Vibration Mitigation on Engine Test Stands using Conventional Analysis Techniques

B. Swaminathan, Affiliated Construction Services (ACS), Madison, WI-53705, USA. Email: bswaminathan@acscm.com

NOMENCLATURE

FRF Frequency Response Function (FRF)

ODS FRF Operating Deflection Shape Frequency Response Function

EMA Experimental Modal Analysis

PTD Polyreference Time Domain

MAC Modal Assurance Criterion

RFP Rational Fraction Polynomial

CMIF Complex Mode Indicator Function

ABSTRACT

A single cylinder optical engine run by an electric motor at a university engine research center was observed to have high levels of vibrations at specific speeds. This paper discusses the case where incorrect isolation of the engine test stand combined with driveline component vibrations caused high vibration levels to be transmitted into the building structure through the floor. Modal and operational tests were carried out on the entire engine - motor assembly to detect the magnitude and the frequency range of the vibrations. Order analysis was performed on the operating data and correlated with the modal data to identify system resonances excited by multiple orders in the operating ranges of the engines. Specific issues such as backlash in the coupling were also detected from the test data and suitable actions were recommended.

Keywords: Engine Test Cells, Vibration Isolation, Rotating Machinery, Modal Testing, Operational Testing

1. INTRODUCTION

Structural dynamics analysis of engine test cells [1][2] have historically focused on development of the single most important piece of equipment within the confines of its 4 walls, the engine. While the internal combustion engine has hogged the limelight of the dynamics world with its complexities, a close second

would be the dynamometer that puts the engine through its test routines in order to prepare it for the real world. When these two complex pieces of equipment, along with their required sub-components, come together to work as a single system, it opens up an entirely new set of dynamic challenges. This paper discusses a case where the behavior of the engine test cell was studied as a whole and various test and analysis methods [3] used to reduce high vibration levels.

The cell that was tested consisted of a single cylinder optical engine used for research in flow visualization techniques at a university lab. The engine is motored by an electric motor from speeds between 600 rpm to 1800 rpm. Test data collected from the cell was analyzed and solutions were proposed to mitigate the excessive levels of vibration. A follow-up test was performed at the cell to establish the vibration levels upon implementation of the recommended vibration isolation.

2. TEST METHODOLOGY, INSTRUMENTATION & DATA ACQUISITION

The engine test cell discussed in this paper was subjected to two tests [4]:

1. An impact hammer modal test 2. An operating test

The two tests were done to understand the complete structural dynamics of the system. The impact test was used to obtain Frequency Response Functions (FRF) to extract modal parameters [5], while the operating test data was used to establish the vibration levels in various components of the system in real operating conditions. This data was further used for order analysis and order extraction [6].

The raw time histories were also processed to generate Operating Deflection Shape (ODS) FRFs [7]. ODS FRFs represent the true mode shape of vibration of the system in the actual running condition.

A 64 channel Bruel & Kjaer Pulse data acquisition system was used for the vibration testing. High sensitivity (1 V/g) tri-axial PCB accelerometers were used for the modal tests, while the operating measurements were made using high temperature resistant tri-axial PCB accelerometers with relatively lower sensitivity (10 mV/g).

X-Modal software from Structural Dynamics Research Lab (SDRL) at University of Cincinnati was used for modal parameter estimation using EMA based algorithms. The B&K Pulse Reflex software was used for Order analysis of operating data. ME Scope from Vibrant Technologies was used to analyze the operating data to obtain Operation Deflection Shapes (ODS) and for computing ODS FRFs.

3. TEST SETUP

Figure 1 shows the engine test stand assembly consisting of the electric motor mounted on an aluminium frame. The single cylinder engine is mounted at the far end of the frame. A set of shafts connect the electric motor to the engine. A Lovejoy coupling is used at the electric motor end as part of the driveline. The engine is driven by the motor during normal operation. A series of belts and pulleys are used to

transfer the torque from the motor to run a cam shaft. The entire test stand assembly is connected to the floor using rigid mounts.

Figure 1 Engine Test Stand Assembly

The entire test stand assembly is represented by a wire frame model as shown in Figure 2, with numbers indicating sensor locations. The driveshaft, pulleys, bearings, etc. along the driveline are represented by a line due to the small size of the components.

Figure 2 Wireframe Model of Engine Test Stand Assembly

40 response locations were picked for the impact hammer test, with 8 reference locations. A set of 20 accelerometers were used in 2 sets to cover the 40 response locations. For the operational testing, a set of 20 accelerometers were used to record time histories at 20 locations.

4. ANALYSIS AND RECOMMENDATIONS 4.1. BASELINE TEST RESULTS Data from the impact hammer modal test was analyzed using the Complex Mode Indicator Function (CMIF) [8] algorithm. The first rigid body and driveline modes that lie close to the operating range of the engine were selected from the CMIF plot shown in Figure 3.

Figure 3 CMIF Plot from Baseline Modal Test

The extracted modal parameters are listed in Table 1.

Frequency (Hz) RPM Mode Shape

19 1142 Lateral Swaying Mode

29 1738 Pitching Mode

33.7 2020 Driveline Bending Mode

38.5 2309 Yawing Mode

Table 1 Modal Parameters from Baseline Modal Test

Modal Assurance Criteria (MAC) values [4] were computed in order to validate the modal vectors. An Auto-MAC plot of the modal vectors obtained using CMIF algorithm is shown in Figure 4. Modal

parameters were also extracted using other algorithms such as Polyreference Time domain (PTD) [4] and Rational Fraction Polynomial (RFP) [9]. These modal vectors were compared to the CMIF results using Cross-MAC calculations in order to identify real modes of the structure.

Figure 4 Auto-MAC plot of Modes from Baseline Test

Figure 5 and Figure 6 show the mode shapes obtained from the impact testing. These primarily represent the rigid body modes and the first bending mode of the driveline.

Figure 5 Mode Shapes from Baseline Modal Test

Figure 6 Mode Shapes from Baseline Modal Test

Time histories were recorded while motoring the engine through its normal operating speeds. Speed sweep data was taken from idle (600 rpm) to rated speed (1800 rpm). Operating test data was processed using order analysis and order extraction techniques. Figure 7 shows order extracted data from several points on the engine and its components from a 600 to 1800 rpm speed sweep. FFT data was processed up to the 10th order and order extraction was done from half order up to the 6th order, in steps of half orders.

From the predominant 1st order peaks on the engine, it can be seen that the movement of the engine is primarily in the axial and lateral directions. Operating deflection shape (ODS) of the peak around 1600 rpm reveals a yawing motion of the engine and a 1st order bending of the driveline. This indicates a strong influence of both the 29 Hz and the 33.7 Hz modes during engine operation. ODS of the peak around 1050 rpm indicates a lateral swaying motion, implying that it is exciting the 19 Hz mode obtained from the impact hammer test.

Figure 7 Order Extracted Data from Baseline Operational Testing – Engine Points

Operating data from points on the test frame and the electric motor are shown in Figure 8. These graphs indicate similar peaks as the graphs from the engine. In addition to this, the electric motor also has a few 3rd order peaks in the operating range. Due to the limited measurement points on the motor, this data is not correlated to impact test results. Being higher order vibrations with lower velocities, these peaks are not expected to cause high levels of displacement.

Figure 8 Order Extracted Data from Baseline Operational Testing – Frame and Motor

4.2. RECOMMENDATIONS

Based on the modal and operational results, the following recommendations were provided to mitigate the 1st order vibrations in the test system:

• Vibration isolators can to be installed at the base of the test stand where it interfaces with the floor. The current setup consists of rigid mounts that do not offer any isolation. Properly sized isolators will reduce the vibrations transmitted to the floor.

• Resonances of the system are excited during engine operation. Some of these resonances are close to the steady state operating speeds of the engine. A finite element analysis of the entire test stand assembly can be used to make modifications to the frame and driveline such that the resonances are shifted away from these frequencies.

• The Lovejoy coupling used in the driveline has high relative displacement in the computed mode shapes. A new coupling with lower backlash would increase the stiffness of the joint and reduce vibration amplitude of the driveshaft.

The first recommendation was implemented on the test stand. Based on the modal test results, a low frequency Neoprene isolator with steel springs was chosen and installed. After installation, the modal impact and operational tests were repeated for validation purposes.

4.3. VALIDATION TEST RESULTS

Figure 9 shows CMIF plot from the validation test performed after installing the vibration isolators. Rigid body modes and 1st driveline bending mode were chosen from the CMIF plot for comparison purposes.

Figure 9 CMIF Plot from Validation Modal Test

Mode shapes and modal frequencies from the CMIF plot indicate that the rigid body modes have shifted lower in frequency when compared to the baseline results. This shift in modal frequencies is beneficial for this particular test cell as the engine is not capable of being motored below 500 rpm. The modal parameters from the validation test are listed Table 2.

Mode No. Frequency (Hz) RPM Mode Shape

1 1.5 90 Lateral Swaying Mode

2 3.4 204 Axial Mode

3 5.9 354 Yawing Mode

4 6.9 414 Rolling Mode

5 9.6 576 Vertical Bouncing Mode

6 13.9 834 Pitching Mode

7 15.4 924 2nd Rolling Mode

8 18.7 1122 Frame Bending Mode

9 32.3 1938 Driveline Bending Mode (cam shaft & motor out of phase)

10 41.9 2514 Driveline Bending Mode (cam shaft & motor in phase)

Table 2 Modal Parameters from Validation Modal Test

An Auto-MAC plot of the extracted modal vectors is shown in Figure 10. It can be observed that a few modes have a high MAC value between them. Upon inspection of the mode shapes, they are found to be unique. The high MAC values could be attributed to the fact that some of the frame components might not have had enough measurement locations to define them completely.

Figure 10 Auto-MAC Plot from Validation Testing

Figure 11, Figure 12 and Figure 13 show the animated mode shapes extracted from the validation test. A few additional response locations were added near the base of the test stand to capture the effect of the vibration isolators.

Figure 11 Rigid Body Mode Shapes from Validation Modal Test

Figure 12 Rigid Body Mode Shapes from Validation Modal Test

Figure 13 Driveline Modes from Validation Modal Test

Operating data was taken at the same points on the test stand assembly as earlier and processed to get order extracted plots. Figure 14 shows the extracted orders from points on the engine. Compared to the baseline tests, the 1st order vibration levels are observed to be lower. The peaks are shifted as expected, considering the fact that the rigid body modes have shifted to lower frequencies.

Figure 14 Order Extracted Data from Validation Operational Testing – Engine Points

Figure 15 shows the order extracted data from the points on the frame. The vibration levels are again observed to be lower compared to the baseline levels. The 3rd order peaks are still prominent in some of the components, although their magnitudes are comparable to the baseline test.

Figure 15 Order Extracted Data from Validation Operational Testing – Frame Points

5. SUMMARY AND SCOPE FOR FUTURE WORK

Structural dynamic behavior of an optical engine test cell was studied as a whole using conventional EMA techniques. A baseline set of measurements were taken to identify the source of excessive vibrations in the test stand assembly. Based on the results of the baseline modal and operating tests, structural modifications were recommended to mitigate the vibration levels. Vibration isolators were installed as a corrective action and the tests were repeated to validate the effectiveness of the isolators. Validation tests indicated that the installation of the isolators caused the rigid body modes to shift lower in frequency and out of the normal operating range of the engine. Vibration magnitude levels were also observed to be lower than earlier.

These tests and diagnosis indicate the benefits of studying an engine test cell as a whole in order to understand the effect of all the components interacting with each other. Vibration analysis of the complete test cell as a single system will lead to effective design that will reduce failures due to damaging levels of vibrations. This eventually means lower cost for replacing components and lesser downtime in the test cell.

Although the vibration isolators installed in the optical engine test cell are observed to be effective, there is scope for further reducing the overall vibration levels by implementing the recommendations discussed in Section 4.2

The test methodology used in this cell has been implemented in another test cell in the same facility that consists of a single cylinder diesel engine hooked to an A.C. Dynamometer. The engine is interfaced to the dynamometer through a driveline consisting of a shaft, dynamometer flywheel and adapter plates at both ends. This system is rated to operate from 600 rpm to 2100 rpm.

The test stand assembly is represented in the form of a wireframe diagram as shown in Figure 16.

Figure 16 Wireframe Model of Diesel Engine Test Stand Assembly

Diagnostic baseline testing was carried out to identify the source of excessive vibrations. Based on analysis of the data, suitable corrective actions were recommended and implemented. This test cell is awaiting validation testing to further verify the effectiveness of this approach to engine test cell design.

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