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INVES N OVC UPD ILS (Es

Technical Repat
NR 203, ON

NAVAL FACILITIES ENGINEERING SERVICE CENTER
Port Hueneme, California 93043-4328

Technical Report
TR-2037-OCN

SEAWATER HYDRAULIC ROCK DRILL
IMPACT MECHANISM MODEL VALIDATION

John P. Kunsemiller
and Bruce W. Farber

DOCUMENT
BDAD\
LIBRARY
is:Hole Oceanccranh:
Woods Hole Oceanographic
Institution

January 1995

Sponsored by
Office of Naval Research

Approved for public release; distribution unlimited.

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1. AGENCY USE ONLY (Leave blank) 2. REPORT DATE 3. REPORT TYPE AND DATES COVERED
January 1995 Final; Oct 1991 through Sep 1994 |

] 4. TITLE AND SUBTITLE | 5. FUNDING NUMBERS
# SEAWATER HYDRAULIC ROCK DRILL IMPACT
| MECHANISM MODEL VALIDATION

PR - RM33F61
WU - DN662005

6. AUTHOR(S)

f John P. Kunsemiller and Bruce W. Farber

| 7. PERFORMING ORGANIZATION NAME(S) AND ADDRESSE(S) | 8. PERFORMING ORGANIZATION REPORT

NUMBER

| Naval Facilities Engineering Service Center
TR-2037-OCN

| 560 Center Drive
" Port Hueneme, CA 93043-4328

10. SPONSORING/MONITORING AGENCY REPORT
NUMBER

3. SPONSORING/MONITORING AGENCY NAME(S) AND ADDRESSES

Office of Naval Research
800 Quincy Street
Arlington, VA 22217-5000

11. SUPPLEMENTARY NOTES

12a. DISTRIBUTION/AVAILABILITY STATEMENT | 12b. DISTRIBUTION CODE
Approved for public release; distribution unlimited.

13. ABSTRACT (Maximum 200 words)

The Naval Facilities Engineering Service Center (NFESC) has developed and evaluated a computer
model of the single poppet-kicker port linear impact mechanism used in the seawater hydraulic rock drill.
The project objective was to identify hydraulic and dynamic elements influencing impact mechanism opera-
tion. A goal of this investigation was to determine whether or not computer modeling of the impact
mechanism for the seawater hydraulic rock drill would lead to an improved linear impact mechanism
suitable for a hand-held, diver-operated rock drill. Having achieved qualitative model response, the model j
| served as an analysis tool for evaluating proposed changes toward achieving a minimum 7 foot-pound blow
energy. Model parametric studies lead to validation of predicted performance improvements through hard-
ware tests using the pre-production seawater rock drill with modified impact mechanism component parts.
This document is the final report on a 3-year effort that has resulted in a better understanding of the
hydraulic and dynamic elements influencing impact mechanism operation as well as a technique for analyz-
ing complex fluid power components. At the conclusion of funding, the validated model showed a reliable 6
foot-pound blow energy. Additional development of the single poppet-kicker port linear impact mechanism
is recommended using the validated model as a basis for optimizing the linear impact mechanism.

| 14. SUBJECT TERMS

15. NUMBER OF PAGES |
45

Seawater, rock drill, tools, diver equipment, hydraulic equipment

17. SECURITY CLASSIFICATION | 18. SECURITY CLASSIFICATION 19. SECURITY CLASSIFICATION
OF REPORT OF THIS PAGE OF ABSTRACT

Unclassified Unclassified Unclassified UL

NSN 7540-01-280-5500 Standard Form 298 (Rev. 2-89)
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CONTENTS

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INTRODUCTION

Navy construction divers currently rely on oil hydraulic powered tools to perform various
underwater construction tasks. In today’s growing awareness for environmental responsibility,
the Navy is actively pursuing technology that will minimize the risk of violating environmental
regulations. One focus for technology development has been improved hydraulic tools for safe
operation in and around environmentally protected waters. Commercial oil hydraulic powered
tools can be replaced by intrinsically safe seawater hydraulic powered tools developed at the
Naval Facilities Engineering Service Center (NFESC).

Seawater has proven to be an effective and environmentally safe substitute for hydraulic
oil (Ref 1). The seawater hydraulic powered diver tool system, developed to satisfy the needs
of the underwater construction diver, includes a bandsaw, rotary disk tool, and rotary impact
tool. The system provides required capability without the hazards associated with using oil as
the hydraulic fluid. The fourth construction tool developed for the system, a seawater hydraulic
powered rock drill, has not been released because the linear impact mechanism was found to be
unreliable and subject to unexplained variations in performance (Ref 2).

This task under the Navy Exploratory Development Technology Program Plan was funded
through the Office of Naval Research in fiscal year 1992 with the objective of identifying the
hydraulic and dynamic elements influencing impact mechanism operation. A goal of this
investigation was to determine whether or not computer modeling of the impact mechanism for
the seawater hydraulic rock drill would lead to an improved linear impact mechanism suitable
for a hand-held, diver-operated rock drill. This document is the final report on a 3-year effort
that has resulted in a better understanding of the single poppet-kicker port linear impact
mechanism and a technique for analyzing complex fluid power components.

BACKGROUND

The seawater rock drill performance problem was attributed to the operation of the single
poppet-kicker port linear impact mechanism. This mechanism provides cycle timing as well as
impact energy to the drill operation. During testing of the original rock drill, small changes to
component dimensions were found to produce wide variations in drill performance. The various
influences on linear impact mechanism performance were not understood.

At the conclusion of the original seawater rock drill development, two recommendations
were presented. The first recommendation was to investigate alternate impact mechanism designs
in hopes of finding one more suited to development of a seawater hydraulic rock drill. The
second recommendation was for independent development of the single poppet-kicker port linear
impact mechanism. Computer modeling was recommended as a means to characterize
performance through parametric studies without the need for expensive hardware fabrications
(Ref 2).

The results of the computer modeling effort are documented in this report. The pursuit
of the first recommendation lead to a separate parallel effort for the development and
demonstration of a water hammer cycle impact mechanism. A successful Phase I proof of
concept and an ongoing Phase II prototype demonstration water hammer drill have been funded

through the Small Business Innovative Research (SBIR) program. At the conclusion of the SBIR
Phase II, the water hammer impact mechanism will be documented separately.

Requirements

The original design requirements for the seawater rock drill served as guidelines for this
impact mechanism development. Drill performance was stated as having a penetration rate
equivalent to the Stanley model HD-20 oil hydraulic hammer drill currently in the Underwater
Construction Team (UCT) inventory. This was restated for our purpose as an impact mechanism
capable of delivering to the drill steel and rock interface 7 foot-pounds of impact energy at a rate
of 30 cycles per second. At this performance level with a 3/4-inch-diameter drill bit, a
penetration of 3.5 inches per minute is calculated for rock having a 12,000 pounds per square
inch compressive strength (Ref 3).

The prototype rock drill as-built weight was 49 pounds. Though 9 pounds heavier than
the design requirement, it was previously agreed that the extra weight did not detract from drill
operability. In the model, the drill weight was set at 50 pounds.

Component Description

The Pre-Production Prototype (3P) rock drill, shown in Figure 1, is configured with a
single poppet-kicker port linear impact mechanism. The functional description of the cycle can
best be understood by reviewing the diagram shown in Figure 2.

During tool operation, water at supply pressure enters the drill through the trigger valve
and is directed into the drive chamber through the initially open supply poppet. This water flow
in turn drives the plunger and the piston down into the drill steel creating a percussive impact
at the rock surface. Near the end of the drive cycle, the relative position of the drive plunger
within the plunger sleeve causes the kicker port to pressurize and closes the supply poppet. The
closing of the supply poppet results in a bleed down of drive chamber pressure allowing the
piston return to reset the plunger. During plunger reset, the relative position of the drive plunger
within the plunger sleeve relieves the kicker port and permits the supply poppet to open. From
this point the cycle repeats.

Exhaust flow from the linear impact mechanism is directed to the 3-horsepower seawater
motor to index the drill steel. Motor exhaust is discharged to ambient out the motor exhaust
port.

DYNAMIC MODEL

The software selected for model development was the Dynamic Analysis and Design
System (DADS). The developed computer model of the linear impact mechanism qualitatively
mimicked empirical test results obtained during prior rock drill evaluation testing. After model
validation, parametric studies were performed for comparison to baseline model results. Model
refinements lead to predictions of performance improvements. These improvements were
validated using the 3P drill to test modified impact mechanism component parts.

The basic building blocks for the model consisted of bodies, some fixed to an inertial
reference frame, and some allowed to move; joints between the bodies; springs; hydraulic

accumulators; valves; actuators; line elements; and control functions. Figure 3 shows the basic
hardware components.
The control and hydraulic functions in the model were divided into four sections:

1 - Supply Pressure Activation
2 - Poppet Motion

3 - Plunger Motion

4 - Piston Motion

Detailed descriptions of each section are provided in Appendix A.

MODEL VALIDATION

Parametric studies were conducted using the computer model as a tool to optimize linear
impact mechanism performance. Model features were added or adjusted to simulate changes to
the linear impact mechanism design. Drill hardware testing validated predicted model results.
Since some of the model parameters were approximated rather than measured values, a precise
matching of the results was not expected nor necessarily desired. The focus of the effort was
to evaluate model trends and relate findings to specific hardware parameters. The first finding
dealt with the sensitivity of the model to small changes in supply poppet valve flow coefficient.

Poppet Valve Coefficient

As reported in Reference 2, the shape of the supply poppet seat affected drill operation.
It was empirically found that a grooved valve seat produced desirable results. This annular
groove had the effect of increasing the valve’s flow coefficient. The model showed weak
impacting from low piston velocities with smaller valve flow coefficients. Conversely, greater
impact energies from high piston velocities were modeled using larger valve flow coefficients.

An experiment was conducted to quantify the differences in behavior between the grooved
and the nongrooved poppet. It was also an objective to determine the validity of the valve
coefficient used in the model. Steady flow and dynamic flow conditions were evaluated. Figure
4 is a schematic of the test fixture.

During the static tests, the displacement of the poppet was fixed either full open or nearly
closed. The pressure drop across the valve was measured for several flow rates. The valve flow
coefficient obtained for the grooved and nongrooved poppet are shown in the graph of Figure
oF

No significant difference between poppets was observed for a nominal displacement of
0.050 inches. The value of the open poppet flow coefficient was close to the modeled value of
5. The smooth poppet yielded a higher valve flow coefficient for a nominal displacement of
0.015 inches. This was opposite of what was expected and has not been explained. The curves
also show that as the valve opens, the grooved poppet flow coefficient increases, whereas the
smooth poppet flow coefficient decreases.

During dynamic tests, poppet valve configurations for grooved and nongrooved poppets
were evaluated with and without the biasing spring. Evaluation of one open-close cycle was
representative of valve operation, however, manual control of the flow to the kicker port limited

cycle speed. Figure 6 presents two graphs comparing dynamic cycle test results for grooved and
nongrooved poppet seats, both with the biasing spring.

There did not appear to be any significant difference between the behaviors of the
grooved and nongrooved poppets. As in the static tests, the expected differences between the
grooved and nongrooved poppets was not substantiated.

One series of model runs focused on changing the area values in the poppet actuator
elements to cause the poppet to open more quickly. In the model, the desired effect was
achieved by increasing the opening area from 0.0182 square inches to 0.0382 square inches and
decreasing the closing area from 0.1329 square inches to 0.1129 square inches. The
improvement was small and suggests that enlarging the poppet area is not detrimental to impact
mechanism performance. Later, drill hardware tests showed equal results when using the
nongrooved poppet suggesting that the grooved feature is not required.

Minimized Drill Body Motion

Movement of the drill body was identified early in the evaluation process as a key
influence on impact mechanism performance. The impact mechanism optimization process,
therefore, used drill body displacement as a measure of cycle performance. During cycling of
the impact mechanism, hydraulic pressure driving the plunger and piston downward also lifts the
drill body up. The concern is the movement of the plunger sleeve in the drill body and the
telative position of the sleeve with respect to the plunger during the cycle. There is a point
where the drill/sleeve has moved far enough up from nominal such that the kicker port is
actuated before piston impact against the anvil. The result is a missed or weak impact caused
by the premature actuation of the kicker port to close off drive chamber supply.

Hardware testing was performed using a test stand that held the drill upright and allowed
the drill unrestricted vertical displacement. A force transducer was used to measure impact
force. This transducer was calibrated by dropping a 10-pound weight from 1 foot to obtain a
measurement for 10 foot-pounds of potential energy. A Temposonic position transducer
measured the vertical displacement of the drill body. One accelerometer was attached to the drill
body and a second to the impact plate at the base of the test stand. The accelerometers did not
provide useful data, since they were somewhat noisy. The measurements using the force
transducer and Temposonic transducer provided repeatable data.

The drill was modeled at the operating condition of 1,500 psi supply pressure with a 50-
pound force applied as the weight of the drill. This model showed a drill body displacement of
over 0.4 inches. Hardware testing confirmed a smaller 0.12-inch drill body displacement and
showed a qualitative improvement in drill cycle rate with the application of downward force to
the drill. This confirmed the relation between drill body movement and impact mechanism
performance.

Two methods were applied to minimize drill body motion. The approach was to first
optimize operating pressure for the existing drill weight of 50 pounds. Reduced operating
pressure resulted in less energy to the tool to cause drill body movement. This is evident in the
model results shown in Figure 7 obtained with an additional 150-pound downward force applied.
Best drill performance was selected at 1,000-psi system pressure.

Then, additional downward force representing increased drill weight was added until
consistent energy transfer was achieved. Adding 100 pounds downward force in the model was
insufficient for reducing drill body displacement while 150 pounds downward force reduced drill
body displacement by a factor of two. Adding 200 pounds or 400 pounds downward force

nearly eliminated drill body displacement. In all cases, the cycle repeatability in terms of
magnitude and repetition of impacts improved with an increase in applied downward force.
Figure 8 shows drill body displacement reductions for several values of added force.

Reductions to Plunger and Piston

From a practical standpoint, a 150-pound or more rock drill was not an acceptable
solution. What was desired was an optimized impact mechanism for the 50-pound drill weight.
Viewed as an energy balance problem, improvement was expected through a scaled reduction of
the plunger and piston size. The energy used to drive the scaled plunger and piston produced
less reaction force to displace the rock drill body.

Table 1 shows a comparison of several parameters used in the baseline 3P drill with those
tested as part of the validation of the reduced plunger and piston areas. Drawings of the
redesigned plunger, plunger sleeve, sleeve cap, and piston are provided in Appendix B.

Of several model runs at various values of plunger and piston areas, the best results were
produced with a plunger area reduction of 30 percent and a piston return area reduction of 36
percent. The ratio of plunger area to piston return area has a measurable effect on cycle timing.
Some slowing of the cycle was desired to preserve timing functions. This necessitated the 6
percent greater reduction in piston return area. Further investigation into timing effects revealed
that by opening the kicker port to the exhaust port sooner (i.e., move the kicker port closer to
the exhaust port in the plunger sleeve), the amount of external applied force could be reduced
to 100 pounds.

Table 1
Comparison of Evaluated Design Parameters

Parameter | Baseline Production Drill Redesign Validation

1,500 psi Tested at 600, 1,000, and
1,500 psi

Added Weight Tested at 0 and 100 lb

Plunger Head Diameter/ | 0.555 in. 0.466 in.
Area 0.24 in.? OM fas?
es | No, single diameter

Sleeve Piston

Plunger Length 3.687 in.
Piston Return Area 0.0807 in.”
Yes No

Operating Pressure

End Cap on Plunger

Plunger Stop (lower)

The decrease in plunger area from 0.24 square inch to 0.17 square inch was accomplished
with a plunger of the smaller diameter without an end cap feature. The plunger mass and travel
limits were adjusted in the model to represent this redesign. The single diameter plunger
traveled in contact with the piston throughout the piston stroke. To maintain cycle timing at the
lower end of the stroke, the plunger length was extended 0.31 inch to position the plunger cutout
between the kicker port and the supply pressure port. This plunger configuration eliminated the
"dead band" in piston stroke inherent in the 3P drill. The portion of the piston stroke in the 3P
drill where the plunger is no longer in contact and driving the piston is referred to as the "dead
band" portion of the piston stroke. The dead band occurred when the plunger came to rest in
the sleeve at the stop for the end cap. During this portion of the stroke, the piston undergoes
deceleration caused by piston seal friction and by a continuously energized piston return.
Eliminating the dead band by driving the piston continuously to anvil impact improved impact
mechanism efficiency.

A reduced upper piston diameter provided a net piston return area of 0.0514 square inches
giving a piston return force of 51 pounds at 1,000-psi supply pressure. The piston mass was
adjusted to represent the change. A spacer was added to the top of the piston sleeve to prevent
over stroking the lengthened plunger. This spacer did not decrease the operating stroke of the
impact mechanism as the stroke length is bound by the timing of the kicker port actuation. The
piston hard stop was modified to reflect the new location between the piston housing and the top
of the piston.

Referring to Table 2, testing of the baseline impact mechanism using the 3P drill parts
produced the results shown as Tests 1, 2, and 3 (direct measurement of body motion was not
available for these tests). The data showed erratic impacts and wide variability. No weight was
added to the drill for the first test set. The results of the baseline model, adjusted to match the
configuration of Test 1, are included in the lower portion of the table. The data for all 1,000-psi
supply pressure sets is in bold typeface for comparison.

A comparison of individual data for Tests 1 through 3 shows a large spread in the range
of impact energy. A closer review of the test data indicates a pattern of short cycling producing
inconsistent impact energy. Comparing Test 1 to the baseline model shows general agreement
in cycle time but the model predicted larger impacts with less variation. The addition of 100
pounds in the model had the expected increase in impact energy but also increased the variability
of the data.

Tests 4, 5, 6 were conducted with the 36 percent reduced area piston installed. This
resulted in longer cycle times and an order of magnitude less drill body displacement than
predicted by the baseline model. Tests at 1,000- and 1,500-psi pressures showed a double
impact with the second impact being far smaller than the first. The reason for the double impact
is discussed in the next section on damping coefficients. It is clear, however, that the double
impacts are responsible for the unexpected longer cycle times at the higher pressures. An
example of this double impact is shown in Figure 9 for the Test 5 data.

Table 3 presents the results from seven tests conducted with the 30 percent reduced
plunger and 36 percent reduced piston installed. For Test A2 and Test A7, a test operator leaned
full weight on the drill, applying an estimated 100 pounds of force downward on the drill. The
baseline model was updated to match the configuration of Test A2. The redesign model results
are displayed in the lower half of Table 3. The data for all 1,000-psi supply pressure sets is in
bold typeface for comparison.

Table 2
Comparison of Baseline Test Results

(psi) Weight (1b) | Length (ms) Range (fi-Ib) | Motion (in.)
es a
ee ee N/A

poe a Oe es sess wa]

| ee ae
DSO |e ss | aa) a eee One
= 450 | oo | | a st04.8

Baseline | 1,000 | 0 31 13to2.5 | 0.4
Model
Bascline || 1000 “| 100" | 31 1.5 to 4.5 0.4
Model |

Table 3
Comparison of Redesign Test Results
Added Cycle Eno

fa 1,000 2 tes ts

Impact

Body
Motion (in.)

aneronoe ee [news] in]
first: 0.7 0.01
Toa

arson | tawad | Poose el

1,500 first: 4.4 0.1, 0.01
0.7 to 2

1,350 first: 5 0.1, 0.01
0.7 to 1.6

ae mee 2ST || oe

— 1,000 2.3 to 3.3 0.27
Model

Comparing Table 3 data to Table 2 data, we find for baseline Test 2 and redesign Test
A1, both with no weight added, the impact energy increased a small amount presumably from
reduced drill body displacement. This was also the case for Test 1 and Test A3. This seems to
validate model prediction. However, this does not hold true when comparing Test 3 to Tests A4
or A5. It appears that the shorter cycle length in the later tests contributed to reduced impact
energy.

During data collection for Tests A5 and A6, the drill was observed to operate in two
modes. The first mode showed an initial displacement of the drill body, and then resulted in a
fast cycle with low force impacts about that displacement. The second mode caused the drill
body to rise and fall at low frequency with higher impact force. The pressure from the supply
also dropped slightly during the second mode operation indicating an increase in flow rate.
During Tests A3, AS, and A6, the initial impact was far higher than subsequent impacts
suggesting that the initial impact had a direct effect on mode selection and that Test A3 may also
have been a first mode operation. Figure 10 shows the drill body displacement and impact
energy for this condition.

Test A7 results are for the drill operating in the second mode with approximately 100
pounds applied to the drill. The larger amplitude force impacts tracked with drill displacement.
When the drill body displacement was small, the impact strength was greatest. This confirmed
model predictions for minimizing drill body displacement to maximize impacts. Figure 11
shows, for Test A7, nearly half of the impacts were at 6 foot-pounds or greater, though there
was still a large spread in the range for impact energy.

At 1,000-psi operating pressure and with 100 pounds weight added to the drill, the
redesign model showed slightly higher impact energy (2.3 to 3.3 foot-pounds versus 1.2 to 3
foot-pounds), significantly greater drill body motion (0.27 inches versus 0.01 inches), and shorter
cycle time (19 ms versus 24 ms). However, when compared with test results from Test A4, run
at 1,500 psi with no added weight, the cycle time and body motion of the redesigned model
match more closely.

Damping Coefficients

The model behavior relating impact energy to added weight/reduced drill body
displacement was evident in the hardware tests; however, model behavior seemed to correlate
better with results from higher pressure test conditions. Inconsistent cycling in the redesign model
and in hardware tests continued to thwart an exact comparison. An analysis of model parameter
for the damping coefficients was conducted to further validate the model to test conditions.

With the longer plunger remaining in contact with the piston until impact with the anvil,
the plunger cutout does not dwell in the "open" position (energizing the kicker port from the
supply port). In the baseline drill, the plunger stopped in a position that allowed sufficient time
to fully energize the kicker port. When the kicker port is not fully energized, the poppet does
not close leaving the plunger chamber pressurized for a secondary impact. This "double impact"
occurs with much less energy than the primary impact. Sometimes, a tertiary impact with even
less energy than the secondary impact may occur before a new cycle can begin.

Test results exhibited suspected "double-impact" events at higher operating pressures
(e.g., Test 6 at 1,500 psi). It was determined that insufficient damping was included in the hard
stops, which explained why double impacting occurred in the DADS model at a lower operating
pressure than in the test results. In the redesign model, a uniform value of 2.5 lb (in./sec) was
used for damping. This corresponded to a range of 8 to 28 percent of critical damping

depending on the element. The damping in each hard stop was changed individually to a value
representing 38 percent of critical damping (corresponding to a coefficient of restitution of 0.3,
a typical value for metal-to-metal contact). The redesign model was adjusted for increased
damping coefficients, with the results shown in Table 4.

Table 4
Comparison of Adjusted Model Results
[Pressure = 1,000 psi, Added Weight = 100 Ib]

Test ——— Length (ms) HUES Eos Body Motion — )
Range (ft-lb) y

Test [ee al 1.2 to 3.0 panscielceig) pad inh 01

Redesign Mode orem 2.3 103.3

38% Damping 25 NZ 0.004

Model

30% Damping 7P5) 1.4 0.007
| Model

Increased damping produced an order of magnitude reduction in predicted drill body
displacement matching test data under similar conditions. The model predicts a more regular
impact energy with little spread in the value. As expected, there was a corresponding decrease
in the impact energy for the model. Compared to Test A2, the 38 percent damping model
yielded an increased cycle length to slightly longer than that in the test data (25 ms versus 24
ms), but a decreased impact energy to the bottom of the test data range (1.2 foot-pounds versus
1.2 to 3 foot-pounds). A small reduction to the damping coefficients (30 percent of critical
damping) increased both the impact energy and body motion to values closer to those measured
in Test A2. Figure 12 presents the impact energy, body motion, and piston stroke relative
motion plots, respectively, for the model with the 30 percent damping coefficients.

RESULTS

A common occurrence during model development was the discovery of several variables
that influenced an observed parameter of impact mechanism performance. One such variable
involved the length of the plunger cutout which was known to control cycle timing. While
adjustment of the damping coefficients eliminated the double impact, the model showed that
lengthening the plunger cutout from 0.75 inches to 0.95 inches also eliminated the double impact
phenomenon even for higher pressures and impact energies. Though appropriate correction of
the damping coefficients achieved the desired effect, it is worth noting the possibility that impact
mechanism sensitivity to one variable may be reduced by smaller adjustments to several
variables.

In several model experiments, the ratio of plunger area to piston return area proved to
be an important variable for achieving stronger impacts. In the first example, the model showed
that for a 58 percent decrease in the piston return area, to 0.0340 square inches, the impact
energy increased by approximately 40 percent in comparison to the increased damping coefficient
model results. Since further reduction in piston return area was not desirable, the ratio was
adjusted by an increase in the plunger area.

In a model experiment, the plunger area was doubled to 0.34 square inches. The drill
cycled evenly, with impact energies averaging 6 foot-pounds (this was close to the desired result
of 7 foot-pounds). Despite the larger plunger area causing larger actuator forces, the drill body
displacement was less than 0.15 inch and there were no short cycles or missed impacts. The
shape of the plunger/piston motion curve has a sawtooth shape, with a long return run, allowing
gravity to act on the drill body to slow its upward speed and begin to return it to its starting
position. Each subsequent drive stroke imparted an upward displacement, but gravity mitigated
drill body displacement during a longer cycle time of 40 ms, or 1,500 beats per minute. Figure
13 shows the impact energy, body motion, and piston stroke relative motion plots for the model
with the doubled plunger area.

Finally, hardware tests showed the impact mechanism to be sensitive to flow restriction
at the exhaust orifice. When the rotary drive motor was connected to the circuit in series to the
impact mechanism, the impact cycles became erratic and the number of cycles per second
decreased. An externally applied force to the drill caused the rotary motor to stall, causing the
impact cycle to become irregular or stop.

Because the stalled motor increases the back pressure at the exhaust orifice, flow through
the impact mechanism is reduced to the point where the pressure across the impact mechanism
is insufficient to produce a cycle. A pressure relief valve installed on the motor supply line and
set to 350 psi produced no change in behavior. When the relief valve was set at 50 psi, the
impact cycles became more regular. In this condition, an externally applied force stalled the
motor but had no affect on cycle performance. Intermediate pressure settings produced a
corresponding reduction in cycle performance.

CONCLUSIONS

The following conclusions are presented as a summary of the development and validation
process for the computer model of the linear impact mechanism used in the seawater hydraulic
rock drill:

1. The computer model of the single poppet-kicker port linear impact mechanism has
been validated for both the baseline (using production parts), and the redesign rock drills. The
damping coefficients, bulk modulus, and other model parameter values have been determined
such that the model results closely match the test data for cycle time, body motion, and impact
energy. The validated model shows a reliable 6 foot-pound blow energy, just short of the
required 7 foot-pound level necessary for the rock drill design.

2. The linear impact mechanism model provides a cost effective tool for evaluating
design modifications. The predictions of drill performance for an applied force, a reduced
supply pressure, a change in poppet area, and a reduced plunger and piston size were
experimentally verified. The pressure and motion curves predicted by the model matched the

empirical behavior well. The use of the Dynamic Analysis and Design System (DADS) software
and the modeling techniques developed under this tasking has extended the Navy’s capability to
design complex seawater hydraulic system components.

3. The model predicts a small improvement in impact mechanism performance from an
increase in the supply poppet valve size. Validation testing confirmed that the groove in the seat
of the supply poppet valve is not necessary. Elimination of the grooved feature from the
production drill and substitution of a larger valve seat can reduce component cost.

4. The seawater motor, presently plumbed in series with the impact mechanism, produces
unacceptable back pressure variations to the impact mechanism resulting in erratic behavior and
poor impact mechanism performance. Model results confirm that the motor supply must be
decoupled from the linear impact mechanism drive chamber pressure if the impact mechanism
is to cycle properly.

5. The smaller, single diameter plunger resulted in a significant reduction in impact
energy with a corresponding reduction in drill body motion. Extending plunger length and
eliminating the dead band portion of the piston stroke improved impact mechanism efficiency.
Use of a single diameter plunger and plunger sleeve can reduce component fabrication cost while
improving impact mechanism performance.

6. The cause of the two distinct modes of drill body motion observed during drill
operation has not been identified. Impact mechanism performance continues to show dependence
on drill body motion. Test results confirm that minimizing drill body displacement is necessary
to maximize impact mechanism performance. Reducing the supply pressure, increasing the
applied force on the drill body, and changing the cycle timing have made demonstrable
improvements to impact mechanism performance because of small drill body displacement.

7. The model predicts impact energy values close to the design goal of 7 foot-pounds for
a plunger area of 0.34 square inches. The oversize plunger produced a slow return stroke and
maintained the drill body position within 0.15 inches. This allowed consistent impact energy.
Further improvements were predicted by the model for a lengthened plunger cutout which
successfully prevented double impacting, without decreasing impact energy.

RECOMMENDATIONS

A goal of this effort was to produce design data for a linear impact mechanism that could
be transitioned into a usable diver-operated rock drill. Complete realization of this goal has not
been attained. The single poppet-kicker port impact mechanism has proven to be a complex
mechanism. There is still much that is not understood such as the observed dual mode drill body
displacement behavior. The validated model now serves as a tool that can be used in the
optimization process to reach the design requirement of 7 foot-pounds impact energy. Because
it is more cost effective to conduct the optimization process using the validated computer model,
further optimization of the impact mechanism model is recommended prior to any hardware
testing. The following recommendations, based on observed model results, should provide the
basis for achieving the design requirement:

1. The relationship between plunger and piston areas should be fully explored. Model
experiments are recommended to determine the best values for plunger and piston sizes that will
produce more than 7 foot-pounds on a consistent basis, without increasing drill body motion.
Model results should be confirmed through testing of modified impact mechanism component
parts.

2. The lengthened plunger cutout should be experimentally evaluated as a means to
eliminate double impacting in combination with any future design seeking higher impact energies.

3. The seawater drive motor for bit indexing and related elements should be added to the
DADS model to determine what parameters can be altered to compensate for the observed
negative effects on hardware performance. The effects of adding the seawater motor in parallel
and in series to the linear impact mechanism should be compared for selection of the best drill
performance.

4. Finally, the Dynamic Analysis and Design System (DADS) software and the modeling
techniques developed under this tasking should be applied to the design of complex seawater
hydraulic system components. Use of this capability will enhance component performance and
result in significant cost savings over traditional build and test methods of hardware development.

REFERENCES

1. J. Kunsemiller and S. Black. "Case study of an environmentally safe diver tool system",
presented at the Underwater Intervention Conference, San Diego, CA, Feb 1994.

2. Naval Civil Engineering Laboratory. Technical Note N-1826: Development of a seawater
hydraulic rock drill, by J. Kunsemiller. Port Hueneme, CA, Mar 1991.

3. William C. Maurer. Advanced drilling techniques, Chapter 2. Tulsa, OK, Petroleum
Publishing Co., 1980.

Figure 1
Pre-Production Prototype (3P) rock drill.

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Basic model hardware components.

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Figure 6
Graph of dynamic test for grooved and nongrooved poppet seat.

Drill Body Motion (150 1b force)

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Drill Body Motion ¢1000 psi>

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of a 100- to 400-pound downward force.

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Appendix A

DETAILS OF DADS MODEL

DYNAMIC MODEL

The basic building blocks for the model consisted of bodies, some fixed to an inertial
reference frame, and some allowed to move; joints between the bodies; springs; hydraulic
accumulators; valves; actuators; line elements; and control functions. The control and hydraulic
functions in the model were divided into four sections:

1 - Supply Pressure Activation
2 - Poppet Motion

3 - Plunger Motion

4 - Piston Motion

Figure A-1 shows the activation of the supply pressure. The nominal supply pressure is
supplied by a large accumulator (SOURCE). The switch (VALVE.SWITCH) is initially closed
and, at a selected time, is opened to supply pressure to the drill (PS).

Figure A-2 shows the effects of poppet motion. The poppet is considered to act as a
valve (VALVE. POPPET) which controls flow to the drive chamber. A curve is used to translate
the displacement of the poppet (DIS.PPHPO1) into a spool setting (SPOOL.POPPET). In the
closed position, the supply pressure applies a force on the poppet (ACT.POPPET1). In addition
to the supply pressure acting to open the poppet, the kicker port pressure (PK) acts to hold the
poppet closed (ACT.POPPET2). Included is a small biasing spring that works to close the poppet
when the pressures are balanced.

When the poppet opens, additional area on the poppet is exposed, and the supply pressure
must be applied to the added area. This was initially modeled using a curve to represent the
added area and multiplying by the supply pressure to obtain the added force. This was later
changed to better reflect a more complex pressure distribution due to the fluid flow velocity
decreasing as the flow cross-sectional area increases. A closer approximation was achieved by
applying the pressure in the poppet chamber to the added area at all times. This change was
made by adding line element (LINE3) discussed below. Poppet chamber pressure (PPO) is
applied to the added area to hold the poppet open (ACT.POPPET3).

Figure A-3 shows the operation of the plunger. Flow from the supply port to the kicker
port, and from the kicker port to exhaust, is regulated by the position of the plunger. Curves
used to represent the plunger as a spool valve were based on the dimensions of the plunger cutout
and galleries in the plunger sleeve. These curves are used to translate the displacement of the

A-1

plunger (DIS.PPHPL2) into spool settings (SPOOL.PLUNGER.S and SPOOL.PLUNGER.E)
for valves controlling flow into and out of the kicker port (VALVE.PLUNGER.S and
VALVE.PLUNGER.E, respectively).

Leakage from the kicker port to exhaust was modeled by a small non-zero offset in the
exhaust curve (CURVE.PLUNGER.E) leaving the exhaust valve partially open at all times. A
similar offset in the supply curve (CURVE.PLUNGER.S) provided for supply leakage. Leakage
values of 2 percent for the supply and exhaust side of the plunger were determined to yield a
truer match to empirical data. However, the model was extremely sensitive to changes in this
parameter.

Line element (LINE2) represents fluid passage between the supply and the kicker port.
An additional line element (LINE3) was added between the poppet chamber and the plunger
drive chamber to provide information on the pressure drop and flow across the supply poppet,
and thus the pressure in the poppet chamber (PPO). The line element (LINE4) added between
the plunger drive chamber and the exhaust orifice allowed for a more accurate exhaust behavior
and produced a desired lag to the drive chamber discharge. This element provided information
on the pressure drop and flow to the exhaust orifice, and thus the pressure in the exhaust
chamber (PEY). Later, a line element (LINES) was added between the plunger and the kicker
port on the supply side. This line element served to reduce pressure pulsation in the kicker port
to match empirical data.

The drive chamber pressure (PPL), which is a function of the poppet setting, is
continuously exhausted through an open orifice (VALVE.EXHAUST). The exhaust condition
is represented by a near-infinite volume accumulator (EXHAUST) at ambient pressure (PEY).
The drive chamber pressure is applied to the plunger through an actuator (ACT.PLUNGER), and
a small force is created by the exhaust pressure acting on the back side of the plunger head
(PFORCE. PL).

Figure A-4 shows the working of the piston. The line element (LINE1) represents the
fluid between the supply and the piston chamber. The pressure in the piston chamber (PPI) is
applied to the differential piston area (ACT.PISTON). The force from this actuator
(ACT.FORCE.PISTON), which is calculated as a by-product of the element, is multiplied by
a fraction, then applied as a seal friction force (FFORCE.PI). The sign of the piston velocity
(DISD.PHPI1) is used to set the sign of the friction force to oppose motion.

Motion of the poppet, plunger, piston, and anvil was limited by hard stops at the upper
and lower limits of body motion from housing contact. Hard contacts were also implemented
at the interfaces for plunger to piston, and piston to anvil. These hard contacts were
implemented through a set of control elements which created a compression only spring damper
force. A typical hard-stop control element diagram is shown in Figure A-5. The string labeled
"XY" represents the names for the two bodies such as "PLPI" for plunger to piston contact. The
displacement (DIS.XY) causes a force (F) to be applied between the two bodies only for values
less than zero. The damping coefficient is a function of velocity (DISD.XY).

The coordinate system locations for the noncentroidal reference frames and centers of
gravity for the redesign model are shown in Figure A-6. Figure A-7 shows the locations for the
hard stops as they were modeled.

A-2

INPUT.SWITCH

Figure A-1
Supply pressure flow chart for model.

A-3

MULT.FFORCE.PI

curve.mu.pi
ood | -{

DISD.PHPI1 FFORCE.PI

AMP.MU.PI

FFORCEO.PI OUT.FFORCE.PI

INPUT.DISD.PHPI1

AMP.FFORRCEO.PI

LINE1 ACT.PISTON

Figure A-4
Piston flow chart for model.

SPRING.XY

INPUT.DIS.XY
SUM.STOP.XY

OUT.STOP.XY

MULT.DAMP.XY
INPUT.DISD.XY

Figure A-5
Sample of hard stop control flow chart.

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Coordinate system locations for the noncentroidal reference frames
and centers of gravity for the redesign model.

A-7

PPHPO|
PPHPL |

PPHPO2

PHP | |
Piel? |

Piriealy2

PIG

AHG2

Figure A-7
Locations for the hard stops as they were modeled for the redesign model.

A-8

Appendix B

IMPACT MECHANISM PARTS DRAWINGS

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DISTRIBUTION LIST

AF / ROBERTS, TALLAHASSEE, FL

APPLIED TECHNOLOGY AND MANAGEMENT / C. JONES, CHARLESTON, SC

ARMY CEWES / (TW RICHARDSON), VICKSBURG, MS

ARMY EWES / PERRY, VICKSBURG, MS; WESCD-P (MELBY), VICKSBURG, MS; WESCV-Z
(WHALIN), VICKSBURG, MS

ARVID GRANT & ASSOC / OLYMPIA, WA

BATTELLE / D. FRINK, COLUMBUS, OH; D. HACKMAN, COLUMBUS, OH

BECHTEL CIVIL, INC / K. MARK, SAN FRANCISCO, CA

BING YEN AND ASSOCIATES, INC / IRVINE, CA

BLAYLOCK ENGINEERING GROUP / T SPENCER, SAN DIEGO, CA

BUREAU OF RECLAMATION / D-1512 (GS DEPUY), DENVER, CO; SMOAK, DENVER, CO

CAL STATE UNIV / C.V. CHELAPATI, LONG BEACH, CA

CHAO, JC / HOUSTON, Tx

CHEE, WINSTON / GRETNA, LA

CHESNAVFACENGCOM / CODE FPO-1C, WASHINGTON, DC; LOWER, WASHINGTON, DC

CHILDS ENGRG CORP / K.M. CHILDS, JR., MEDFIELD, MA

CITY OF MONTEREY / CONST MGR (REICHMUTH), MONTEREY, CA

COLLINS ENGRG, INC / M GARLICH, CHICAGO, IL

COLORADO SCHOOL OF MINES / CHUNG,

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EXXON PRODUCTION RSCH CO / OFFSHORE OP DIV, HOUSTON, TX

GEOCON INC / CORLEY, SAN DIEGO, CA

GRUMMAN AEROSPACE CORP / TECH INFO CENTER, BETHPAGE, NY

HALEY & ALDRICH, INC. / CETIN SOYDEMIR, CAMBRIDGE, MA

HAN-PADRON ASSOCIATES / DENNIS PADRON, NEW YORK, NY

HARDY, S.P. / SAN RAMON, CA

HARTER, J.V. / METAIRIE, LA

HERONEMUS, W.E. / AMHERST, MA

INTL MARITIME, INC / D. WALSH, MYRTLE POINT, OR

IOWA STATE UNIV / CE DEPT, AMES, IA

LANDAU ASSOC / STIRLING, EDMONDS, WA

LONG BEACH PORT / ENGRG DIR (ALLEN), LONG BEACH, CA; ENGRG DIR (LUZZ]),
LONG BEACH, CA

MARATHON OIL CO / GAMBLE, HOUSTON, TX

MARITECH ENGRG / DONOGHUE, AUSTIN, TX

MBARI / SHANE, MOSS LANDING, CA

MC CLELLAND ENGRS, INC / LIB, HOUSTON, TX

MC DERMOTT, INC / E&M DIV, NEW ORLEANS, LA

MCAS / 16, MAG, TUSTIN, CA; PWO, KANEOHE BAY, HI

MCRD / PWO, SAN DIEGO, CA

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NAVSTA / ENGR DIV, PWD, FPO AA; PWO, ROTA, SPAIN, FPO AE

NAVSTA PUGET SOUND / CODE 922, EVERETT, WA

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NOAA / JOSEPH VADUS, ROCKVILLE, MD

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NSF / (N32), NORFOLK, VA

NUSC / CODE 0031 DIVING OFFR, NEWPORT, RI

NUSC DET / NEWPORT, RI; CODE 2143 (VARLEY), NEW LONDON, CT; CODE TA131,
NEW LONDON, CT; DOC LIB, NEW LONDON, CT

OCEANEERING TECH / VENTURA, CA

OCNR / CODE 1121 (EA SILVA), ARLINGTON, VA

OMEGA MARINE, INC. / SCHULZE, LIBRARIAN, HOUSTON, TX

PACIFIC MARINE TECH / M. WAGNER, DUVALL, WA

PACNAVFACENGCOM / CODE 102, PEARL HARBOR, HI

PURDUE UNIV / CE SCOL (ALTSCHAEFFL), WEST LAFAYETTE, IN; CE SCOL (LEONARDS),
WEST LAFAYETTE, IN; ENGRG LIB, WEST LAFAYETTE, IN

PWC / CODE 421 (KAYA), PEARL HARBOR, HI; CODE 421 (QUIN), SAN DIEGO, CA

Q ASSOCIATES / QUIRK, J PANAMA CITY, FL

R J BROWN & ASSOC / R PERERA, HOUSTON, TX

SAN DIEGO PORT / AUSTIN, SAN DIEGO, CA

SAN DIEGO STATE UNIV / CE DEPT (NOORANY), SAN DIEGO, CA

SCHUPACK SUAREZ ENGRS INC. / SCHUPACK, NORWALK, CT

SEAL TEAM / 6, NORFOLK, VA

SEATTLE UNIV / CE DEPT (SCHWAEGLER), SEATTLE, WA

SOUTHWEST RSCH INST / KING, SAN ANTONIO, TX

T.C. DUNN / SHREWSBURY, MA

TEXAS A&M UNIV / CE DEPT (HERBICH), COLLEGE STATION, TX; OCEAN ENGR PROJ,
COLLEGE STATION, TX

TWELVE OAKS BUSINESS PARK / WEST HAVEN, CT

UNIV OF CALIFORNIA / CE DEPT (MITCHELL), BERKELEY, CA; NAVAL ARCHT DEPT,
BERKELEY, CA

UNIV OF HAWAII / LOOK LAB, OCEAN ENGRG, HONOLULU, HI; MANOA, LIB, HONOLULU, HI;
OCEAN ENGRG DEPT (ERTEKIN), HONOLULU, HI

UNIV OF MICHIGAN / CE DEPT (RICHART), ANN ARBOR, MI

UNIV OF NEW HAMPSHIRE / LAVOIE, DURHAM, NH

UNIV OF RHODE ISLAND / PELL MARINE SCI LIB, NARRAGANSETT, RI

UNIV OF TEXAS / CONSTRUCTION INDUSTRY INST, AUSTIN, TX

UNIV OF WISCONSIN / GREAT LAKES STUDIES CEN, MILWAUKEE, WI

USACOE / CESPD-CO-EQ, SAN FRANCISCO, CA

USAE / CEWES-IM-MI-R, VICKSBURG, MS

USDA / FOR SVC, EQUIP DEV CEN, SAN DIMAS, CA

USN / CAPT COLIN M JONES, HONOLULU, HI

USNA / OCEAN ENGRG DEPT, ANNAPOLIS, MD

VSE / OCEAN ENGRG GROUP (CHASE), ALEXANDRIA, VA

WESCR-P / HALES, VICKSBURG, MS

WESTNAVFACENGCOM / CODE 407, SAN BRUNO, CA; ROICC, SILVERDALE, WA
WISWELL, INC. / SOUTHPORT, CT

WOODS HOLE OCEANOGRAPHIC INST / DOC LIB, WOODS HOLE, MA
WOODWARD-CLYDE CONSULTANTS / R. CROSS, OAKLAND, CA; WEST REG, LIB, OAKLAND, CA

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