Process Engineering Equipment Handbook Episode 1 Part 10 potx

C-292 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) FIG. C-321 Vibration spectrum (rpm = 20,000, P d = 1200 psig). 13 FIG. C-322 Vibration spectrum (rpm = 20,000, P d = 1250 psig). 13 FIG. C-323 Vibration spectrum (rpm = 20,000, P d = 1270 psig). 13 FIG. C-324 Vibration spectrum (rpm = 20,000, P d = 1320 psig). 13 C-293 1200 psig discharge pressure. Note a synchronous peak of 0.5 mil at 20,000 rpm (possibly due to unbalance). In Fig. C-324 machine rpm and suction pressure stay unaltered. Discharge pressure has been raised to 1250 psig. Now a 0.2 mil subsynchronous component shows up at 9000 rpm. Frequently such components may be intermittent and hard to capture without the use of the peak hold mode on the analyzer. In Fig. C-323 with all other conditions remaining unchanged, it is noted that just a 20 psig increase in discharge pressure raised the 9000 rpm component from 0.2 mil to 1.5 mil. It took an increase in suction pressure by 50 psig while maintaining the same discharge pressure to allow the unit to regain stability. This shows the importance of keeping track of process conditions in addition to mechanical items that change within the machine system. Other frequency orders. When originated within the machine, these are generally related to meeting contact surfaces. Blades. For instance, blade tip rub would show up a signal of “number of teeth in the ‘contacting’ stage ¥ rpm.” Figure C-325 shows an accelerometer’s signature from an axial flow compressor with strong frequency component of the first, second, and third harmonic of the fifth-stage stator blade row. An inspection of this stator row indicated cracks caused by high-cycle fatigue. Gears. A frequency of “number of gear teeth ¥ rpm” may indicate resonance with the natural frequency of the concrete foundation, a fabricated base, or supporting beams. C-294 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) FIG. C-325 Axial-flow compressor spectrum showing blade-passing frequency. 13 Figures C-326 and C-327 are typical of the kind of information that can be provided by accelerometers. These data would not be possible in the low-frequency spectra provided by proximity probes. Figure C-326 shows two gears in good condition (accelerometer is at the low-frequency end of the gearbox). Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) C-295 FIG. C-326 Gearbox signature (low-frequency end). 13 FIG. C-327 Gearbox signature (high-frequency end). Potential for damaged tooth. 13 Figure C-327 shows a problem with gear A that may be a chipped or cracked tooth. A frequency of 1¥ rpm may be observed when a 1¥ rpm signal elsewhere in the machine (e.g., an unbalanced orbit) is transmitted to a gear and gets it to run eccentrically. This eccentric running in turn produces a 1¥ rpm signal with a high amplitude in the direction of the imaginary line joining the centers of the two mating gears. At 90° to this position, the 1¥ rpm signal would be lower, producing an orbit that is a flat ellipse. This orbit in turn magnifies any resonance to gear mesh frequency if present. When troubleshooting resonance, remove sources for lower order frequencies first, to help analyze the higher frequency vibration. If gear misalignment occurs, a vibration at 2¥ rpm would probably show up. To prove misalignment, use Prussian blue to coat the gear surfaces and run for a few minutes. A clear indication of contact surfaces and wear pattern shows when the gears are examined. Loose assembly. Looseness of a part generally causes vibration at 2¥ the frequency of the rotating part. To visualize this, consider the case of a loose machine base and compare it to a bench that has two uneven supports. First one support touches in a cycle, then the other. So the frequency is 2¥ rpm. Drive belts. If drive belts are loose, a vibration is caused. This can be observed with a strobe shone on a reference mark on the belt. This is often confused with an unbalanced belt and belts are thrown away unnecessarily. Sources of a vibration-causing force from outside the machine include: 1. Piping stresses—static, cantilevered 2. Foundation problems ᭿ Foundation settling ᭿ Frost melting/permafrost problems ᭿ Moving soil (muskeg or other shifting soil insufficiently removed) ᭿ Foundation inclusions (grout problems, soft feet, and so forth) 3. Extreme climatic change Contingency measures in a mature model (old or approaching scheduled overhaul) and retrofit. In the case of a mature operating model, the problems prevalent with a new train or prototype model can give way to those posed by: 1. Changing field composition 2. Changing environmental regulations: new burner designs, water and steam injection to reduce NO x , and so forth These changes may or may not have anything to do with the aerodynamic and mechanical behavior of the machine in question. At any rate, they will have to be analyzed as they come up and an attempt should be made to approximate some budget figures for retrofit items that are anticipated due to tightening legislation. Vibration signatures may also be a good indicator of when a machine is approaching the point of requiring overhaul. See Figs. C-328 and C-329. Figure C- 328 compares baseline signature with one taken after two years of operation. The increase in high-frequency levels was found attributable to blade flutter caused by cracked blades. Figure C-329 is similar. It shows an increase in the component due to one stator stage’s resonant frequency, indicating high blade flutter, that was found to be caused by cracks in that stator. C-296 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) Precautions on new turbomachinery. To help avoid problems on new turbomachinery: 1. Ask that vibration specifications be included with preliminary information— prior to formal quote request stage. 2. Make sure vibration specifications include data on allowable vibration levels, types of probes to be used, and whether the probes are seismic or proximity. 3. Arrange to be present for in-factory balance tests as well as final (full or partial load) tests. Final assembly should also be witnessed if at all possible, particularly in the case of a prototype. Baseline signatures should be taken. 4. Be aware of machinery shipping plans. 5. During commissioning, run the drive unit alone and take vibration readings. 6. Run the drive unit with the machine coupled, but not in complete process loop. 7. Record readings during machine rundown. Audit surrounding systems, piping, and so forth. 8. Before the machine arrives on site, check piping, grouting, and so forth. 9. Do hot and cold alignments on the train. 10. With electric motors, ensure half key is allowed for during balance. Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) C-297 FIG. C-328 Machinery analyses showing comparison of baseline signature to signature before overhaul. 13 Troubleshooting philosophy In any problem situation, the indicators may include vibration readings as well as gas path parameters. There may also be other indicators, such as bleed valve behavior, bearing cavity temperature, and so forth. These other measured quantities may not be conveniently available (although they may be monitored at some intervals) in a nonexpert comprehensive system. However, most problems— over 95 percent for nonprototype applications—that occur with turbomachinery can be solved with good VA and PA data. A basic philosophy for troubleshooting is as follows: 1. Spend money on diagnostic equipment only if you can use and interpret the data. If you are new to troubleshooting, VA, and so forth, get help, but with a view to learning how to do all this yourself. A good troubleshooter has the right ᭿ Physical equipment ᭿ Mental knowledge ᭿ Relevant training 2. One of the things that is really useful to have is a portable spectrum analyzer. If you have a vibration system already installed, but you need to see if it would benefit you to ᭿ Retrofit more probes on your installation ᭿ Work out how many you need for similar future installations then a portable system with: ᭿ A portable probe or probes (velocity or acceleration transducer) ᭿ A spectrum analyzer, including storage capacity to store successive plots and a chart recorder to make a hard copy of the spectrum C-298 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) FIG. C-329 Machinery analyses showing comparison of baseline signature to signature before overhaul. 13 is very useful. You can now build up your own store of information on every item of machinery you are responsible for. 3. You should study the instrumentation—OEM-supplied or otherwise—on your installation and learn about its accuracy, usefulness, and ability to have its signal fed into a retrofitted PLC (programmable logic controller) or a computer. Consider what additional instrumentation, if any, might be useful. 4. Concentrate on gas path monitoring parameters as these are the most useful. Generally, most systems, however basic, supplied by an OEM will have enough data for you to fit a PA system. This is useful for ᭿ Determining the health of the gas path ᭿ Helping diagnose failed blades, combustion liners, crossover tubes, and so forth ᭿ Determining when a module (compressor or turbine) needs to be washed ᭿ Determining if premature shutdown/maintenance is required For further discussion of PA systems, see Life-Cycle Assessment (LCA). 5. Consider what the return on investment (ROI) might be if you were to get a comprehensive online (perhaps real-time) condition-monitoring system. Consider also if it would ever make life trouble-free for the operator. Problem diagnosis. Let us assume that a problem has occurred. Ask these questions with reference to the occurrence: ᭿ What? ᭿ When? ᭿ How? ᭿ Why? Then ask: ᭿ What needs to be done? ᭿ When? ᭿ How? ᭿ Why? ᭿ Will this affect anything else? ᭿ What is the cost of doing nothing? ᭿ How much production will we lose meanwhile? ᭿ Can we correct anything else while we correct this problem? ᭿ What can we learn for future installations? Summary rules 1. There is no one consistently right answer for any symptom in condition monitoring. 2. Separate the elements of plant, process, and personnel. 3. Do not spend money to get more data than you can thoroughly understand or be taught to understand. 4. Fully automated intelligent systems might not be worth the money. 5. When you think you know all the answers, see rule 1. Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) C-299 Tables C-21 through C-27 give commonly accepted guideline limits for vibration readings. These limits apply to turboexpanders and all associated machinery in the process train. Figures C-330 through C-333 are a few of the diagnostic charts available in industry. They are not new but then neither is much of the machinery being monitored in older plants. There is no hard-and-fast rule about which is best. Knowledge of a particular machine and process determines which are appropriate. Guide. Note that the limits expressed in Fig. C-332 are based on experience in refineries. This guide reflects the typical proximity probe installation close to and supported by the bearing housing and assumes the main vibration component to be of 1¥ rpm frequency. The seemingly high allowable vibration levels above 20,000 rpm reflect the experience of high-speed air compressors (up to 50,000 rpm) and jet engine–type gas turbines with their light rotors and light bearing loads. Readings must be taken on machined surfaces with runout less than 0.5 mil up to 20,000 rpm and less than 0.25 mil above 12,000 rpm. Warning. Judgment must be used especially when experiencing frequencies in multiples of operating rpm on machines with standard bearing loads. Such machines cannot operate at the indicated limits for frequencies higher than 1¥ rpm. In such cases, enter the graph with the predominant frequency of vibration instead of the operating speed. C-300 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) TABLE C-21 “Normal” Vibration Levels on BRG Housings in IPS (Peak) Highest Noted on Smooth Machine 20 Machine Type 1 2 3 4 VP1 VP2 GMF1 GMF2 BP1 BP2 Blowers (6000 rpm max) .05 .02 .01 .01 .04 .01 Horizontal centrifugal compressors .05 .02 .01 .01 .04 .01 Barrel compressor .03 .01 .005 .005 .05 .005 Gears Parallel Shaft .1 .05 .02 .02 .05 .02 Epicyclic .05 .02 .02 .02 .05 .01 Steam turbines .1 .02 .02 .02 .05 .01 Gas turbines/axial compressor .2 .02 .01 .01 .05 .01 Pumps .1 .05 .01 .01 Motors .1 .1 .05 .05 LP1 LP2 LP3 LP4 LP5 Screw compressor .1 .01 .1 .1 .1 .05 .05 VP = vane pass. BP = blade pass. GMF = gear mesh frequency. LP = lobe pass. [...]... purp Parallel, gen purp Epicyclic 1 Gear Mesh Freq 1 Blade Pass 10 10 10 10 10 05 05 Intermed Freq 10 00–3000 Hz 15 25 20 15 15 50 15 40 10 25 10 25 10 1 Lobe Pass Screw compressors 2¥ Gear Mesh Freq 2¥ Lobe Pass 3¥ Lobe Pass 4¥ Lobe Pass 5¥ Lobe Pass 20 20 20 20 20 25 20 05 10 10 15 10 05 05 Note that filtered components add up to unfiltered total amplitude of vibration (1) The significance of vane, blade,... Type 1 rpm 2¥ rpm 3¥ rpm 4¥ rpm 1 Vane Pass 2¥ Vane Pass Blowers, up to 6000 rpm maximum 50 40 25 25 10 050 Contrifugal compressors Horizontal Barrel 25 15 20 10 15 10 15 10 10 05 050 025 Steam turbines Special purpose General purpose 30 50 25 40 15 25 15 25 Gas turbines and axial compressors 50 40 25 25 Contrifugal pumps Between brgs Overhung type 25 50 20 40 15 25 15 25 Electric motors 25 20 15 15 ... Condition–Monitoring System(s) C-3 01 TABLE C-22 Maximum Allowable Vibration Limits on BRG Housing in IPS (Peak) for Operation Up to Earliest Possible Corrective Shutdown20 Machine Type 1 2 3 4 VP1 VP2 Blowers (6000 rpm max) 5 4 25 25 1 05 Horizontal centrifugal compressors 25 2 15 15 1 05 Barrel compressor 15 1 1 1 05 025 Gears Parallel Shaft 25 2 15 15 1 05 Epicyclic 15 1 1 1 1 05 Steam turbines 25 2 15 Gas turbines/axial... 1 1 05 025 Gears Parallel Shaft 25 2 15 15 1 05 Epicyclic 15 1 1 1 1 05 Steam turbines 25 2 15 Gas turbines/axial compressors 5 4 Pumps 25 Motors 25 15 1 05 25 25 1 05 2 15 15 2 25 BP2 15 15 LP1 Screw compressors BP1 2 2 1 LP3 LP4 2 2 2 GMF2 05 LP2 GMF1 LP5 2 VP = vane pass BP = blade pass GMF = gear mesh frequency LP = lobe pass Introduction to rotor dynamics: vibration theory The two main categories... System(s) C-309 FIG C-333 General machinery vibration severity chart .15 which, if we substituted into the previous equation, we get Ê r 2 + c r + k ˆ e rt = 0 Ë m m¯ (C -10 ) c2 k 2 4m m (C -11 ) which, if we solve for t r1,2 = -c ± 2m or in more general terms x = e - c 2 m )t c1e[ ( [ c2 ( 4 m2 )-( k m ) ]t + c2 e [ c2 ( 4 m2 )-( k m ) ]t ] (C -12 ) C- 310 Condition Monitoring; Condition-Monitoring System(s); Engine... velocity, and acceleration .13 FIG C-338 Single degree of freedom system (spring mass system) .13 FIG C-339 Free vibration with viscous damping .13 C- 311 C- 312 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) FIG C-340 Overdamped decay .13 FIG C-3 41 Critical damping decay .13 We now define the damping factor t= c cc (C -14 ) to specify the amount... Condition–Monitoring System(s) C-325 FIG C-3 51 Accelerometer locations on impeller tested .13 FIG C-352 Impeller showing nodal points .13 In the case, for instance, of an impeller with 20 blades, a design speed of 3000 rpm (or 50 Hz), and a critical frequency of 10 00 Hz, the impeller may have a serious problem at these conditions (the intersection of the running speed line, 10 00-Hz frequency line, and line of... q) (C -18 ) C- 314 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) FIG C-343 Forced vibration system .13 FIG C-344 Free body diagram of mass (M) .13 where D is displacement of the steady-state oscillation Motion lags force by q So for velocity and acceleration, we have p ˙ q = x = Dw cos(wt - q) = Dw sinÊ wt - q + ˆ Ë 2¯ (C -19 ) p ˙˙... r1,2 = ±i (C -15 ) Then the response becomes [ ( x = e - c 2 m )t c1e i ( ( k m )- c 2 4 m2 ) + c2 e - i ( ( k m )- c 2 4 m2 ) ] which can be written ( x = e - c 2 m )t [ A cos w d t + b sin w d t] (C -16 ) where x is the response amplitude Forced vibration In forced vibration there is an external excitation force See Figs C-343 and C-344 Now the equation of motion is ˙˙ ˙ mx = F sin wt - kx - cx (C -17 )... mwn = critical damping coefficient) D= F k 2 w2 ˆ Ê w ˆ Ê 1 - 2 + 2z Ë wn ¯ Ë wn ¯ w wn tan q = 2 Ê w ˆ 1 wn ¯ (C-24) 2z (C-25) C- 316 Condition Monitoring; Condition-Monitoring System(s); Engine Condition Monitoring; Engine Condition–Monitoring System(s) c k = natural frequency and z = = damping factor cc m We can see that damping has a large part to play in determining amplitude and phase angle in . .20 .15 .15 .10 .050 Barrel .15 .10 .10 .10 .05 .025 Steam turbines Special purpose .30 .25 .15 .15 .10 General purpose .50 .40 .25 .25 .10 Gas turbines and axial compressors .50 .40 .25 .25 .10 Contrifugal pumps Between. .2 .15 .15 .1 .05 Epicyclic .15 .1 .1 .1 .1 .05 Steam turbines .25 .2 .15 .15 .1 .05 Gas turbines/axial compressors .5 .4 .25 .25 .1 .05 Pumps .25 .2 .15 .15 .1 .05 Motors .25 .2 .15 .15 LP1 LP2. brgs. .25 .20 .15 .15 .10 .05 Overhung type .50 .40 .25 .25 .10 .05 Electric motors .25 .20 .15 .15 Gear units Intermed. Freq. Parallel, 10 00–3000 Hz spec. purp. .25 .20 .15 .15 .15 .10 .05 Parallel, gen.