix C g = Φ T C Φ is assumed to be diagonal from the start by assuming modal damping factors C g = diag[ξ i ω i M i ] = i =,,...n. (.9) in which n is the number of DOF, M i = φi T M φ i is the i th modal (generalized) mass, and ξ i is the assumed the damping factor for the i th natural mode. A compact matrix form of (.9) used in.5. C g = M g, = UMU (.0) in which = diag[ξ i ω i ]. As a result of this assumption, the modal EOM (.8) decouple, reducing the EOM to n canonical second-order scalar equations in the modal amplitudes η i + ξ i ω i η i + ω i η = f i (t), i =,,...n. (.) These equations can be solved using the methods described in Lecture. The solutions can be superposed via the mode decomposition assumption u(t) = Φ η(t) to get the physical response. The method is straightforward. Two technical difficulties remain. First, how are modal damping factors picked? Second, what is the error incurred by the decoupling assumption?.5.. Damping Factor Guessing One time honored approach is educated guessing. This is of course the only possibility if (a) little is known about the damping level and sources and (b) there are no experimental results; for example the structure only exists on paper. The structural engineer then makes recourse to experience with similar systems. With an air of authority she says: No problema. Let s assign % to modes through 5, % to modes 6 through 0, and 4% to all higher ones. Done. There is some method in the madness. First, damping factors of well constructed structures are typically small compared to unity: to 5% is typical. Second, damping generally increases with frequency, the reason being that more hysteresis cycles take place within a fixed time interval. Third, the effect of light damping is not significant on the response unless the force (or base motion) happens to excite a resonance; so the difference between, say, % and % may be well within modeling uncertainty. Note that this procedure directly delivers a diagonal C g rather than C. If it is desirable to obtain C a posteriori, one may use the following relations,the second of which assumes M g = I: C = Φ T C g Φ = M Φ C g Φ T M = n i= ξ i φ i (Mφ i )(Mφ i ) T. (.) Taking account of mode orthogonality gives the useful relation φ T i Cφ i = ξ i ω i. (.3) 8

9 .9 DAMPED MATRIX EOM.6. Diagonalization by Modal Projection This method assumed that the C matrix of a n-dof system has already been constructed from the model data, as in the EOM example (?). Compute the vibration modes of the undamped system and orthonormalize them with respect to the mass matrix. Let those modes be denoted as usual by φ i, i =,...n. Compute the projection of C on each mode via c i = φt i C φ i φi T, φ i i =,...n. (.4) The diagonalized damping matrix is taken to be C g = diag[c i ] = diag[ξ i ω i M i ]. (.5) The last expression implicitly defines the damping ratio ξ i assigned to the i th mode to be ξ i = c i /( ω i M i ). This construction effectively ignores modal cross coupling. In numerical analysis, the expression (.4) is called the Rayleigh Quotient (RQ) of matrix C with respect to the φ i vector. For this reason, this technique may be also called RQ-diagonalization..7. Diagonalization by Proportional Damping This diagonalization procedure is widely used for civil structures, especially for seismic response calculations. The viscous damping matrix C is directly defined as a linear combination of the mass and stiffness matrix: C = a 0 M + a K. (.6) Here a 0 and a are selected constants of appropriate physical dimensions. The method is called proportional damping in the literature. Applying the modal matrix congruential transformation to this C results in C g = Φ T C Φ = diag[c r ] = diag[a 0 + a ω i ] = diag[ξ i ω i ], (.7) whence the effective modal damping ratio is ξ i = ( ) a0 + a ω i. (.8) ω i Choosing the damping ratio for two modes of different frequencies and solving (.8) for a 0 and a yields C from (.6). In practice the stiffness proportional term is more physically relevant (because, as previously noted, damping usually increases with frequency). Thus structural engineers tend to adjust the choice of those frequencies so that ξ i is roughly minimized for the lowest frequency mode..8. Remainder of Lecture The reminder of this lecture just presents the spring-mass-dashpot example sytem of previous lectures. For various damping levels, this is solved exactly (as a coupled system, by direct numerical integration) and also modally, with a particular damping-matrix-diagonalization method. The results are compared for various damping levels. 9

10 Lecture : MODAL ANALYSIS OF MDOF FORCED DAMPED SYSTEMS (a) (b) Static equilibrium position k c F = k u s. F = c u d u = u (t) u = u (t) Static equilibrium position k Mass m c Mass m s s F = k (u u ) F = m u.. I d F x F d.. F = c (u u ) F = m u.. I Figure.. Two-DOF, forced, damped spring-mass example system: (a) configuration, (b) DFBD..9. Damped matrix EOM Consider again the two-dof mass-spring-dashpot example system of Lecture 9. This is reproduced in Figure. for convenience. The physical-coordinate EOM in detailed matrix notation are [ ] [ ][ ] [ ][ ] [ ] m 0 ][ü c + c + c u k + k + k u p =. (.9) 0 m ü c c u k k u p Suppose m =, m =, k = 6, k = 3, c = c = c, p = 0, and p = F cos t, in which c (a dashpot coefficient), (excitation frequency of harmonic force), and F (harmonic force amplitude) are for the moment kept as free parameters. Then [ ] [ ][ ] [ ][ ] [ ] 0 ][ü c c u 9 3 u =. (.0) 0 ü c c u 3 3 u F cos t The data matrices and vectors are [ ] [ ] [ ] [ ] 0 c c M =, C =, K =, p =. (.) 0 c c 3 3 F cos t The undamped circular natural frequencies and associated modal matrix (orthonormalized with respect to the mass matrix) were obtained in Lecture 0 and are reproduced here for convenience: ω = 3 3 [ ], ω = 6, Φ = [ φ φ ] = =. (.)

11 . COMPARISON OF EXACT VERSUS DIAGONALIZED RESPONSES With the foregoing choice of Φ, the generalized matrices and forces are [ ] c M g = Φ T 0 3 c M Φ = = I, C 0 g = Φ T C Φ = 3 c 3, 5c [ ] [ ] (.3) 6 K g = Φ T 3/ 0 K Φ = = diag[3/, 6], f(t) = Φ T p(t) = F 0 6 cos t The modal EOM in terms of the amplitudes η (t) and η (t) collected in vector η, are η(t) + C g η(t) + diag[3/, 6] η(t) = f(t). (.4) Since C g is not diagonal if c > 0, the modal EOM (.4) are coupled through that matrix..0. Diagonalization By Energy Balance, aka RQ Disgonalization We will investigate the following Rayleigh quotient (RQ) diagonalization technique [ Cg RQ = diag[c RQ, C RQ C RQ ] 0 ] = 0 C RQ, C RQ = φt Cφ φ T φ = c 5, C RQ = φt Cφ φ T φ = 5c. (.5) This relation is obtained by equating the energy dissipated by damping over one cycle. Replacing C g with C RQ g gives the RQ modal EOM [ c η(t) + diag 5, 5c ] η(t) + diag[3/, 6] η(t) = f(t). (.6) These now decouple to η (t) + c 5 η (t) + 3 η (t) = F cos t, η (t) + 5c 6 η (t) + 6η (t) = F cos t. 3 (.7) The effective modal damping factors are ξ RQ = c 5(ω ) = c, ξrq = 5c (ω ) 3 = c. (.8) Ther IC required for solving the ODE (.7) are, from Lecture 0, η(0) = Φ T M u(0) and η(0) = Φ T M u(0), and replacing numbers: [ ] 6 [ ][ ] (u (0) + u (0)) η0 η 0 = = 6 η 0 0 u (0) = 6 0 u (0) (u (0) u (0)), [ ] 6 [ ][ ] ( u (0) + u (0)) (.9) η0 η 0 = = 6 η 0 0 u (0) = 6 0 u (0) ( u (0) u (0)), Once the solutions η RQ (t) and η RQ are obtained thay can be combined via the modal matrix as u RQ (t) = Φ η RQ to produce the response in physical coordinates. Do not confuse with the Rayleigh damping, also known as proportional damping, approach. Same name, different matrices.

12 Lecture : MODAL ANALYSIS OF MDOF FORCED DAMPED SYSTEMS.. Comparison of Exact Versus Diagonalized Responses This will be demonstrated in class using Mathematica from a laptop if there is time, else it will be skipped. The damping level will be varied as well as the IC and forcing function.