Nonlinear Vibration Analysis for a Jeffcott Rotor with Seal and Air-Film Bearing Excitations

Abstract

The nonlinear coupling vibration and bifurcation of a high-speed centrifugal compressor with a labyrinth seal and two air-film journal bearings are presented in this paper. The rotary shaft and disk are modeled as a rigid Jeffcott rotor. Muszynska's model is used to express the seal force with multiple parameters. For air-film journal bearings, the model proposed by Zhang et al. is adopted to express unsteady bearing forces. The Runge-Kutta method is used to numerically determine the vibration responses of the disk center and the bearings. Bifurcation diagrams for transverse motion of the rotor are presented with parameters of rotation speed and pressure drop of the seal. Multiple subharmonic, periodic, and quasiperiodic motions are presented with two seal-pressure drops. The bifurcation characteristics show inherent interactions between forces of the air-film bearings and the seal, presenting more complicated rotor dynamics than the one with either of the forces alone. Bifurcation diagrams are obtained with parameters of pressure drop and seal length determined for the sake of operation safety.

1. Introduction

The motion stability of high-speed rotor systems has drawn extensive attention throughout the past several decades. It is now well known that the stability of the rotor’s equilibrium can be lost as a result of the Hopf bifurcation, which leads to finite-amplitude whirls of oil-film inside the bearings. The mechanism of oil whips developed from escalating whirling motions has been thoroughly investigated both experimentally and theoretically (see, e.g., [1–4]). Several models have been developed to investigate oil-film forces of short bearings and bearings with finite lengths [5–9]. Various studies of the oil-film forces were carried out to present nonlinear vibrations, for example, super- and subharmonic motions, of the rotor system related to the bearing dynamics [10–12]. Aside from the bearing forces, seal forces play significant roles in vibration and stability of air compressors and steam turbines. Seal forces are usually generated due to the fluid-solid interaction in the clearance between the shaft and the stator which may cause self-excited motions of the rotor. Previous investigations showed that the seal force provides not only supportive reactions to the rotor in the radial direction but also cross-coupling forces in the tangential direction that excites severe vibrations in some occasions. An effective model was proposed by Muszynska to express nonlinear seal forces based on experimental results [13, 14]. This model was later adopted by Ding et al. [15] in their study on the Hopf bifurcation of a symmetric rotor-seal system and by Hua et al. who numerically obtain the nonlinear vibration and bifurcation characteristics of an unbalanced rotor-seal system [16]. Similar research was provided in Zhang et al. [17] where subharmonic motions and bifurcation diagrams were demonstrated with parameter of rotation speed. In spite of the numerous publications that separately dealt with rotor-bearing and rotor-seal systems, very few literatures have been focused on the dynamics of rotor-seal-bearing systems which is a great concern of air-compressor and steam turbine engineers. It is worth emphasizing that the interaction between the seal and bearing excitations should not be ignored since complicated, large-amplitude motions can be developed for rotors of compressors and turbines.

The numerical analysis for nonlinear vibration and bifurcation behavior of a high-speed centrifugal compressor with a labyrinth seal and two journal bearings is presented in this paper. What differentiates the current rotor system from others is the application of air-film bearings rather than conventional oil-film journal bearings. Practically, compressors supported by this kind of bearings operate under circumstance where only inflammable lubricants (i.e., air or pure water) are allowed. It should be noticed that the air-film bearings complicate the dynamics of the rotor in two aspects: (1) since the viscosity of the air is very small, the amplitude of whirling orbit is remarkably large, which brings rich nonlinear characteristics into the rotor response; (2) the airflow inside the clearance of journals is much more irregular and turbulent than the oil-film bearings, which makes most of existent theories unable to provide realistic prediction of the bearing dynamics. In the first case, the vibration response is strongly nonlinear and must be solved numerically with consideration of both bearing and seal forces. In the second case, an effective model for unsteady air-film force should be adopted to express time-varying boundaries of the film that whirls rapidly around the journal center. In the present study, the oil-film force proposed by Zhang et al. [18, 19] is used to model the nonlinear, unsteady air-film excitation in the current study. For the seal force Muszynska’s model is adopted with parameters of pressure drop, rotation speed, and seal length. The complexity in the rotor motion is demonstrated through bifurcation diagrams with those parameters as well as through the Poincaré maps, time history of displacement, and rotor orbits. For seal pressure drop of 0.2 MPa the bifurcation sequence is given with increasing rotation speed, showing subharmonic motions of periodic-1, 12, 11, 10, 9, 8, 7 and quasiperiodic motions. The results are compared to the ones without bearing forces to present the interaction between the air-film bearing and the seal forces. Periods-4 and -11 bifurcations and quasiperiodic motion are observed with a 0.4 MPa pressure drop. The bifurcation diagrams of motion with parameters of pressure drop and length of the seal provide suitable values of these quantities for improvement of operation safety of the machinery. The intricacy in the motion’s bifurcation presents complicated dynamics of the system in contrast to the rotors with either of bearing forces or of the seal excitations.

2. Problem Modeling

A Jeffcott rotor with a rigid disk, a segment of labyrinth seal, and two supporting air-film journal bearings is shown in Figure 1, where 𝑜₁ is the geometric center of the disk; 𝑜₂ and 𝑜₃ are centers of the left and right bearings. Denote by (𝑥₁,𝑦₁), (𝑥₂,𝑦₂), and (𝑥₃,𝑦₃) the displacements of the disk center, the left bearing, and the right journal bearing, respectively. The equation of motion of the system is expressed as follows:𝑚₁𝑥̈₁+𝐷_𝑒(𝑥̇₁−𝑥̇₂)+𝐷_𝑒(𝑥̇₁−𝑥̇₃)+𝐾_𝑒1(𝑥₁−𝑥₂)+𝐾_𝑒2(𝑥₁−𝑥₃)=𝐹_𝑥+𝑚₁𝑒𝜔²𝑚cos𝜔𝑡,₁𝑦̈₁+𝐷_𝑒(𝑦̇₁−𝑦̇₂)+𝐷_𝑒(𝑦̇₁−𝑦̇₃)+𝐾_𝑒1(𝑦₁−𝑦₂)+𝐾_𝑒2(𝑦₁−𝑦₃)=𝐹_𝑦−𝑚₁𝑔+𝑚₁𝑒𝜔²𝑚sin𝜔𝑡,₂𝑥̈₂+𝐷_𝑒(𝑥̇₂−𝑥̇₁)+𝐾_𝑒1(𝑥₂−𝑥₁)=𝑓_𝑥2,𝑚₂𝑦̈₂+𝐷_𝑒(𝑦̇₂−𝑦̇₁)+𝐾_𝑒1(𝑦₂−𝑦₁)=𝑓_𝑦2−𝑚₂𝑚𝑔,₃𝑥̈₃+𝐷_𝑒(𝑥̇₃−𝑥̇₁)+𝐾_𝑒2(𝑥₃−𝑥₁)=𝑓_𝑥3,𝑚₃𝑦̈₃+𝐷_𝑒(𝑦̇₃−𝑦̇₁)+𝐾_𝑒2(𝑦₃−𝑦₁)=𝑓_𝑦3−𝑚₃𝑔, (2.1) where 𝑚₁ is the mass of the disk; 𝑚₂ and 𝑚₃ are masses of the left and the right bearings. 𝐾_𝑒1 and 𝐾_𝑒2 are equivalent stiffness coefficients of the left and the right shafts; 𝐷_𝑒 is the factor of viscous damping; 𝑒 is the mass unbalance of the disk; 𝐹_𝑥 and 𝐹_𝑦 are directional components of the seal force; 𝑓_𝑥2,𝑥3 and 𝑓_𝑦2,𝑦3 are directional force components of the left and the right bearings, respectively. 𝜔 is the rotation speed and g is the gravitational acceleration. The symmetry of the fluid field inside the seal clearance is destroyed as the rotor is perturbed from its equilibrium position with a nonzero rotation speed. Muszynska’s model [13, 14] is used to express the seal forces in both x- and y-directions, as 𝐹_𝑥𝐹_𝑦=−𝐾−𝑚_𝑓𝜏²𝜔²𝜏𝜔𝐷−𝜏𝜔𝐷𝐾−𝑚_𝑓𝜏²𝜔²𝑥₁𝑦₁−𝐷2𝜏𝑚_𝑓𝜔−2𝜏𝑚_𝑓𝑥̇𝜔𝐷𝑦̇¹₁−𝑚_𝑓00𝑚_𝑓𝑥̈₁𝑦̈₁, (2.2) where K and D are coefficients of stiffness and damping of the air that flows through the seal clearance, respectively; 𝑚_𝑓 is the effective mass of the air; 𝜏 is the factor of average angular speed of fluid that rotates along with the rotor, determined by 𝜏=𝜏₀(1−𝜀)^𝑏, (2.3) where 𝜏₀ is the average angular speed for the unperturbed rotor; b is an empirical coefficient; 𝜀=𝑥³_2,3+𝑦²_2,3 is the nondimensional amplitude of whirling motion of the bearings. The model of the bearing force adopted in the current study is the one proposed by Zhang et al. [18, 19] for unsteady oil-film journal bearings, expressed as follows:𝑓_𝑥=−𝐶₁𝜀̇−𝐶₂(̇𝜔𝜑−2)𝜀,𝑓_𝑦=−𝐶₂𝜀̇−𝐶₃(̇𝜔𝜑−2)𝜀, (2.4) where 𝜑̇ is the whirling speed of the journal; 𝐶₁, 𝐶₂, and 𝐶₃ are damping coefficients of the lubricant [20]. Unlike most existent bearing theories that handle time-invariant boundaries of the lubricant film with, for example, the Gümbel condition and the π-oil-film assumption, the unsteady force model of (2.4) is capable of dealing with time-varying boundary of the film arising from large whirling velocity of the journal center, which is appropriate for weakly viscous systems with air- or water-film bearings such as the present one. Introducing the following nondimensional parameters:𝑇=𝜔𝑡,𝑋₁=𝑥₁𝑐,𝑌₁=𝑦₁𝑐,𝑋_2,3=𝑥_2,3𝛿,𝑌_2,3=𝑦_2,3𝛿,𝑋^′₁=𝑥̇₁(𝜔𝑐),𝑌^′₁=𝑦̇₁,𝑋(𝜔𝑐)_′2,3=𝑥̇_2,3(𝜔𝛿),𝑌^′_2,3=𝑦̇_2,3,(𝜔𝛿) (2.5) where (⋅)^′≜𝑑(⋅)/𝑑𝑇 denotes the derivative of a quantity with respect to T, and c and 𝛿 are clearances of the seal and the journal bearings, respectively, the equation of motion is then rewritten as𝑋₁^′′+2𝐷_𝑒+𝐷(𝑚₁+𝑚_𝑓𝑋)𝜔^′1+2𝜏𝑚_𝑓𝑚₁+𝑚_𝑓𝑌^′₁−𝐷_𝑒𝛿(𝑚₁+𝑚_𝑓)𝜔𝑐(𝑋^′₂+𝑋^′₃𝐾)+^𝑒1+𝐾_𝑒2+𝐾−𝑚_𝑓𝜏²𝜔²(𝑚₁+𝑚_𝑓)𝜔²𝑋₁+𝜏𝐷(𝑚₁+𝑚_𝑓𝑌)𝜔¹−𝐾_𝑒1𝛿(𝑚₁+𝑚_𝑓)𝜔²𝑐𝑋₂−𝐾_𝑒2𝛿(𝑚₁+𝑚_𝑓)𝜔²𝑐𝑋₃=𝑚₁𝑒(𝑚₁+𝑚_𝑓𝑌)𝑐cos𝑇,_1′′−2𝜏𝑚_𝑓𝑚₁+𝑚_𝑓𝑋^′₁+2𝐷_𝑒+𝐷(𝑚₁+𝑚_𝑓𝑌)𝜔^′1−𝐷_𝑒𝛿(𝑚₁+𝑚_𝑓)𝜔𝑐(𝑌^′₂+𝑌^′₃)−𝜏𝐷(𝑚₁+𝑚_𝑓𝑋)𝜔¹+𝐾_𝑒1+𝐾_𝑒2+𝐾−𝑚_𝑓𝜏²𝜔²(𝑚₁+𝑚_𝑓)𝜔²𝑌₁−𝐾_𝑒1𝛿(𝑚₁+𝑚_𝑓)𝜔²𝑐𝑌₂−𝐾_𝑒2𝛿(𝑚₁+𝑚_𝑓)𝜔²𝑐𝑌₃=𝑚₁𝑒(𝑚₁+𝑚_𝑓𝑚)𝑐sin𝑇−¹𝑔(𝑚₁+𝑚_𝑓)𝜔²𝑐,𝑋₂^′′−𝐷_𝑒𝑐𝑚₂𝑋𝜔𝛿^′1+𝐷_𝑒𝜔𝛿+𝑆₀𝐶_11𝐿𝑚₂𝜔²𝛿𝑋^′₂+𝑆₀𝐶_12𝐿𝑚₂𝜔²𝛿𝑌^′₂−𝐾_𝑒1𝑐𝑚₂𝜔²𝛿𝑋₁+2𝐾_𝑒1𝛿−𝐶_2𝐿𝑆₀2𝑚₂𝜔²𝛿𝑋₂+𝐶_3𝐿𝑆₀2𝑚₂𝜔²𝛿𝑌₂𝑌=0,_2′′−𝐷_𝑒𝑐𝑚₂𝑌𝜔𝛿^′1+𝑆₀𝐶_12𝐿𝑚₂𝜔²𝛿𝑋^′₂+𝐷_𝑒𝜔𝛿+𝑆₀𝐶_22𝐿𝑚₂𝜔²𝛿𝑌^′₂−𝐾_𝑒1𝑐𝑚₂𝜔²𝛿Y₁+2𝐾_𝑒1𝛿−𝐶_2𝐿𝑆₀2𝑚₂𝜔²𝛿𝑌₂−𝐶_3𝐿𝑆₀2𝑚₂𝜔²𝛿𝑋₂𝑔=−𝜔²𝛿,𝑋₃^′′−𝐷_𝑒𝑐𝑚₃𝑋𝜔𝛿^′1+𝐷_𝑒𝜔𝛿+𝑆₀𝐶_11𝑅𝑚₃𝜔²𝛿𝑋^′₃+𝑆₀𝐶_12𝑅𝑚₃𝜔²𝛿𝑌^′₃−𝐾_𝑒2𝑐𝑚₃𝜔²𝛿𝑋₁+2𝐾_𝑒2𝛿−𝑆₀𝐶_2𝑅2𝑚₃𝜔²𝛿𝑋₃+𝑆₀𝐶_3𝑅2𝑚₃𝜔²𝛿𝑌₃𝑌=0,_3′′−𝐷_𝑒𝑐𝑚₃𝑌𝜔𝛿^′1+𝑆₀𝐶_12𝑅𝑚₃𝜔²𝛿𝑋^′₃+𝐷_𝑒𝜔𝛿+𝑆₀𝐶_22𝑅𝑚₃𝜔²𝛿𝑌^′₃−𝐾_𝑒2𝑐𝑚₃𝜔²𝛿𝑌₁−𝑆₀𝐶_3𝑅2𝑚₃𝜔²𝛿𝑋₃+2𝐾_𝑒2𝛿−𝑆₀𝐶_2𝑅2𝑚₃𝜔²𝛿𝑌₃𝑔=−𝜔²𝛿, (2.6) where l and r are length and radius of the bearing, respectively; 𝜇 is the dynamic viscosity of the lubricant; superscripts L and R represent the left and the right bearings, respectively, and𝐶₁₁=𝐶₁cos²𝜑+𝐶₃sin²𝜑−2𝐶₂𝐶sin𝜑cos𝜑,₁₂=𝐶₂₁=𝐶₂(cos²𝜑−sin²𝜑)+(𝐶₁−𝐶₃𝐶)sin𝜑cos𝜑,₂₂=𝐶₁sin²𝜑+𝐶₃cos²𝜑+2𝐶₂𝑆sin𝜑cos𝜑,₀=6𝜇𝜔𝑙𝑟³𝛿⁻². (2.7)

Figure 1 

A Jeffcott rotor-seal-bearing system.

3. Subharmonic Motions and Bifurcation Behavior

Notice that parameters 𝐾, 𝐷, and 𝜏 and coefficients 𝐶₁, 𝐶₂, and 𝐶₃ are functions of displacements of the disk centers and the bearings. Hence, (2.6) is a group of highly nonlinear ordinary differential equations that can hardly be solved through conventional perturbation methods [21]. Instead, the vibration responses of the disk center and the two bearings are computed by using the fourth-order Runge-Kutta method with adaptive-step control to reduce local truncation error of every single step. The parameters selected for the current study are𝑚₁=50kg,𝑚₂=3.5kg,𝑚₃=3.5kg,𝐷_𝑒𝐾=3000N⋅s/m,_𝑒1=3.4635×10⁶N/m,𝐾_𝑒2=3.8127×10⁶𝑒N/m,₁=0.2mm,𝑟=0.035m,𝑙=0.06m,𝑐=0.3mm,𝛿=0.3mm,𝜇=1.47×10⁻⁵Pa⋅s,𝜏₀=0.4,𝑏=0.45. (3.1) Additionally, the length and radius of the seal are 0.102 m and 0.067 m, respectively. The system’s parameters are chosen based on a single-staged centrifugal compressor manufactured by Shenyang Turbo-machinery Cooperation. To investigate the bifurcation we chose the rotation speed as the parameter under two pressure drops of the seal, that is, the pressure differences between the entrance and the exit of the seal. The initial displacements and velocities of the disk center and the two are (0.01,0).

Let pressure drop Δ𝑃 be 0.2MPa. The bifurcation diagrams of displacement x are illustrated in Figure 2, where 𝑆=𝜔/(𝐾_𝑒1+𝐾_𝑒2)/𝑚₁ is the nondimensional rotation speed. In the current computation 𝑆=1 corresponds to a rotation speed of 60.71 Hz or 3642.77 rpm.

(a) The disk center

(b) The left bearing

(c) The right bearing

Figure 2 

Bifurcation diagrams of the rotor system with seal and air-film excitations. $\mathrm{\Delta }P=0.2$ MPa.

It can be seen that the disk and the bearings are in motions of period-1, that is, motions with the same frequency as the rotation speed, when the rotation is slow. The primary resonance happens at 𝑆=1.1272. The stability of the period-1 motions is lost at 𝑆=2.1496, and the motion becomes quasiperiodic. Various subharmonic motions can be observed when the rotation speed is increased. A period-12 bifurcation takes place at 𝑆=3.1457. Following that, the motions become quasiperiodic again with escalating rotation speed. At 𝑆=3.2505 the displacements undergo a period-11 bifurcation and return quasiperiodic with higher S afterwards. A period-10 bifurcation is encountered with speed 𝑆=3.355. The Poincaré map of displacement 𝑥₁ is presented in Figure 3(a) to show the existence of a periodic-10 motion. The time history of 𝑥₁ is illustrated in Figure 3(b), and the orbits of the disk center and the left bearing are shown in Figures 3(c) and 3(d), respectively. Further, a period-9 bifurcation is observed at 𝑆=3.5389 followed by a period-8 bifurcation at 𝑆=3.8010. The bifurcation cascade continues at 𝑆=4.0632 when a period-7 bifurcation takes place. Following that, quasiperiodic motions are obtained with higher rotation speed. Figures 4(a), 4(b), and 4(c) depict the orbits of the disk center and the right bearing as well as the Poincaré map of displacement 𝑥₁ at 𝑆=4.352, respectively.

(a) Poincaré map of displacement x1

(b) Time history of displacement x1

(c) Orbit of the disk center

(d) Orbit of the left bearing

Figure 3 

Motions of the rotor system with rotation speed $S=3.355$.

(a) Orbit of the disk center

(b) Orbit of the right bearing

(c) Poincaré map of displacement x1

Figure 4 

Motions of the rotor system with rotation speed $S=4.352$.

To investigate the interaction between the bearing and the seal forces a comparative computation is carried out for a Jeffcott rotor with two rigid supports (hence, the seal force is the only excitation of the system) and exactly the same geometrical and seal properties as aforementioned. The bifurcation diagram is shown in Figure 5. For rotation speed less than 𝑆=1.432, the motion is period-1 with the same frequency as the rotation speed. With an increasing speed, the motion remains quasiperiodic up to 𝑆=3.52, where a period-8 bifurcation is observed from the disk’s displacements. The motions turns into quasiperiodic again with advancing rotation speed. The comparison between the responses to the coupling forces and to the seal force alone reveals rich bifurcating behavior of the system vibration: the interaction of the seal and the air-film forces results in more period-multiple bifurcations (see Figures 2(a) and 5).

Figure 5 

Bifurcation diagram of the disk center with rigid supports.

We now change the pressure drop of the seal to 0.4 MPa. The bifurcation diagrams of displacement x of the disk and the two journal bearings are presented in Figure 6.

(a) The disk center

(b) The left bearing

(c) The right bearing

Figure 6 

Bifurcation diagrams of the rotor system with seal and air-film excitations. $\mathrm{\Delta }P=0.4$ MPa.

It is found that the x-directional displacements of the disk and the bearings are period-1 with small rotation speed. The primary resonance in the motion is found at 𝑆=1.2582. Then, the bifurcation starts and the motions become quasiperiodic. A period-4 bifurcation takes place with speed 𝑆=2.0709 followed by quasiperiodic motions as the rotor is accelerated. For speed 𝑆∈[3.2243,3.3816]∪[3.4340,3.6437] the motions are period-4. Figures 7(a) and 7(b) show the orbits of the disk center and the left bearing at 𝑆=3.4340. Figure 7(c) depicts the Poincaré map of the disk motion. The motions become quasiperiodic with higher rotation speed. Figures 8(a) and 8(b) plot the orbits of the disk center and the right bearing at 𝑆=3.8797. The Poincaré maps of displacements 𝑥₁ and 𝑥₃ are shown in Figures 8(c) and 8(d), respectively. A period-11 bifurcation is observed at 𝑆=4.2204 followed by another series of quasiperiodic motions. With higher pressure-drop from the entrance to the exit of the seal, some previously notified bifurcations are not observed again. Nevertheless, the bifurcation behavior is still more complicated than the one with the seal force only.

(a) Orbit of the disk center

(b) Orbit of the left bearing

(c) Poincaré map of displacement x1

Figure 7 

Motions of the rotor system with rotation speed $S=3.434$.

(a) Orbit of the disk center

(b) Orbit of the right bearing

(c) Poincaré map of displacement x1

(d) Poincaré map of displacement x3

Figure 8 

Motions of the rotor with rotation speed $S=3.8797$.

In the following analysis we adopt the pressure drop as the bifurcation parameter. Let rotation speed 𝜔 be 1200rad/s. The bifurcation diagrams of x-displacements of the disk center and the left and the right bearings are presented in Figure 9. For low-pressure drops the motions are found quasiperiodic with large amplitude until ∆P up to 0.048 MPa. The motions of the disk center and the bearings then become period-1, and the amplitudes step up with the advancing pressure drop. The synchronous motions are lost at a critical drop Δ𝑃=0.168 MPa without undergoing primary resonances in the motions. The vibrations afterwards are basically quasiperiodic, and it is very difficult to distinguish the bifurcation points. The average amplitudes of the displacements remain almost unchanged with increasing pressure drops, showing the remarkable air-film whip in the journal bearings. This implies that the whole system cannot be stabilized by increasing the pressure drops larger than the critical value.

(a) The disk center

(b) The left bearing

(c) The right bearing

Figure 9 

Bifurcation diagram of the rotor with varying pressure drop. $\omega =1200\hspace{0.17em}\text{rad}/\text{s}.$

Finally, the evolution of the bifurcation in the rotor motions is investigated by taking the length of the seal as the control parameter. Let rotation speed 𝜔 be 1200rad/s and let pressure drop Δ𝑃 be 0.2 MPa. The bifurcation diagrams of 𝑥-displacements of the disk center and the left and the right bearings are depicted in Figure 10. The period-1 motion is found for length: 0.082m≤𝑙≤0.098m, where the orbital whirling motions grow monotonously with the increasing seal-length. Beyond this range of length, the motions are mainly quasiperiodic with considerably large amplitudes. Therefore, a suitable length of seal should be chosen between 0.082 m and 0.098 m to keep the rotor distant from strong vibration responses that may jeopardize the safety of the machine in operation. From the manufacturer’s point of view, a labyrinth seal with a medium length of between 0.082 m and 0.098 m is feasible for it can be conveniently processed, assembled, and positioned by using conventional tools.

(a) The disk center

(b) The left bearing

(c) The right bearing

Figure 10 

Bifurcation diagram of the rotor with varying seal-length. $\omega =1200\hspace{0.17em}\text{rad}/\text{s}\text{;}$$\mathrm{\Delta }P=0.2$ MPa.

4. Conclusions

The nonlinear coupling vibration excited by a labyrinth seal and two air-film journal bearing is investigated through numerical simulations for high-speed centrifugal compressors. The results obtained with various rotation speeds and seal pressure drops show complexity of nonlinear vibration and bifurcation behavior in the displacements of the rotor system. Further, the motions of the system reveal period-multiple bifurcations compared to the system excited only by the seal force, presenting an intricate interaction between the seal and the bearing forces. Suitable seal pressure drop and seal length are determined for the sake of operation safety through the bifurcation analysis for rotor displacements as well.

Acknowledgments

The authors gratefully acknowledge the Natural Science Foundation of China (Projects 10472021, 10721062), the Chinese National Programs for High Technology Research and Development (2007AA04Z405), and the State Key Development Program for Basic Research of China (Project 2009CB724300) for their fundings.