CHAPTER 7 FAULT DIAGNOSIS OF CENTRIFUGAL PUMP AND IMPLEMENTATION OF ACTIVELY TUNED DYNAMIC VIBRATION ABSORBER IN PIPING APPLICATION

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1 125 CHAPTER 7 FAULT DIAGNOSIS OF CENTRIFUGAL PUMP AND IMPLEMENTATION OF ACTIVELY TUNED DYNAMIC VIBRATION ABSORBER IN PIPING APPLICATION 7.1 INTRODUCTION Vibration due to defective parts in a pump can be an annoying problem resulting in unnecessary maintenance and can affect the pumping system performance and endurance. This chapter focuses on diagnosing of faults of centrifugal pump by vibration analysis. Also, to control the amplitude of vibration in a pipe line due to hydraulic pulsation frequency, this is the speed result of operating condition of the pump. The developed SMA based actively tuned dynamic vibration absorber was used to control the amplitude of vibration for varying excitation frequency. 7.2 CENTRIFUGAL PUMP Centrifugal pumps are one of the most important elements in almost all industries. The pumps are the key elements in food industry, waste water treatment plants, agriculture, oil and gas industry, paper and pulp industry, etc. Its purpose is to convert energy of a prime mover (an electric motor or turbine) first into velocity or kinetic energy and then into pressure energy of a fluid that is being pumped. The energy changes occur by virtue of two main parts of the pump, the impeller and the volute or diffuser. The impeller is the rotating part that converts driver energy into the kinetic energy. The volute or diffuser is the stationary part that converts the kinetic energy into pressure energy. The cut section model and fluid path through centrifugal pump is shown in Figure 7.1.

2 126 Figure 7.1 Fluid path through the centrifugal pump (Suhane A, 2012) 7.3 SIGNIFICANCE OF FAULT DIAGNOSIS USING VIBRATION ANALYSIS The most revealing information on the condition of rotating machinery is a vibration signature. Vibration parameters provide the needed frequencies due to the flow and recirculation. When analyzing the vibration data, an FFT vibration spectrum may be broken down into several frequency ranges to help to determine the machine problem. Vibrations externally measured on a pump have been used to monitor the operating condition of the pump and diagnose the fault, if there is any, without interfering with the normal operation. The most common method employed for examining mechanical vibration is spectral analysis. Condition monitoring and fault diagnostics are useful to ensure the safe running of machines. Vibration signals are often used for fault diagnosis in mechanical systems because they carry dynamic information from mechanical elements. These mechanical signals normally consist of a combination of the fundamental frequency with a narrowband frequency component and the harmonics. Most of these are related to the

3 127 revolutions of the rotating system since the energy of vibration is increased when a mechanical element is damaged or worn. Some of the conventional techniques used for fault signal diagnosis include power spectra in time domain or frequency domain. These can provide an effective technique for machinery diagnosis provided the signals are stationary. 7.4 CAUSES OF VIBRATION IN CENTRIFUGAL PUMP Vibration due to unbalance Dynamic imbalance in centrifugal impeller or shaft can cause heavy vibration and transmits it to piping which can be cured by balancing the fan with shaft on balancing machine. Simple unbalance, uncomplicated by other problems can be identified by the following characteristics (Brain PG 2011): a) Amplitude occurs at 1 X RPM of the shaft b) The radial vibration is reasonably uniform and not highly directional c) If the specific component such as impeller or fan is the source of unbalance, it will have high amplitude at 1 X RPM frequency Vibration due to misalignment Misalignment of direct coupled machines is the most common cause of machinery vibration. In spite of self-aligning bearings and flexible couplings, it is difficult to align two shafts and their bearings, which will cause vibration. Although machines may be well aligned initially, several factors can affect alignment, namely, operating temperature, setting up of the base or foundation and deterioration or shrinkage of grounding. Misalignment can be clearly identified from the following characteristics:

4 128 a) Predominantly occurs at 2 X RPM b) Amplitude is high in axial direction when compared to horizontal direction Vibration due to hydraulic pulsation The problems caused by hydraulic pulsation in pumps are easy to recognize because the resultant vibration will occur at frequency which is the product of number of impeller vanes and machine speed (rpm). The amplitude of vibration due to hydraulic pulsation in a pipe line is inevitable because of the working of the pump. It is not unusual to detect some vibration at the vane or blade passing frequency on nearly every pump. It would be impossible to build a machine where no hydrodynamic forces are present. However, when the amplitude of hydraulic pulsation is excessive, a problem is indicated. Vibration due to cavitation Pumps are designed to operate at certain flow conditions including suction and discharge pressures, flow rates, head pressures, product density or specific gravity, etc. If operated beyond or outside these designed parameters, a high amplitude of vibration generally results. Pumps that are forced to operate amount of fluid enters the pump is insufficient. This creates vacuum pockets in the fluid that are unstable and can even collapse or explode. Cavitation can be identified by the following characteristics: a) Cavitation occurs between CPM ( Hz) and CPM (2500 Hz) b) Vibration can be detected at any location in the pump

5 129 c) A hydraulic problem and it may be due to the design of vanes d) Causes haystack of vibration Vibration due to bearing defects When a rolling-element bearing develops flaws on the raceways and/ or on rolling elements, there are actually a number of vibration frequency characteristics that can result, depending on the extent of deterioration. Thus, identifying these characteristic frequencies can not only help to verify that a bearing is definitely failing, but it can also give some indication on the extent of deterioration. Bearing defects can be identified by the following characteristics: a) Defects on bearing occurs between CPM ( Hz) to CPM (2500 Hz) b) The vibration can only be detected at the place of bearing c) The haystack of vibration will increase and spread out as the day progresses Apart from the above stated reasons, piping system might experience vibrations due to improper supports, fittings and water hammer. In case of gas pipe lines, the vibration can also come from the pulsations generated by reciprocating compressors. 7.5 INFLUENCING PARTS OF PIPING SYSTEM working of the pump. The following parts are subjected to vibration and defect due to the Pipe All piping supports Hangers Snubbers

6 130 Pipe to pipe interfaces Machinery or devices attached to the pipe All these items can influence the pipe vibration patterns. The vibrations produced in the pipelines contain various risks concerned with the industry as well as the domestic applications. A pipe will not vibrate if it is prevented from moving. However, this does not necessarily help the piping system design from the standpoint of its ability to absorb differential thermal expansion. Therefore, when addressing a vibration problem, the flexibility design of the piping system must also be considered. Restraints that are added to reduce vibration must not increase the pipe thermal expansion stresses or end-point reaction loads to unacceptable levels. 7.6 SPECIFICATION OF PUMP It is necessary to obtain amplitude vs frequency spectrums or FFTs absolutely no value to the vibration analyst unless some specific details about the machine are known. Specific machinery problems are identified by relating their vibration frequencies to the rotating speed (RPM) of the machine components, along with other machine features such as the number of teeth on gears, the number of blades on a fan etc., are shown in Table 7.1. Table 7.1 Specification of centrifugal pump 1. Rotating speed of shaft RPM 2. Type of bearings - Roller bearing 3. No of rolling element in each bearing No of impeller vanes Head range feet 6. Output range lpm

7 EXPERIMENTAL PROCEDURE FOR FAULT DIAGNOSIS Experimental setup used in this study for fault diagnosis of a centrifugal pump is shown in Figure 7.2 (a). Centrifugal pump is rigidly fixed to the foundation, so that the vibration caused due to the looseness with the foundation could be avoided in experimental results. Accelerometers are mounted at the inboard bearing of centrifugal pump as shown in Figure 7.2 (b). (a) Major parts (b) Location of accelerometer Figure 7.2 Experimental setup for fault diagnosis of centrifugal pump Experimental setup is made to run as per the testing condition standards. By changing the accelerometer orientation, data was recorded in other two axes at the same bearing location. The vibratory forces generated by the rotating components of a machine are passed through the bearings. Vibration readings for both detection and analysis are taken on the bearings whenever possible. Ideally, vibration readings taken in horizontal and vertical directions are taken directly on or as close as possible to the bearings with the accelerometer pointing towards the centerline of the shaft. Axial vibration readings are taken on the bearing as close to the shaft as possible. Adhesive mounting is used to mount the

8 132 accelerometer over the bearing which provides frequency response up to CPM (9000 Hz) that is sufficient for this experimentation. The things that must be taken care while adhesive mounting is machine surface must be flat, smooth and clean to ensure secure bonding. The layer of adhesive should be kept as thin as possible to provide maximum frequency response. Figure 7.3 shows the Lab VIEW block diagram for this experimentation. NI USB DAQ 9461 with 4 input channels and delta type DAQ is used to connect the accelerometer with Lab VIEW software. Initially, the measurement was taken to check for background noise, which clearly indicates that there is no significant level of surrounding as shown in Figure 7.4. Figure 7.3 Lab VIEW block diagram

9 133 Figure 7.4 Lab VIEW measurement for surrounding noise level Also, the frequency spectrum of the healthier pump was recorded, so this Figure 7.5. Then, the frequency spectrum of defected pumps was recorded. From the data collected, the defects are identified through vibration analysis by finding the difference in amplitudes. Also, according to pump standards the amplitude of vibration is within 0.15 inches which is acceptable due to any cause of vibration (Brain PG 2011). Figure 7.5 Vibration level of a good pump

10 134 Few pumps were taken from pump industry which has customer complaints for vibration analysis in order to find out the causes for complaints. From Figure 7.6 (a) & (b), it can be clearly noticed that the peak amplitude of vibration occurs at 50 Hz, which is clearly a characteristic of an unbalanced impeller. The magnitude of unbalance is little higher in case of second pump when compared to first one. But in both cases, at frequency of 50 Hz, amplitude of vibration is higher than the permissible value which has to be addressed. Vibration was measured both in axial and radial direction. (a) Radially (b) Axially Figure 7.6 Vibration measured for an unbalanced pump

11 135 The vibration levels measured in the radial and axial axes of the pump are shown in Figure 7.7 (a) & (b). From the Figure 7.7 (b), it can be clearly seen that the axially measured amplitude vibration of a pump is 700% greater than that of radially measured vibration. Both are located at 100 Hz. Also, since it has different magnitudes in different axes, it is found to be directional. It shows that the amplitude at 100 Hz is clearly the characteristic of misaligned shaft pump. Further, it can be understood that the vibration is at risk because of the misalignment which has to be addressed. (a) Radially (b) Axially Figure 7.7 Vibration measured for misalignment in a pump

12 136 Similarly a number of pumps were taken for experimentation to study the effect of cavitation and hydraulic pulsation in various pumps. Common pump defects and its corresponding frequency is shown in Figure 7.8 which will be helpful for further fault detection in defected or normal pumps. Figure 7.8 Vibration characteristics of centrifugal pump with defects 7.8 VIBRATION CONTROL IN A PIPE LINE USING SMA BASED ATDVA Vibration in the pipelines may be caused by various factors as discussed above. However, the vibration due to the hydraulic pulsation is caused when the impeller in a pump continuously transmits tiny pockets of water into the pipeline. The hydraulic pulsation is calculated as given in Equation 7.1. Hydraulic pulsation frequency = NZ (7.1) where N is the speed of the shaft and Z is the number of vanes on the impeller. An attempt is made to develop an absorber system to control the amplitude of vibration caused by the frequency which is the result of impeller rotation. Hydraulic pulsation frequency caused by the speed is a result of operating

13 137 condition of the pump. The variation in the speed due to voltage fluctuation, slippage in rotor and variation in load conditions that leads to increase in vibration in the pipe line. This may increase the chances of causing damages to pipeline, the seals and joints which results in leaks. The leakage of the toxic chemical or gases may pollute the atmosphere or may even affect the surroundings. Further, the pipe line produces sound and causes loss of the fluid. These undesirable effects are produced due to the variation in hydraulic pulsation frequency is addressed by the development of SMA based actively tuned vibration absorber. Figures 7.9 & 7.10 show the centrifugal pump apparatus and the experimental setup including the interfacing the system with LabVIEW. Figure 7.9 Experimental setup Figure 7.10 Experimental setup with Lab VIEW

14 138 Based on the centrifugal pump selected for vibration analysis, the hydraulic pulsation frequency was calculated by using Equation 7.1. Speed of the motor (N) = 2880 rpm Total no of vanes (Z) =7 Excitation frequency of the pump = 2880 x 7 = rpm (hydraulic pulsation frequency) =20160/60 Excitation frequency of the pump = 336 Hz When the vibration is measured on the pipe line, it is evident from the Figure 7.11 that the peak occurs at nearly 336Hz and it is the frequency which has the highest amplitude. This amplitude of vibration is due to hydraulic pulsation (Excitation Frequency) which was considered for active vibration control. But in the experimental results, it was found that the peak frequency is varying between 336 Hz and 339 Hz. This may be due to the voltage fluctuations, varying load conditions and slippage in rotor. For the varying frequencies from 336 Hz to 339 Hz, the conventional absorber has to be redesigned which is cumbersome. Also, the SMA spring available may not offer the stiffness for the frequency around 336 Hz, Hence, it is decided to develop a combined absorber system with conventional spring and SMA spring so the equivalent stiffness was considered for calculation of excitation frequency.

15 139 Figure 7.11 Amplitude of vibration due to hydraulic pulsation Development of conventional absorber In this study, the conventional absorber is developed for the excitation frequency of 336 Hz. The variation in excitation frequency due to the impeller rotation is taken care by the SMA spring. Approximately a change speed of 180 rpm (3 Hz) is accounted for reduction in amplitude of vibration. Angular velocity x 3.14 x f where f = 336 Hz = 336 x 2 x 3.14 = rad/s modulus E = 2.1x10 5 N/mm 2 (spring steel) =0.3 Wire diameter d= 1.8 mm Mean diameter D=3.72mm No of turns N=9 Stiffness of the conventional spring K = N/mm

16 Development of an absorber with conventional SMA spring The stiffness offered by the SMA spring is not sufficient for controlling this very high excitation frequency. Hence, to improve the stiffness, a parallel system of springs has been used which consists of both the conventional and SMA springs as shown in Figure Figure 7.12 DVA with conventional spring and SMA spring K 1 = stiffness of the conventional spring. K 2 = stiffness of the SMA spring. Equivalent stiffness (K e ) = K 1 + K 2 In martensite state, Equivalent stiffness of the parallel system = K 1 + K 2 K 1 = N/m K 2 = N/m in martensite state of SMA spring K e = N/m and the corresponding frequency is calculated as 2 x x where stiffness K e = N/m and suspended mass m=50g rad/s. Frequency f = Hz.

17 141 In austenite state, K 1 = N/m K 2 = N/m in austenite state for SMA spring K e = N/m and the corresponding controlling frequency is calculated where K e = N/m and for the same suspended mass of 50g 2 = /0.050 = rad/s = 2 x 3.14 x f f= 339 Hz Hence, the absorber was designed for controlling the frequency in the range of 336 to 339 Hz. 7.9 EXPERIMENTATION USING SMA BASED ATDVA The excitation frequency of the pump is calculated as 336 Hz. But in the experimental results, it was found to be varying between 336 and 339 Hz. For the varying frequencies of 336 to 339 Hz, an SMA based actively tuned absorber system has been developed to take care of this 3Hz change in excitation frequency due to hydraulic pulsation. Figure 7.13 shows that the peak occurs at 337 Hz with amplitude of 2.3mm. But the theoretical calculation shows that the excitation frequency of the pump occurs at 336 Hz according to the Equation 7.1. This variation in frequency may be due to voltage fluctuation, varying load conditions and slippage in rotor. An absorber designed with the help of SMA spring along with conventional spring was used to control the amplitude of vibration due to the

18 142 change of 3 Hz in excitation frequency. The amplitude of reduction in vibration at 337 Hz is shown in Figure From this, it is evident that the amplitude of vibration reduces from 2.3 mm to 0.8 mm which accounts for around % reduction of amplitude due to the SMAs stiffness changing ability. Also, the experiments were carried out by intentionally varying the excitation frequency within the permitted level of change in stiffness of SMA spring. Figure 7.13 Amplitude of vibration without absorber at 337 Hz Figure 7.14 Amplitude of vibration with absorber at 337 Hz When the excitation frequency is varied to 338 Hz, the amplitude of vibration is measured on the pipeline. Figure 7.15 shows that the peak amplitude occurs at 338 Hz, owing to the variation in the excitation frequency. The SMA spring will be supplied with required current by the control system,

19 143 in order to produce the desired stiffness to reduce the amplitude of vibration, which is the result of change in frequency from 337 to 338 Hz. The amplitude of vibration is reduced from 2.1 mm to 0.8 mm which is shown in Figure This results in 58.09% reduction in vibration of the pipe line. Figure 7.15 Amplitude of vibration without absorber at 338 Hz Figure 7.16 Amplitude of vibration with absorber at 338 Hz When the excitation frequency is increased further to 339 Hz, it tends to develop peak amplitude in pipe line at 339 Hz. The amplitude is found to be 2.25 mm at 339 Hz as shown in Figure 7.17.

20 144 Figure 7.17 Amplitude of vibration without absorber at 339 Hz This leads to 62.2% reduction in amplitude of vibration at the frequency of 339 Hz which is depicted in Figure If the excitation frequency is reduced from 336 to 334 Hz. It is not possible to use the combined DVA with conventional and SMA, because the conventional DVA was designed to take care of the frequency of 336 Hz. Figure 7.18 Amplitude of vibration with absorber at 339 Hz The SMA spring is attached with conventional spring will increase the stiffness leads to account only increase in excitation frequency for vibration control. This problem can be addressed when only SMA springs of high stiffness ranges can be combined parallel nature or co-axial in nature results in more variation in excitation frequency range both below and above the frequency of 336 Hz.

21 145 Figure 7.19 shows the reduction in amplitude of vibration in the pipe line for the frequencies of 336 to 339 Hz. Table 7.2 shows the reduction in amplitude of vibration with percentage. 2.5 Amplitude (mm) Frequency (Hz) Without absorber With SMA spring (mm) Figure 7.19 Reduction in amplitude of vibration for the frequency range of 336 to 339Hz Table 7.2 Comparison of percentage reduction in amplitude of vibration for 336 to 339 Hz of excitation Frequency (Hz ) Without absorber (mm ) With SMA spring (mm) Percentage of reduction (% ) CONCLUDING REMARKS Thus, the vibration due to different defects in centrifugal pumps has been analyzed and the frequency at which the defects are happening has been found out by the experimentation procedure. These frequencies would be useful for the fault diagnosis of the centrifugal pumps without dismantling the assembly. Experiments were further carried out with the help of SMA springs

22 146 on centrifugal pump apparatus. In order to demonstrate the concept of actively tuned dynamic absorber for varying excitation frequencies due to hydraulic pulsation results in more amplitude of vibration in the pipe lines connected with centrifugal pump. The change in excitation frequency in the inclining trend was addressed with the help of ATDVA. The reduction in amplitude of vibration around 50-65% was attained with the help of developed SMA based actively tuned dynamic vibration absorber. The variation in excitation in decreasing trend, which can be addressed only with SMA springs of high stiffness ranges can be combined parallel nature or co-axial in nature resulting in more variation in excitation frequency range in the decreasing trend.