Case Study - Fluidic Instability

Transcription

1 Case Study - Fluidic Instability Date : October 6, 2014 Fluidic Instability - Its Detection, Causes and Rectification Fluidic instability is one such malfunction in rotary machines which is uncommon and need proximity shaft vibration data for its detection and data analyses. There are two major agents of fluidic induced instability in rotary machines- Lube Oil induced, Process fluid induced. This paper discusses about the Lube Oil induced fluidic instability in sleeve / journal bearings, its detection & analyses through shaft vibration data and rectifications. Lube Oil Instability is one of the most detrimental malfunc-tions which are sudden and very sensitive to its depending factors. When they remain unattended, would cause severe damage to bearings and ro-tor and other casing components at times. Modern age technology has identified its root causes and has upgraded the machines with suitable bearing modifications and lubrication methodology. Still there are field cases where slight changes in those depending factors can cause significant impact on the system dynamic stiffness knowingly or unknowingly. Machine operators sometimes fail to bridge between these two things and end up in high rotor vibrations. This paper involves those case studies which show how changes in lube oil parameters, bearing design etc can affect the system damping causing fluidic instability and thereby high vibra-tions in the rotor. The case studies are supported with vibration data plots from online diagnostics tools like ADRE and System1 of Bently Nevada (Division of GE Measurement & Controls). Case Study presented by: Padmanabhan G 1.0 Introduction All rotating machines experience various forces during its operation or mo-tion. These forces can act in various directions and can be static or dy-namic in nature [1]. These forces are the causes which produces the effect in terms of vibration. To balance these forces and to keep the rotor in equilibrium within the stator, there is something called a complex dynamic stiff-ness term. Hence the final vibration vector will surely depend upon the forces versus the dynamic stiffness vectorial terms. The below is a simple equation of Synchronous Response Motion of the rotor: (1) Synchronous Response Motion R = Unbalance Force F /Dynamic Stiffness D The complex dynamic stiffness terms comprises of two major stiffness contributors. They are the direct stiffness and the quadrature stiffness. The following is the simple equation of the Dynamic Stiffness term [1]: (2) Dynamic Stiffness D = (K- M?2) + jd (1-?)??'- The term Lambda, is the Fluid Circumferential Average Velocity Ratio.

2 Not only unbalance forces, but changes in dynamic stiffness conditions also contribute to changes in machine vibration behavior. As explained above during steady state and normal operation, the spring stiffness and Mass stiffness doesn t change as they are the design material properties of the rotor. The most vulnerable term to change and sensitive to external factors is the quadrature stiffness. For a rotor running with constant speed or angular velocity?, its quadrature stiffness depends upon the terms? - Lambda and D the damping of the system. 2.0 Fluidic Instability Introduction The simple rotary machine has a stator and a rotor supported with fluid film bearings, seals etc. The bearings carry lubricating oil which supports the rotor in all operating conditions and the seals cover both ends of the con-trolled process fluid zone within the rotor and stator area. Thus both seals and bearings represent a simple model of a cylinder within which another cylinder rotates and fluids are entrapped in between them [4]. A plain sleeve bearing supports the weight of the rotor by developing a hydrody-namic pressure in the converging wedge formed by the shaft and the bearing surface. 3.0 Case Studies There are 3 case studies discussed in this paper as follows [5]: Case Study- 1 This case study was about a Steam Turbine driving a Generator with be-low machine train diagram Figure 4.

3 Fig. 4 Machine Train Diagram- Steam Turbine Generator The machine was operating at steady state with normal vibrations since long time. From 4th to 10th June 2013 vibrations increased gradually with fluctuating behavior for HP/IP and LP Turbine Bearings 1, 2 & 3. Since the increased maximum amplitudes were just within trip levels the machine didn t trip. Lube oil temperature & pressure and bearing temperatures, steam parameters were checked and observed normal. From 11thJune 2013 onwards the vibration trends came back to normal values. Not1X component (Forward 0.5X) increase was observed during this period as shown in below Figure 5.

4 Fig. 5 Direct Vibration Trend (Left) versus Not 1X Component (Right) increase. Waterfall Plot (bottom) showing Forward 0.5X component in Brg 3 similar to Brg 1 & 2 Data Analyses showed circular orbits with scattered Keyphasor dots excit-ing X frequency all indicating towards Fluidic Instability symptoms. The instability was severe enough to produce suspected rub also as shown in below Orbit Plots in Figure 6. Fig. 6 Direct Orbit in Left) showing suspected rub (Blue) versus normal (Orange). Direct Orbit in Right showing fluidic instability as the cause of rub Data Analyses showed circular orbits with scattered Keyphasor dots excit-ing X frequency all indicating towards Fluidic Instability symptoms (Oil Whip). Rectification: It was analyzed that the moisture content was ob-served gradually increasing during 4th to 10th June 2013 causing this vibra-tion increase. Early monsoons were suspected as cause for this. In-creased moisture content in oil could have deteriorated the lube oil viscosity hence affecting damping triggering fluidic instability in the bear-ings. As and when the moisture % was restored to normal values by centri-fuge operation the vibrations came back to normal Case Study- 2 This case study was about an Expander assembly with below machine train diagram in Figure 7.

5 The expander first Stage underwent a bearing (tilt-ing pad) replacement due to minor damage in pads. During restart ma-chine was tripping just before reaching full speed where sudden Not 1X component increasing from 20 to 110 microns within 10 sec.

6 Fig. 7 Machine Train Diagram- Motor- Expander Forward X component suddenly appearing near to full speed with circular orbits and scattered Keyphasor dots. Sudden shaft centerline movement also observed during the time indicating bearing unloading or equilibrium disturbance. The below Figure 8 shows relevant data plots: Fig. 8 Waterfall Plot (Reverse 0.438X due to reverse probe connection). Below: Left- Direct Orbit with scattered KPH dots, Right- Shaft Centerline Plot with movement.

7 Data analyses showed Fluidic Instability symptoms (Oil Whip) in the new tilting pad bearing. Similar symptoms were observed before in 2006 with similar frequency excitation. Rectification: In 2006 also, the new bearing showed similar symptoms when replaced, and nozzle & pads size modifications were done that time and managed. Hence new bearing installed in 2012 didn t have that modi-fication hence the same was replaced with old one and machine was successfully restarted Case Study- 3 See Figure 9 for a Steam Turbine driving Compressors & Expander, suf-fering from varying vibrations with respect to ambient conditions. This was in May 2013 month where summer peaks in India. Fig. 9

8 Machine Train Diagram- Steam Turbine- Compressor- Expander Figure 10 shows Not 1X varying linearly with ambient temperature. Direct orbits with loops & multiple Keyphasor dots indicating Oil Whirl. Fig. 10 Left- Not 1X trend linear with ambient conditions, Right- Direct orbit High vibra-tion with internal loops (Blue) v/s Normal orbit (Orange) Rectification: After oil cooler cleaning in June 13 there was 2-3 deg reduc-tion in oil inlet temperatures. Also the 0.5X component almost disappeared after oil cooler cleaning as shown in Figure 11.

9 Fig. 11 Waterfall plots showing 0.5X excitation before v/s after Oil Cooler Cleaning 4.0 Data Analyses & Results In each of the cases discussed above the vibration pattern was distinct in the way which the data plots matched with the theoretical concepts. The frequency excitations were distinct as each of them was relating directly to rotors natural frequencies with typical orbit behavior. Variation of process fluid flow or pressure (steam or gas) was not responding to vibration pat-tern, hence process fluid instability was ruled out and Lube Oil induced in-stability was detected. To summarize the data analyses & results see be-low Table 1: Case Study & Malfuncti on Root Cause Brg Type. Symptom Ca se-1 Elliptical Journal Brg Ca se-2 Tilting Pad Brg Case-3 Plain Sleeve Brg Oil Whip Oil Whip Oil Whirl 5.0 Conclusion Increased Moisture Level in Oil. Rectificati A on ffe cted Parameter Continuou Viscosity s Lube hence? Centrifuge Operation. Improper Installed Oil Noz-zleold & Pad dim bearing en-sions inwhich had Bearing. Poor Oil Cooler Ap proaches. proper nozzle & pad size Lube Oil Cooler Cleaning. Fluid Flow Wedge-D Oil Temp,? Fluidic instability is so detrimental and can even lead to permanent dam-age of the machine or its components. It is even more dangerous to oper-ate the machine without shaft vibrations measurement. However modern technology & design accommodate protects and monitors these problems, but operational efficiency should also be accurate enough to support it. It could be

10 Powered by TCPDF ( O understood that in the Case 1 & 3 the bearing types were Ellipti-cal & Plain Sleeve bearings and they are susceptible to fluidic instability. In fact in both the Cases 1 & 3 the rotor position was already operating near to center and was very much vulnerable to instability, with any minor changes in lube oil parameters. In Case 2 the bearing type was tilting pad bearing and even they are also not safe if not designed properly. It is not just important to maintain the right lube oil and bearing types in place but it is also equally critical to maintain the lube oil parameters within limits and ensure 100% reliability in bearing design & quality as even minor changes in these can affect either Damping Coefficient D or Lambda?. It will just take some seconds to excite Fluidic instability and put the machine into dangerous situation. References [1] Donald E Bently, C. Hatch, Fundamentals of Rotating Machinery Diagnostics [2] Donald E Bently, March Dynamic Stiffness in Whirl & Whip. Orbit Article- GE Measurement & Controls. [3] Agnes Muszynska, September Oil Whip of a poorly supported bearing. Shaft Centerline Article, Bently Nevada, U.S.A. [4] Luis San Adres, Hydrodynamic Fluid Bearings and their effect on the Stability of Ro-tating Machinery. Texas A & M University, U.S.A. [5] Padmanabhan G, Data & Observations from various case studies, GE Oil & Gas, Measurement & Controls, India.