Multidisciplinary Senior Design Conference

Kate Gleason College of Engineering Rochester Institute of Technology

Rochester, New York 14623

Copyright © 2012 Rochester Institute of Technology

Project Number: P13471

UL VIBRATION TEST APPARATUS

Sean Coots Mechanical Engineering

Walter Bergstrom Mechanical Engineering

Spencer Crandell

Mechanical Engineering Mark Ellison

Mechanical Engineering

Abstract A vibration testing system was designed, fabricated and tested at the Rochester Institute of Technology through the Winter and Spring quarters of 2012-2013. Sponsored by Cooper Crouse-Hinds (now Eaton Corporation), the system was designed to follow UL 844 Testing Standard, which details the vibration testing procedure for luminaires to be used in hazardous locations. This project is to be completed over multiple Senior Design phases. This particular phase focused on the design and fabrication of the mechanical portion of the system; that is, the mechanism that translates rotary motion to linear, the eccentricity adjustment and locking mechanism, and the linear slider that provides the connection between the suspended luminaire and the motion mechanism. Testing was completed on the system to verify that the displacement (or stroke) achieved by the reciprocating mechanism followed UL 844.

Introduction Project Scope This project is to be completed over multiple Senior Design phases. The final product will consist of a mechanical system with LabView communication that will supply sufficient displacement to a luminaire at the required frequency. This particular phase includes the full design of the mechanical system; that is, the design-related deliverables of this project include a full CAD model of the displacement mechanism, the connection to the vertical conduit, motor selection, and a preliminary frame design, all accompanied by detailed drawings. The primary focus for this Senior Design group was to design, fabricate and assemble the displacement mechanism. Testing was completed to verify that the displacement achieved through the designed mechanism meets the UL Standard.

Summary of Current Apparatus The current vibration testing apparatus in use at Cooper Crause-Hinds is in need of replacement. The system runs in a separate area from the rest of the laboratory, which includes an isolated floor to prevent vibrations affecting the other areas of the facility. The current system is largely exposed, however; lubricant is uncontained and mechanical components are open to contaminants. The frame design makes it difficult for a single technician to prepare the system for testing. A single speed motor is used in unison with a speed reducer, which power an eccentric shaft to supply the required displacement. The system does not have the ability to modify the stroke associated with the eccentric. Many of the components have not been replaced in decades, and the design includes a number of custom components that are not documented or described. It is likely that the system does not account for part wear and tolerance stack-up. Additionally, it is difficult to verify with high accuracy that the frequency and displacement supplied by this apparatus truly follows the UL Standard.

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Summary of UL 844 Testing Standard Section 33 of UL 844 details the testing procedure and setup for luminaires to be used in hazardous locations [1]. The luminaire is to be attached via a threaded connection to a 26-1/2 inch (673 mm) vertical length of rigid conduit. The vertical conduit is to be screwed into a rigidly mounted frame. A horizontal rod is to be rigidly connected to the vertical conduit at 4 inches (102 mm) above the lower end of the conduit. A mechanism must drive the conduit-rod system through a cyclic displacement of 1/16 inch (1.6 mm), or 1/32 inch (0.8 mm) in each direction from the vertical position, at 2000 cycles per minute. The system setup must allow for stanchion, yoke, and trunnion mounting styles. This test is to be completed over a period of 35 hours. Proper completion of the test will show no damage to the joints or other explosion-proof or dust-ignition-proof features of the luminaire subjected to testing. There is no indication that this standard will change in the foreseeable future, but the mechanism design allows for minor adjustment of the key parameters: frequency and displacement.

Figure 1: UL 844 Supplemental Diagram

Methodology Theory To calculate the required power to deflect the vertical conduit at the preferred frequency, the analysis was split into three segments: the calculation of the acceleration of the system, the calculation of the force induced on the displacement mechanism due to the inertia of the system, and the calculation of the torque due to the inertial force.

The acceleration of the system was completed using two distinct methods: a dynamic analysis of a crankshaft mechanism and a frequency analysis [2]. Considering the crankshaft as two rigid links as shown in Fig. 2, the following acceleration equations were applied for aA and aB:

The acceleration at A was an intermediate calculation, as the acceleration at B translates directly to the vertical conduit. The maximum acceleration at B was found to be 57.4 ft/s (17.5 m/s).

Figure 2: Crankshaft Linkage

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Since the acceleration at B was known, it was then necessary to find the acceleration at the location of the center of mass of the luminaire, which was assumed to be a maximum of 100 lbf (444.8 N). For this analysis, a rigid conduit was considered in pendulum motion with small angle approximations applied, as shown in Fig. 3. A similar dynamic analysis was applied to solve for the angular acceleration of the conduit:

The maximum angular acceleration calculated was found to be equal to 30.4 rad/s

2.

Figure 3: Vertical Conduit

The alternate method involved the use of the simple harmonic motion equation. The derivative of the equation provided the velocity, and the second derivative provided the acceleration. This method is detailed directly below, where A is the amplitude, or stroke, of the displacement mechanism and ω is the desired frequency:

This method provides identical answers to the dynamic analysis for maximum velocity. The reason for this is simple: since the crank arm (length ‘r’ Fig. 2) is significantly smaller in magnitude than the link (length ‘l’ Fig. 2), the crankshaft system essentially acts in simple harmonic motion. As the crank arm and link approach the same length, however, the acceleration no longer follows a harmonic motion pattern and becomes less uniform.

The force calculation is directly dependent on both the inertia of the conduit-luminaire system and, since it is assumed to be rigidly connected, the pure bending of the conduit. Eqn. 7 represents the maximum-acceleration inertial calculation, while Eqn. 8 represents the pure bending of the conduit [3]. The moment of inertia, Itotal, for the system is calculated by assuming two cylinders to model the conduit and the luminaire. Rconduit refers to the length from the rigid connection C to the mechanism connection B, E is the Modulus of Elasticity, and δ is the maximum deflection of the beam at point B.

The force due to inertia is 280 lbf (1245.5 N) and the force due to bending is 80 lbf (355.9 N). Applying superposition to these forces provides a total required force of 360 lbf (1601.4). It was further assumed to be 400 lbf (1779.3 N) due to the fact that there may be deflection of the frame to some degree at such a frequency. A factor of safety of 2.0 was applied at this point, making the force due to the conduit-luminaire system (which will be referred to as FB from this point forward) 800 lbf (3558.6 N).

A modified single-cylinder engine dynamic analysis was used to trace FB through the crankshaft mechanism [4]. This analysis includes the inertial effects of the connecting rod, crankshaft and linear slider, ultimately providing the bearing forces on the crankpin. The maximum x-component of this force was found to be 1138 lbf (5062.1 N) and the maximum y-component to be 12 lbf (53.4 N). The almost negligible y-component is due to the very small crank arm length compared with the connecting rod. From this point forward, a force of 1200 lbf (5337.9 N), FS, acting on the driving shaft was assumed.

The torque required for the system at steady state (2000 cycles per minute) was calculated in Eqn. 9, while Eqn. 10 details the required torque during start-up. Tref represents the reflected torque on the shaft due to FS, Isys refers to the total moment of inertia of the motor-shaft-crankshaft system, and α refers to the start-up acceleration of the motor until it reaches a steady-state speed.

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Treq during steady-state was found to be 1.1 lbf-ft (1.5 N-m), which corresponds to 0.4 HP. Assuming a start-up time between 50 and 60 seconds for the motor to reach steady-state, Treq was found to be ~3.0 lbf-ft (4.1 N-m). This start-up torque will be discussed further in the motor selection segment of Detailed Design.

Detailed Design The final design can be split into three main subsystems: the adjustment mechanism, which also serves as the crankpin, the linear slider, which includes the connection between the displacement mechanism and the vertical conduit, and the frame. Additionally, the motor selection, lubrication, and hardware used in the design will be discussed.

Due to the small displacement described in UL 844, it was decided that the displacement mechanism should have a built-in adjustment capability to allow for minor modifications to the stroke magnitude. This necessitated a robust method to not only adjust at the order of 1/64 in (0.397 mm) but to also securely clamp the crankpin at the desired eccentricity. The design therefore employs a friction locking mechanism that is allowed freedom of motion radially from the center of a steel disk. This design was taken from [5] and modified to provide greater clamping force for this application. The disk is directly connected to the driving shaft for the system (Fig. 4). The locking mechanism consists of two main components: a base that is rigidly connected to the rotating disk and an upper slider which controls the eccentricity (Fig. 5). The base employs a T-slot with two slider blocks. Bolts thread through the upper slider and into the T-slot slider blocks, effectively pulling the components together. Nord-Lock lock washers are used to lock the bolts in place so that will not loosen during operation. The friction surfaces exist between the T-slot slider blocks and the base and the 45° angled surfaces between the base and the upper slider.

The crankpin threads into the upper slider (and is locked in with a Nord-Lock washer from the underside) and provides the eccentric mechanism for the displacement system. To adjust the eccentricity of the mechanism, two features were incorporated into the design. First, dowel pins located in the disk and upper slider mate at the exact center of the disk; this orientation provides zero eccentricity for the crankpin system. A precise shim can then be inserted between the dowels, providing the desired stroke. The second feature is a die spring-set screw combination that allows for very fine adjustment. The springs provide a force on one side of the upper slider, forcing it against the set screw. The set screw can then be turned to adjust to the desired displacement. Figure 6 provides isometric views of the base and upper slider, while Fig. 7 provides an isometric view of the entire mechanism.

Figure 4: Rotating Disk and Drive Shaft

Figure 5: Adjustment Mechanism Section View

Figure 6: Adjustment Mechanism Exploded View

Figure 7: Adjustment Mechanism Isometric View

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The function of the linear slider subsystem (shown in Fig. 8) is three-fold; the primary function is to provide a means to translate the rotational motion into linear motion through a connecting rod. The slider box is constrained to move only in the plane of desired displacement. Linear pillow block bearings are attached to the sides of the slider system and are guided by steel rails. The rails are rigidly connected to the base plate for the entire vibration system. The connecting rod, shown in Fig. 9, is a solid steel component. Simple ball bearings are press-fit into the connecting rod and then attached to the crankpin on one side and a similar pin on the other, which is rigidly connected to the slider box. This provides a connection between the rotational system and the linear.

Figure 8: Linear Slider Subassembly

Figure 9: Connecting Rod

The second function of the linear slider is to provide a means to translate the linear motion to the vertical conduit and therefore to the luminaire. The extension rod (Fig. 10) connects to a threaded rod that can be attached to a collar that clamps on the vertical conduit. The extension rod is attached to the linear slider by two shaft clamps that provide sufficient force to connect the rod rigidly to the slider. This rigid connection constrains the conduit to move at the same displacement and frequency as the linear slider. The third and final function of the slider subassembly is to enclose the mechanical system. The slider in the current apparatus is separate from the eccentric mechanism. It was decided in the course of this design that it would be preferred from a safety standpoint to combine the eccentric mechanism and linear slider into an enclosed system. The adjustment mechanism and associated subcomponents (rotating disk, flange, hardware) adds significant mass to the system; enclosing these fast-moving components in a steel box provides a significant safety advantage over separating the linear slider and eccentric system. A clear polycarbonate top and backing complete the enclosure. These polycarbonate shields can quickly be removed from the slider assembly to allow for maintenance on the enclosed subsystems. The rearmost supports for the linear box were designed to allow for accessibility to the inner components.

Figure 10: Extension Rod Detail

Figure 11: Linear Slider Frame

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Figure 11 provides a view of the slider box frame, including hardware connection points. The front plate does not carry the load from the connecting rod; the horizontal plate centered in the slider carries this load. The support ribs below the horizontal plate provide the connection to the outer box frame; the horizontal plate is bolted directly to these support ribs, not directly to the outer frame. The connection points and associated hardware in the lower support ribs are loaded in shear but have been found to provide sufficient safety factors for operation.

The general frame for this system is a recommended design. Fabricating the frame falls outside of the scope of this phase of the project; a future Senior Design team will construct and test the frame. A proposed design has been completed, however, and may be modified during future work on this project. The current apparatus in use at Cooper employs a frame with a cantilever. The vertical conduit is threaded up from below into the cantilever. It was decided that a more robust design could be created if a cantilever design was not employed. Rather, three sets of vertical posts support a horizontal beam that provides the foundation to the frame. One set of vertical supports is located at the point of connection between the frame and vertical conduit; this placement ensures maximum support at the critical connection point in the displacement mechanism. The other two sets of vertical supports provide the means to fasten the base plate of the displacement mechanism to the frame. Damping pads are located on each foot of the design as well as between connection surfaces. These pads should prevent excess vibration of the frame due to the displacement of the luminaire. The frame is to be bolted to the concrete floor in the testing room. Figure 12 displays the proposed frame design.

Figure 12: Proposed Frame Design

Figure 13: Flange Detail

An additional design feature is detailed in Fig. 13. The vertical conduit is threaded into a large flange at the point of rigid connection to the frame. The luminaire can then be threaded to the lower end of the conduit. This entire system can then be inserted into the groove on the upper horizontal beam component of the frame and bolted to the frame through five fasteners. This feature allows for more efficient test setup and ensures that a single operator can prepare a luminaire for testing.

This purchase and implementation of a motor and associated variable-speed drive do not fall within the scope of this project. A proposed motor is included with this design, however. As shown through Eqn. 9, the steady-state torque required for this system corresponds to 0.41 HP. Therefore a 1 HP motor is recommended. This poses an issue with start-up, however, as the required torque is dependent upon the time the motor takes to reach steady-state. Assuming 50-60 seconds to reach steady-state, Eqn. 10 provides a required torque value of ~3 lbf-ft (4.1 N-m). Referencing Reliance Electric’s AC Motor Performance Curves [6], the torque that can be provided by a general purpose 1 HP AC motor dips during start-up to ~3.8 lbf-ft (5.2 N-m) at 400 RPM but then increases to 5.7 lbf-ft (7.7 N-m) at the operating frequency for this system at 2000 RPM. Since the calculated required torque (3 lbf-ft) is below the provided torque from the motor (minimum 3.8 lbf-ft), a 1 HP AC motor is recommended for use in this system.

Lubrication in this system is vital for proper operation. Bearings are used in a variety of applications; ball bearings provide freedom of motion to the crankshaft and drive shaft and linear bearings allow unidirectional motion of the slider system. It was decided during the course of the design to use ball bearings in the connecting rod as opposed to

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a journal-sleeve connection because it is much more difficult to contain lubricant in a journal-sleeve bearing, especially in the orientation required in this design. Heavy-duty flange bearings provide support for the drive shaft. These bearings were purchased with grease fittings to ensure continual and effective lubrication. The linear pillow block bearings were also purchased with grease fittings. All bearings are fully sealed to contain lubricant.

The hardware selection for this design is important from a strength perspective. Many of the screws used in the slider mechanism are load bearing in shear. All fasteners used in this design are comparable to Grade 8 steel and have a black oxide finish for lubricity and mild rust resistance. The locking washers chosen to be used in the adjustment mechanism are Nord-Lock lock washers. These lock washers are used in four places; two are used on the fasteners that clamp the upper slider component of the adjustment mechanism to the base and two are used to prevent loosening of the connecting rod pins. Nord-Lock lock washers are considered the highest performing lock washers that also are reusable [7, 8]. They feature serrated teeth between washers that lock together when tightened to the prescribed torque.

Testing Process

Correct displacement of this mechanical system is one of two requirements detailed in UL 844. Since the purchase and implementation of a motor is not within the scope of the project, frequency could not be verified for the system. Therefore a robust method to verify the displacement (or stroke) of the crankshaft mechanism is required. The proposed test method employs a standalone aluminum frame that can be rigidly supported. A vertical conduit (schedule 80 pipe) is to be rigidly connected to this frame at the prescribed distance specified in UL 844. The mechanism system, which is mounted on a table, can then be connected to the conduit through a collar.

To test the torque that is required to provide the correct displacement, the adjustment mechanism must first be locked in the eccentric location to provide the desired stroke. Once in position, a fixture is to be attached to the bottom of the drive shaft. This fixture has a fitting that allows a torque wrench to supply a known torque to the system. The displacement of the vertical conduit is measured at the point of connection to the crankshaft mechanism with a dial gauge to verify that it meets the standard. The torque required to achieve this displacement must be less than the calculated steady-state torque for the system to verify motor selection. Figure 14 details the test setup and final assembly of the system.

Figure 14: Test Setup and Final Assembly

Results and Discussion Mechanical analysis was completed on the subsystems of this design and is reflected in the final assembly. Budynas and Nisbett provide the theoretical methods used to validate the design in [4]. The displacement of the system was tested to verify that the crankshaft mechanism operates per design. First, the zero-displacement mark of the system was measured. The adjustment mechanism has two dowels that mate at the exact center of the rotating disk, which aligns the crankpin and shaft and therefore provides zero displacement of the slider box. A dial indicator was placed in two locations for measurement: one on the slider box itself and on the vertical conduit. Both measurements provided identical displacements of 0 in (0 m). This validates that the crankshaft system’s center point is truly at the center of the shaft.

The second measurement required the use of gauge blocks to verify displacement of the system at eccentric values greater than 0 in (0 m). Measurements were taken at three eccentric values: 0.250 in (6.35 mm), 0.100 in (2.54 mm), and the UL Standard required displacement of 1/64 in (0.397 mm). At 0.250 in, the dial indicator showed a deflection within ± 0.010 in (0.254 mm) of the desired value. It is postulated that there are uncertainties due to the fact that the dowels used to provide the eccentric deflection were not pressed into the adjustment mechanism and rotating disk. Tapered dowels are to be implemented in the next phase of this project. At the lower deflections, the

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dial gauge indicated significant discrepancies. At 0.100 in, the measured deflection was between 0.045 in (1.143 mm) and 0.055 in (1.397 mm), which is nearly half the desired deflection. At 1/64 in (0.015625 in), the measured deflection was between 0.003 in (0.076 mm) and 0.006 in (0.152 mm), which is considerably lower. It is thought that the dial indicator used in this application was not precise enough to measure the lower displacements accurately. It is recommended that a more exact linear measuring device is purchased and the tests repeated for validation.

As detailed previously, the required steady-state torque required to move the system is 1.07 lbf-ft (1.45 N-m). The torque wrench used during testing was rated for torques considerably higher than what this system will see; therefore the accuracy at which measurements could be made contains a high degree of uncertainty. It was, however, verified that the required steady-state torque for this system is below 2 lbf-ft (2.7 N-m). That is the limit that the torque wrench output could be measured accurately. That does, however, show that the torque seen in this system is similar to the estimated steady-state value. It is recommended that a more accurate torque wrench is applied to this system and the test repeated to verify the torque value.

Conclusions and Recommendations The mechanical system outlined in this document is a completed assembly and has been proven through testing to provide the correct displacement according to UL 844. The required torque to provide the desired displacement was found to be below 2 lbf-ft (2.7 N-m) during testing of the design. The torque wrench used in testing was not ideal; that is, it was not precise enough at lower torque values to provide an exact torque value, though it is in the range of the required steady-state torque. It is recommended that both a more precise torque wrench and a more exact linear indicator are purchased to verify these results. This design meets the desired customer needs outlined by the customer at the start of this Senior Design phase. Most importantly, the system provides a method to fine-tune the stroke of the crankshaft to meet the UL Standard and a robust locking mechanism to provide the desired displacement over the course of a 35 hour test. The system is entirely enclosed within the linear slider to prevent contamination of components and as a safety feature.

A number of recommendations have been outlined in the Detailed Design portion of this document, but they will be reiterated. A 1 HP AC Motor is recommended with a start-up time of at least 50 seconds. The recommended frame design is detailed in this document. It is also recommended that the frame design be further analyzed to verify that its resonant frequency (the first mode) falls above the operating frequency for the motor. To achieve this, the frame can be lengthened along the plane of motion. Customer feedback has been positive, though it was recommended that the next Senior Design group incorporate the following features: (1) a variable-frequency drive that operates within an open feedback loop with a LabView interface for in-test verification (and adjustment) of vibration frequency, (2) a means to isolate the motor from directly driving the drive shaft to the system due to the fluctuating loads of the system, (3) a shield to protect the motor from lubricant and other contaminants due to the system above, and (4) an alternative method to anchor the crankpin to the adjustment mechanism other than through interference fits.

References [1] Underwriters Laboratory, 2006, Electric Lighting Fixtures for Use in Hazardous (Classified) Locations,” 844 [2] Wilson, C. E., and Sadler, J. P., 2003, Kinematics and Dynamics of Machinery, Pearson Education Inc., Upper

Saddle River, NJ, Chap. 2-4. [3] Rahnejat, H., Ed., Boedo, S., 2010, Tribology and Dynamics of Engine and Powertrain, Woodhead Publishing,

Philadelphia, PA, Chap. 19. [4] Budynas, R. G., Nisbett, J. K., 2010, Shigley’s Mechanical Engineering Design, McGraw-Hill, New York, NY [5] Jensen, P. W., 1991, Classical and Modern Mechanisms for Engineers and Inventors, Marcel Dekker, Inc., New

York, NY [6] Reliance Electric, 1992, “Reliance Electric A-C Motor Performance Curves,” from

http://www.reliance.com/cgi-bin/mtrquery.pl [7] Nord-Lock, 2012, from http://www.nord-lock.com/products/wedge-locking/washers/introduction/ [8] Youtube, 2007, “Nord-Lock Demonstration Video,” from http://www.youtube.com/watch?v=mgwmuZuJ02I

Acknowledgements The authors would like to acknowledge Joe Manahan, Ed Leubner, and the technicians and engineering employees of Cooper Crouse-Hinds for their recommendations, Dr. Michael Parthum for his unbiased design review, Dr. Stephen Boedo for reference material, Dave Hathaway, Rob Kraynik and *** for machining advice and assistance, and especially Dr. Benjamin Varela, our project guide.