## Compressible Flow Through Nozzles, and the Vulcan 06 Valve

Most of our fluid mechanics offerings are on our companion site, Chet Aero Marine.  This topic, and the way we plan to treat it, is so intertwined with the history of Vulcan’s product line that we’re posting it here.  Hopefully it will be useful in understanding both.  It’s a offshoot of Vulcan’s valve loss study in the late 1970’s and early 1980’s, and it led to an important decision in that effort.  I am indebted to Bob Daniel at Georgia Tech for this presentation.

## Basics of Compressible Flow Through Nozzles and Other Orifices

The basics of incompressible flow through nozzles, and the losses that take place, is discussed here in detail.  The first complicating factor when adding compressibility is the density change in the fluid.  For this study we will consider only ideal gases.

Consider a simple orifice configuration such as is shown below.

The mass flow through this system for an ideal gas is given by the equation

$\dot{ m }=A'_{{o}}\rho_{{1}}\left ({\frac {p_{{2}}}{p_{{1}}}}\right )^{{k}^{-1}}\sqrt {2}\sqrt {g_{{c}}kRT_{{1}}\left (1-\left ({\frac {p_{{2}}}{p_{{1}}}}\right )^{{\frac {k-1}{k}}}\right )\left (k-1\right )^{-1}}{\frac {1}{\sqrt {1-{A_{{o}}}^{2}\left ({\frac {p_{{2}}}{p_{{1}}}}\right )^{2\,{k}^{-1}}{A_{{1}}}^{-2}}}}$

where

• $\dot{m} =$ mass flow rate, $\frac{lb_m}{sec}$
• $A_o =$ throat area of orifice, $ft^2$
• $A'_o =$ adjusted throat area of orifice (see below,) $ft^2$
• $\rho_1 =$ upstream density, $\frac{lb_m}{ft^3}$
• $p_1 =$ upstream pressure, psfa
• $p_2 =$ downstream pressure, psfa
• $g_c =$ gravitational constant $= 32.2 \frac{lb_m-ft}{lb_f-sec^2}$
• $k =$ ideal gas constant or ratio of specific heats $= 1.4$ for air
• $R =$ gas constant $= 53.35 \frac{ft-lb_f}{lb_m\,^\circ R}$
• $T_1 =$ upstream absolute temperature $\,^\circ R$

At this point we need to state two modifications for this equation.

First, we need to eliminate the density, which we can do using the ideal gas equation

$\rho_1 = {\frac {p_{{1}}}{RT_{{1}}}}$

Second, we should like to convert the mass flow rate into the equivalent volumetric flow rate for free air.  Most air compressors (and our goal is to determine the size of an air compressor needed to run a test through this valve) are rated in volumetric flow of free air in cubic feet per minute (SCFM.)  This is also the basis for the air consumption ratings for Vulcan hammers as well, both adiabatic and isothermal.  This is accomplished by using the equation

$\dot{m} = {\frac {1}{60}}\,{\it SCFM}\,\rho_{{{\it std}}}$

Making these substitutions (with a little algebra) yields

$SCFM = 60\,A'_{{o}}p_{{1}}\left ({\frac {p_{{2}}}{p_{{1}}}}\right )^{{k}^{-1}}\sqrt {2}\sqrt {-g_{{c}}kRT_{{1}}\left (-1+\left ({\frac {p_{{2}}}{p_{{1}}}}\right )^{{\frac{k-1}{k}}}\right )\left (k-1\right )^{-1}}{\rho_{{{\it std}}}}^{-1}{R}^{-1}{T_{{1}}}^{-1}{\frac {1}{\sqrt {-\left(-{A_{{1}}}^{2}+{A_{{o}}}^{2}\left ({\frac {p_{{2}}}{p_{{1}}}}\right )^{2 \,{k}^{-1}}\right){A_{{1}}}^{-2}}}}$

In this article the coefficient of discharge $C_D$ is discussed.  It is also the ratio of the effective throat area to the total throat area, or

$A'_o = C_DA_o$

We are basically considering the energy losses due to friction as an additional geometric constriction in the system.

One final–and very important–restriction on these equations is the critical pressure, given by the equation

$p_c =p_{{1}}\left (2\,\left (k+1\right )^{-1}\right )^{{\frac {k}{k-1}}}$

The critical pressure is the downstream pressure for a given upstream pressure below which the flow is “choked,” i.e., the mass or volumetric flow rate will not increase no matter how much you either increase the upstream pressure or decrease the downstream pressure.  This limitation, which was observed by Saint-Venant, is due to achieving the velocity of sound with the flow through the nozzle or valve.  A more common way of expressing this is to consider the critical pressure ratio, or

$p_{cr} = \frac{p_c}{p_{{1}}} = \left (2\,\left (k+1\right )^{-1}\right )^{{\frac {k}{k-1}}}$

As you can see, this is strictly a function of the ideal gas constant.  It’s certainly possible to get around this using a converging-diverging nozzle, but most nozzles, valves or orifices are not like this, and certainly not a Vulcan 06 valve.  We now turn to the analysis of this valve as an example of these calculations.

## Application: the Vulcan 06 Valve

The first thing we should note is that pile driving equipment (except that which is used underwater) is designed to operate at sea level.  Using this calculator and the standard day, free air has the following properties:

• Temperature: 518.67 °R
• Density: $\rho_{std} = 0.00237 \frac{slugs}{ft^3} = 0.0763 \frac{lb_m}{ft^3}$
• Pressure: $2116.22 \frac{lb}{ft^2}$ (or psfa)

Now let’s consider the valve for the 06 hammer (which is identical to the #1 hammer.)  A valve setting diagram (with basic flow lines to show the flow) is shown below.

Note the references to steam.  Until before World War II most of these hammers (along with most construction equipment) was run on steam.  With its highly variable gas constant and ability to condense back to liquid, steam presented significant analysis challenges for the designers of heavy equipment during the last part of the nineteenth century and the early part of the twentieth.  For our purposes we’ll stick with air.

There are two cases of interest:

• The left panel shows the air entering the hammer and passing through the valve to the cylinder.  Pressurising the cylinder induces upward pressure on the piston and raises the ram.  The valve position (which shows the inlet port barely cracked) is shown for setting purposes; in operation the valve was rotated more anti-clockwise, opening the inlet port.
• The centre panel shows exhaust,  where air is allowed to escape from the cylinder.  The piston is no longer pressurised and the ram falls to impact.

According to the vulcanhammer.info Guide to Pile Driving Equipment, the rated operating pressure for the Vulcan 06 at the hammer is 100 psig = 14,400 psfg = 16,516.22 psfa = 114.7 psia.  For simplicity’s sake, we can consider the two cases as mirror images of each other.  In other words, the upstream pressure in both cases is the rated operating pressure.  This should certainly be the case during air admission into the hammer.  For the exhaust, it should be true at the beginning of exhaust.  Conversely, at the beginning of intake the downstream pressure should be atmospheric (or nearly so) and always so for exhaust.

From this and the physical characteristics of the system, we can state the following properties:

• Upstream pressure = 114.7 psia
• Downstream pressure = 14.7 psia
• Upstream area (from hammer geometry, approximate) $A_1 = 0.00705 ft^2$
• Throat area $A_o = 0.00407 ft^2$
• Coefficient of Discharge, assuming sharp-edge orifice conditions $C_D = 0.6$
• Adjusted throat area $A'_o = 0.00407 \times 0.6 = 0.002442 ft^2$

At this point calculating the flow in the valve should be a straightforward application of the flow equations, but there is one complicating factor: choked flow, which is predicted using the critical pressure ratio.  For the case where $k = 1.4$, the critical pressure ratio $p_{cr} = .528$.  Obviously the ratio of the upstream pressure and the downstream pressure is greater than that.  There are two ways of considering this problem.

The first is to fix the downstream pressure and then compute the upstream pressure with the maximum flow.  In this case $p_1 = \frac{p_{atm}}{p_{cr}} =$ 27.84 psia = 13.14 psig.  This isn’t very high; it means that it doesn’t take much pressure feeding into the atmosphere to induce critical flow.  It is why, for example, during the “crack of the exhaust,” the flow starts out as constant and then shortly begins to dissipate.  The smaller the orifice, the longer the time to “blow down” the interior of the hammer or to fill the cylinder with pressurised air.

The reverse is to fix the upstream pressure and then to vary the downstream pressure.  The critical downstream pressure is now $p_2 = p_1 \times p_{cr} = 114.7 \times 0.528 =$ 60.59 psia = 45.89 psig.  This means that, when the cylinder is pressurising at the beginning of the upstroke, the cylinder pressure needs to rise to the critical pressure before the flow rate begins to decrease.

We will concentrate on the latter case.  If we substitute everything except the downstream pressure (expressed in psia,) we have

$SCFM = 0.05464605129\,{\frac {{{\it p_2}}^{ 0.7142857143}\sqrt { 3126523.400-806519.7237\,{{\it p_2}}^{ 0.2857142857}}}{\sqrt { 0.9999999996-0.0003806949619\,{{\it p_2}}^{ 1.428571429}}}}$

If $p_2$ falls below the critical pressure, the flow is unaffected by the further drop and is constant. In this case the critical flow is 795 CFM.  For downstream pressures above the critical pressure, the flow varies as shown below.

As noted earlier, when air is first admitted into the cylinder the flow is constant.  Once the critical pressure ratio is passed, the flow drops until the two pressures are equal.

It was this large volume of flow which prevented the use of the 06 valve (which could have been separated from the cylinder using a valve liner) in the valve loss study.  The smaller DGH-100 valve was used instead.

It is interesting to note that the rated air consumption of the hammer is 625 cfm.  This is lower than the instantaneous critical flow.   Although on the surface it seems inevitable that the hammer will “outrun” the compressor, as a further complication the hammer does not receive air on a continuous basis but on an intermittent one.  For much of the stroke the compressor is “dead headed” and no air is admitted into the cylinder from the compressor.  To properly operate such a device, a large receiver tank is needed to provide the flow when it is needed.  The lack of such large tanks on modern compressors is a major challenge to the proper operation of air pile hammers.

## The Valve Loss Study

All fluid flow in Vulcan hammers is regulated and directed by a valve.  For most Vulcan hammers (the California series being a notable exception, the #5 is another) the valve is a Corliss type valve modified from those used in steam engines.  Simple and reliable, it, like any other valve, is subject to losses as the air or steam passes through it.  These are reflected in the mechanical efficiency of the hammer.

The losses due to air or steam flowing through the valve are generally not the most significant source of energy losses in a pile hammer.  In the late 1970’s and early 1980’s, with the increase in sheer size of the hammers, these losses became of more concern.  It was necessary to at least attempt to quantify these losses instead of using a “standard” back pressure value.

In May 1979 Vulcan contacted the Georgia Institute of Technology in Atlanta about using a Vulcan #1 series valve (like used in the #1, 06, etc.) in a test to determine the losses of air flowing through these valves.  At this point a major problem was encountered: the air flow required to properly test the valve was too large for Georgia Tech’s equipment.  Reaching out to Lockheed didn’t help either; they couldn’t do it.  At this point Vulcan came up with an alternative: use the DGH-100 valve, which was a Corliss valve albeit much smaller, for the test.  Making things easier was the fact that the DGH-100 used a small aluminium valve chest, which made the valve mounting simpler.

This proved feasible and Vulcan received a proposal from Brady R. Daniel at Georgia Tech for these tests.  The valve was tested in two “configurations”:

The tests were run and the report was presented in October 1980.  The immediate results were as follows:

1. The report showed that the valve could be modelled essentially as a sharp-edge orifice.  In the context of incompressible fluids, this is explained here.
2. A numerical method was developed to analyse the hammer cycle, as opposed to the closed-form solutions that had been used since the beginning of Vulcan pile hammers.  This led to some design changes, and was also adapted for the Single-Compound hammer design.

The report also contained some suggestions for “streamlining” the design of the valve.  These were not adopted, and the reason should be noted.

With the Corliss type valve, the Valve Port 1 is continuously pressurised, and this in turn forces the valve against the valve chest (or liner in the case of most newer Vulcan hammers.)  With proper lubricant this seals the valve and further sealing (rings, seals, etc.)  are unnecessary.  This is a major reason why Vulcan hammers are as reliable as they are under the dire circumstances many operate.  But that comes with a price.  As with any design, there are trade-offs, and in this case the simplicity of the valve is traded off for efficiency.  The simplest way to deal with this is to properly size the valve, and this was the main reason for the Valve Loss Study.

The Valve Loss Study is an interesting example of design analysis (others are here) which even an old product line like Vulcan’s can benefit from.

## General Arrangements and Assemblies

One of the typical information items Vulcan would send out would be the “general arrangement” (or assembly, to use the Raymond terminology) of a hammer, or a sub-assembly such as a capblock follower. These were also included in the offshore field service manuals. Sometimes they would feature the specifications of the hammer. They are useful for basic clearance and other dimensions or to understand the basic layout of the machine.

Some of these were put in data format. We feature for download some collections of these as follows:

• Vulcan 020 Offshore Hammer Specification Sheet. Not a general arrangement per se, but a specification sheet (in US and SI units) along with parts of the general arrangement on the back. These were issued in the 1970’s and were very popular for many years.
• Vulcan 040 Offshore Hammer Specification Sheet.
• Vulcan Offshore Hammers
• Auto-Jack Cable Tensioning Device for most Vulcan offshore hammers
• Vulcan 535 Hammer, 54″ and 80″ Jaws (similar to the 530)
• Vulcan 530/535 Capblock Follower Assembly (80″ Jaws)
• Vulcan 560 Hammer
• Vulcan 5110 Hammer
• Vulcan 5100 Capblock Follower Assembly
• Vulcan/Raymond Hammers
• Vulcan 513 Hammer
• Vulcan 515 Hammer
• Vulcan 517 Hammer
• Vulcan 525 Hammer
• Vulcan/Raymond 60X Hammer, with and without Vari-Cycle II
• Vulcan/Foster Vibratory Hammer. Vulcan manufactured L.B. Foster vibratory hammers during the 1990’s on a “private label” basis. These are the general assemblies for the 1050 and 4200.

Some of our general arrangements are in image format; we present some of them below.

We also have an extensive collection of these (including the specification sheets) in other “traditional formats.” If you would like to contact us about obtaining these, click here. We also have extensive information in our Vulcan Data Manual.

## Decelflo Pile Hammer Muffer, and the Thruflo Muffler

One thing most people notice first about pile driving jobs is that they generate an elevated level of noise. Until the 1960’s, most people simply put up with this and many other aspects of industrialisation and development. In the early 1970’s, Vulcan and other pile driving equipment manufacturers were confronted with new regulations–both at the federal and local level–which sought to regulate the noise output of construction equipment. Needless to say, pile driving equipment was high on the list.

Vulcan’s first reaction was to study the issue. It retained the services of United Acoustical Consultants in Glastonbury, CT, and its principal, Stannard Potter, to study the nature of noise output of Vulcan hammers. In December 1972 they conducted a study of a Vulcan #1 at the Chattanooga facility, and the report on the test is shown here.

Vulcan hammers, although durable and simple, suffered from two specific difficulties for noise abatement: a) their open construction gave little natural noise attentuation, and b) their lack of a recoil dampener increased the impact load on the frame, thus making it difficult to attach shrouds and other devices to attenuate sound. However, one of the results of the study was that a large proportion of the sound emission from an operating Vulcan hammer came from the exhaust. Since muffling the exhaust was simpler than doing same with the impact, Vulcan commissioned Potter to design an exhaust muffler, which it called the Decelfo Muffler.

Below: a diagram of the Decelflo Muffer concept, from Potter’s U.S. Patent 3,981,378.

The muffer was simple, a box which directed the air or steam output of the exhaust through perforated pipe surrounded by acoustical foam. The drawing shows a stacked arrangement for the muffler, but Vulcan never employed this arrangement.

The first test of the Decelflo took place in October 1973 in the Alameda yard of Santa Fe construction in Alameda, California. It involved muffling a Vulcan 020 hammer.

As shown below, the test was successful; the muffer performed as anticipated and its used resulted in reduction of hammer noise.

Flush with this success, Vulcan continued in its development of the muffler. In July 1974 it had another opportunity to demonstrate (and verify) the Decelflo’s capabilities, this time in Chicago at a sheet piling project. Below: the Decelflo mounted on top of the Vulcan hammer, in this case a 50C. The hose connection from the exhaust port to the muffler can be clearly seen, along with its connection to the hammer via the sheave pin. For a photo of the muffer in action during pile driving, click here.

We have an audio clip from this test which compares the hammer sound with and without the muffer; you can click here to listen to it and compare for yourself.

Vulcan had great plans for the Decelflo; at this time it was working on a method to mount the muffler directly on the hammer, as shown below.

But then things took a strange twist.

To begin with, there was considerable contractor resistance to the concept of having to add another device to the hammer assembly. Mounting it above the hammer lengthened the leaders required to operate the hammer, and the large installed base of Vulcan hammers dictated that this would be the normal way the muffler would be mounted.

Beyond that, the level of noise emissions, and how people perceive them, vary widely from one jobsite to another. This variation is a function of the location of the job (urban, remote, etc.), the presence or absence of neighbouring buildings to reflect the sound, and whatever ambient noise is in proximity to the jobsite. For example, driving piling next to an existing interstate, with the road noise already present, may not be very perceptible.

Finally, as far as those working on the jobsite are concerned, contractors (and OSHA) found it simpler to deal with noise emissions from pile drivers and other equipment on site by providing hearing protection to the workers, which of course is standard on jobsites today.

In any case, the Decelflo muffler was never very popular, “noise pollution” never achieved the notoriety of air and water pollution, and both Vulcan and its customer base moved on to other concerns. For his part Stan Potter moved on to patent the Decelflo concept independently of Vulcan.

### Thruflo (Geothermal) Muffer

The need to attenuate noise combined with another concern of the era, the need for alternative energy resources, with the geothermal muffer. An experimental product, it nevertheless touched on issues that are still important today.

Geothermal energy is possible when the hot magma which exists in the earth is close enough to the surface and the underground water to turn the latter into steam, which can be used to drive the turbines and generators to produce electric power. The means that the source of geothermal energy is not only free economically, but also that carbon dioxide (greenhouse gas) is not emitted in the production of electricity.

In the course of producing energy, the steam is vented to the atmosphere, and unmuffled this can produce a high noise level.

Vulcan built a prototype and tested it in its own facility in August 1974.

Below: measuring the sound as the steam passes through the straight pipe (left) and the muffler (right.) You can hear the difference by clicking here to hear the audio clip of the test.

Unfortunately the Thruflo Muffer did not get past the prototype shown above. Some of the mufflers that did make it to The Geysers had a difficult time of it, as this report attests.

## Vulcan Hammer Noise Study

Note: This study was commissioned by Vulcan and conducted in early December 1972 by United Acoustical Consultants of Glastonbury, CT, and dated 23rd June 1973. The report was submitted by the President of UAC, Mr. Stannard M. Potter. It has been released publicly by Vulcan at various times since its completion. This is the text for “Volume II” to the entire study. The graphs and other external references given in the report are not available. The fine print for this document applies.

## A. DESCRIPTION OF TESTS

During the week of December 4th through December 9th, 1972, tests were conducted on a standard Vulcan Hammer, Size 1, at the Pile Hammer Test Stand in Chattanooga. Tennessee.

### 1. Test Site

The normal structural steel test rig was removed and a special wooden support was provided to support the hammer in the test pit. The wooden beam structure prevented the secondary noise source from the steel structure from intervening with the hammer noise.

Unfortunately, the test stand site is within 75′ of the main manufacturing area of the Vulcan Plant. This provides a serious reflection of the noise from the hammer. There are other reflecting surfaces nearby. though they are smaller and at greater distances (250′). To the east of the test stand, the ground surface is hard pavement for all of the microphone locations. Although the site had acoustic shortcomings, it was felt that the proximity of the Vulcan Plant and its personnel was an obvious advantage over a possible “free field” at some distance from the plant. UAC’s Instrument Van was located in a supply shed next to the boiler room for the test stand at a distance of about 60′.

For the noise tests, the boiler and steam supply for the hammer were replaced by a truck-mounted air compressor unit. The truck was located in the front parking lot, as far away as the hoses would reach. A special muffler was designed to augment the exhaust noise reduction of the air compressor’s muffler. During the first two tests, this muffler proved inadequate and was replaced by a second unit, starting with Test 3 of the Muffled Auxiliary Exhaust.

The weather throughout the testing period was frequently cold and rainy. Wind protection was provided for the microphones and a plastic shield was built of thin vinyl in tent form over the hammer test rig to facilitate working during the rain during Test 11.

### 2. Instrumentation

Four microphone outputs were recorded on magnetic tape simultaneously. Three were data positions and one was for general announcements.

 Microphone Location A Close to The Hammer – generally 6″ away from the radiating surface. (See Vulcan Drawing D-10249) 7 different microphone placements. B 25′ from the centreline of the Hammer 3 positions: Northeast, East, Southeast C 50′ from the centerline of the Hammer 3 positions: Northeast, East, Southeast

Direct readings of Peak Impact Noise were made on an oscilloscope for later verification of laboratory playback.

For most of the tests. a direct field chart was made of the A-weighted output for use in direct assessment of the signal character during each of the tests.

A detailed list of the instrumentation is at the end of this section.

### 3. Hammer Configurations

There were eleven basic changes made to the hammer and its cushion, as listed below:

 Test No. Hammer Configuration 1 Bare Hammer – Steel Cushion 2 Muffled Exhaust 3 Mufflex Auxiliary Exhaust and Air Compressor 4 Nylon Slide Bar 5 Ascon Cushion 6 Micarta Cushion 7 Wood Cushion 8 Wrapped Base + Double Auxiliary Exhaust 9 Wrapped Cylinder 10 Damped Ram – 1″ Plate and EAR on Ram 11 Ram Cover – Armaplate Steel Strike Plate Micarta Strike Plate

#### Test No. 1

The Bare Hammer – Steel Cushion was a standard Vulcan Hammer, Size 1.

#### Test No. 2

During the second test. the exhaust noise us muffled by coupling a flexible hose (heavy duty rubber) to the hammer exhaust port and piping the noise about 40′ away behind an embankment. The hose terminated in a wooden box lined with fibreglass. Unfortunately the seal between the rubber hose and the exhaust discharge duct leaked and it was not until Test No. 10 – Damped Ram – that the leak was properly sealed.

#### Test No. 3

A single chamber plywood box stuffed with fiberglass was affixed to the Auxiliary Exhaust Port with an opening transverse to the normal auxiliary exhaust airflow.

#### Test No. 5

The base of the hammer was covered so that as much of the exterior radiating surface as possible was enclosed with 2″ of polyurethane foam. This was covered with a second layer of foam attached to a lead vinyl sheet manufactured by Ferro Composites. Norwalk, Conn. At the same time the Auxiliary Exhaust Muffler was rebuilt to contain a second chamber. Again, the plywood box was lined with flberglass and protected with netting and the gas flow exhausted to atmosphere through a side port transverse to the normal auxiliary exhaust air flow.

#### Test No. 9

The cylinder was treated in the same manner as the base.

#### Test No. 10

A sheet of rubber-like damping material (nominally 1/4″ thick) with the trade name EAR, manufactured by the National Research Company, a Division of Cabot Industries Cambridge. Mass., was clamped between the outer surfaces of the ram and a 1″ steel plate. The plate was attached to the ram with nuts and lock washers on previously installed ram studs. Unfortunately, due to the non-uniformity of the ram castings. only a small percentage of the radiating area was actually damped even after machining some surfaces of the ram.

#### Test No. 11

A cover comprising a steel plate bonded to rubber (about 1/4″ thick) with the trade name ARMAPLATE, manufactured by Goodyear, was fashioned to cover the entire exterior part of the ram. Tests were run with a) a Steel Strike Plate and b) a Micarta Strike Plate.

## B. METHODS OF ANALYSIS

### 1. instrumentation for Impact Analysis

Three different types of instruments were compared to evaluate the Peak Level of impact. Since this part of the noise cycle is a sharp transient the response characteristics of instruments are bound to give different values.

#### a. General Radio Impact Analyser

This Is an instrument which is designed to retain the Peak Sound Level on a meter so that the meter can he read easily. Though it has several different settings such as Time Average, Peak and Quasi Peak, we only tested the Peak.

#### b. Dumont Oscilloscope with a compressed time base

The vertical deflection of the scope trace was read from a persistent screen with a recticle which had been calibrated previously. With the exception of the inaccuracy in the visual readout the scope has the fastest and most accurate response of all the instruments.

#### c. Sanborn Graphic Level Recorder

The electronic circuitry controlling the writing pen ballistics was redesigned to give a nominal 0.005 seconds rise time. In any such display device with transient stimuli, the pen tends to over ride the actual peak level due to inertia. For the analysis used in this report, an actual writing speed of 6000 millimetres/sec. without overshoot was obtained. By moving the paper fast enough under the deflecting pen, a very clear display of the noise amplitude variation (Sound Pressure level in dB vs. time) is obtained. A comparison of all three methods indicated similar spectra. Though the Peak Impact levels from the Sanborn charts were lower (between 3 and 9 dB) than the Impact Analyser and Scope, it appeared to be consistent. The Sanborn Graphic Level Recorder has the very important advantage that it shows all the detail in the complete noise cycle while the Impact Analyser and the Scope show only Peak Impact level. This allowed us to make a detailed analysis of the various parts of the cycle affecting the total noise emanations. It was decided that the knowledge of the detailed parts of the cycle was more important than the absolute value of the Peak Impact level. Correction factors were added to all Band Pressure Levels affected by the high writing speed circuitry to provide a flat response. Accordingly. the method of analysis used for tape playbacks on tile Sanborn Graphic level Recorder

### 2. Sanborn Charts

#### a. Noise Signatures

To determine the uniformity of the noise spectra and the reliability of using a single cycle for analysis, copies of the charts showing the variation in Noise Signatures were made for linear, dB(A), 63 Hz and 2000 Hz. These data will be found in Volume III of this report. To obtain the dB Level on the ordinate (vertical direction) for any of the charts, reference should be made to the digital information presented for the Peak Impact Level in dB re 20 u N/m&Mac178; given in Tables IA thru F in Volume II. The Noise Signature charts are identified by a GR73 No. in the upper right-hand corner of each page. Turning to the tabulated digital data, the Noise Signature identifying numbers will be found under the Column “Graph 73-” in the right-hand column of each table. The noise levels in dB are listed under the appropriate column for each of the Octave Bands, Lin and dB(A).

To obtain the Sound Pressure Level of any portion of any cycle, relate the digital data to the Impact Peak. For example, GR73-024 is the analogue output for three successive cycles of the Bare hammer – Test 1, for the Microphone A located at the Base of the Hammer with a steel cushion. The synchronous readout on four channels, top to bottom, is for –

1. LIN short for Linear Weighting Network: includes all frequencies with a flat frequency response for engineering purposes and some contract specifications.
2. dB(A) short for “A” Weighting Network: it weighs the frequency response of the spectra closely with human auditory response and is frequently used in legislation.
3. 63Hz short for 63 Hertz Octave Band Pressure level: the lower practical band for noise evaluation both for engineering and subjective purposes.
4. 2000 Hz same as above except the Octave Band Center frequency is 2000 Hertz instead of 63 Hz.

Thus, in studying these charts. one can see

1. how repetitive the noise signatures are, and
2. roughly whether the peaks comprise low frequencies (costly to fix) or high frequencies (easier to reduce).

The smallest divisions are 1 dB for the ordinate and 0.02 seconds for the abscissa. The heavy lines are 5 dB and 0.1 seconds, respectively. A “blip” at one second intervals is given at the bottom or 5th trace on each chart.

On certain charts the 4th channel is the “Ticker” signal from the trip switch installed to insure identification of the ram position. It replaces the 2000Hz trace. Another variation of chart (GR73-165 thru 179), the dbA(A) channel has moved from channel 2 to 4.

For discussion or the noise signatures, we have arbitrarily divided the cycle of GR73-024 as follows:

 Seconds Identification 0-0.14 Impact – the vertical rise of the trace indicating the instant of impact is used as the time reference. 0.14-0.48 Rise – the period of lifting the ram. 0.48-0.64 Exhaust – that part of the cycle normally dominated by the exhaust. 0-64-1.06 Fall – the remainder of the cycle before the next impact.

Obviously, the time periods will vary slightly with the variation of pressure and ram weight (especially with the EAR, Test 10). but the four basic parts of the cycle will be useful for reference.

To determine the sound pressure level of any part of the analogue trace, say the exhaust peak of 024 LIN, refer to the digital tables in Volume II for the appropriate reference Line GR73-024, find the Peak Impact Sound Pressure Level = 137 dB. On GR73-024 in Volume III, we have written in these values above the peaks. Again. referring to the LIN trace, note the dB level reduces after impact to 113 dB at the beginning of the Rise period and reaches as low as 103 dB before Exhaust. The Peak Exhaust of this cycle is about 123.5 or 124 dB. The next Exhaust Peak is 123 dB.

#### b.Noise Cycle Spectra

The Noise Cycle Spectra are a composite of the Linear, db(A) and eight Octave Bands from 63 Hz – 8000 Hz Noise Signatures for the same hammer noise cycle. These are published for each test and source microphone location as well as 50′ East. The number above the impact part of the cycle is the Sound Pressure level for the Octave Band or Weighting Network at Peak Impact. It is from these spectra that the digital data were obtained. For any part of the noise cycle, the Sound Pressure level in dB re 20 u N/m&Mac178; can be determined referencing the Sanborn Chart grid system and file units described above. Since the subjective reaction to the hammer noise Is a function of frequency, it is important to know how the noise varies with frequency. This spectral information also is important as to the control measures that are required to reduce the noise.

### 3. Digital Data

When there is a large body of steady state analysis to be done, it is most efficiently done with a Real Time Analyser. In the case of the many events during the pile hammer noise cycle, the Real Time Analyser falls short. The analyser needs to be started at the precise instant the event of interest occurs. The synchronization problem associated with the event time and the analyser starting time is, at present, unsolved. If the hammer cycle was constant to within a fraction of the impact duration, one could then program a computer to start the analyser at the instant desired. Of course, the hammer cycles are not that constant, particularly during a test program involving changes to the ram weight. For this reason, the digital data, determined in this study, have all been obtained manually from the analogue readouts on the Sanborn charts.

### 4. Statistical Distribution Analysis

Though major events such as ram impact, exhaust and impacts from the valve tripping mechanism are discernible from the noise signature, the affect of different hammer configurations is not readily determined for the entire noise signature. To assess the changes in the total signature, Statistical Distribution Analyses were made for certain microphone locations and hammer configurations.

These data are determined from an inspection of the noise signature at each 50th/sec interval. If the cycles are dissimilar, more than one cycle is analysed. The results are then tabulated for each level and the cumulative percentages are determined. The data are plotted on probability graph paper to show the deviation from Gaussian distribution. From each individual random source. the curve should be a straight line when plotted on the probability paper. Thus, sharp discontinuities in the curvature of the Statistical distribution indicate the dominance of another type of random source. For the most part, the curves show few straight line portions other than ambient. We interpret this to mean that several of the sources are intermingled in the noise signature at any one time.

The data are very useful because, at any individual percentage of the time, the effectiveness of each hammer configuration can be assessed by merely subtracting the differences between the Statistical Distribution Levels.

One will notice on each Statistical Distribution plot that there are two different abscissas. At the bottom of the graph, the percentages are Indicated as the “% Time Noise is Lower Than the Indicated Level”. At the top of the graph, the reverse percentage in plotted as the “‘% Time Noise is Higher Than the Indicated Level”. It is the latter scale which is finding increasing acceptance by legislative bodies and standards organizations.

These Sound Pressure Levels in dB are simply referred to as L followed by the percentage. For example, “L40” means that 40% of the time the noise is above the level quoted for “L40”.

## C. RESULTS OF TESTS

The results of file tests are contained in Volumes II and III as follows:

 VOLUME II Section E UAC Drawing No. DWG. NO. 730V2 Section E is actually a graphical summary of file tests which are on a drawing labelled “Peak Impact”. The drawing is folded and tucked in the pocket in the front of Volume II’s binder. This summary shows the variation of the Linear Sound Pressure Level as a function of time for each test and each microphone location. These are representative cycles taken from the noise signatures. This drawing affords an overview of the entire test. F Tables IA thru F G Noise Cycle Spectra – Graphs No. 188 thru 277 H Statistical Distribution Analyses – Tables II, III, IV Graphs No. 398 thru 402e, 435 thru 475 VOLUME III Noise Signatures Graphs No- 024 thru 179B

For access to any particular data. use the GR73 number cross referenced in the Data Index.

## D. ANALYSIS OF DATA

### a. Noise Signatures and Cycle Spectra

#### a. Test 1 – BARE

At each of the microphone positions. there are two dominant noise sources and their Peak Sound Levels are listed below:

 Mike Impact Exhaust 1- Base 138 124 2 – Ram 136 126 3 – Exhaust 134 144 4 – Trip 136 125 5 – Top Cyl. 132 124 6 – Bot. Cyl. 133 119 7 – Aux. 134 118 25′ 112 106 50′ 104 101

Except when the mike is at the Exhaust the Impact noise dominates.

As the mike is moved away from the base, the level is lower except at the Trip and the Auxiliary Exhaust. It may well be that the piston at impact transmits to the cylinder wall and certainly the columns do. The Auxiliary Exhaust ports open directly to the interior of the cylinder which acts like a reverberation chamber. These differences are minor but do show clues to the sources. Generally, the whole hammer radiates during Impact as one would expect with such rigid connections between the supporting structure. The levels range around 135 dB and project to 25′ at 112 dB and 104 dB at 50′. It is recognized that the area is reverberant around the Test Stand and, if anything, levels at 25′ and 50′ are probably high. From this data, we calculate an equivalent spherical radius equal to roughly 1.5′. Calculated radius is 2.9′. Obviously. the whole hammer is not radiating or it has directional characteristics.

Certainly the Exhaust appears to be directional as one might expect from examination of the above table. Moving from Position 3 to Position 2, a distance of only 4′, changed the Impact noise by only2 dB while the Exhaust noise changed 18 dB. Measuring perpendicular to the Exhaust axis at distances of 25′ and 50′, the levels drop by 38 and 43 dB, respectively, showing the source radius to be only 0.315′ in contrast to the estimated physical radius of 0.47′. Obviously. the Exhaust is not only small but directional. The Exhaust source, though small in size and directional, is a very potent 144 dB or 6 dB higher than the apparent peak Impact noise.

See Table A for the estimated Equivalent Radiating Area Spherical Radii of other hammer parts.

The Exhaust directionality is partially attested, as the far field mikes are moved to the Southeast Positions (more in line with the Exhaust). Though the increase is a modest 2 dB for 50′, it is a drop of 2 dB for the 25′ mike distance. The reason for this is not clear except for the less reverberant condition (no direct reflections – See Vulcan drawing P168). Note that the Trip Mike Position 4 is only 3′ away but the level is lower than the Ram at 4.5′. This shows how easy the shadowing of the valve chest can hide the dominantly short wave lengths of the Exhaust which peaks at 2000 Hertz as shown in the Noise Cycle Spectra GR73-190. All of these factors must be considered in designing the control measures.

 VULCAN HAMMER NOISE SOURCE RADIATING AREAS Total Area = 104.8 Square Feet Source % of Total Mech. Area Square Feet Equivalent Spherical Radius Feet A. Mechanical 1. Ram 32.2 33.7 1.63 2. Cylinder 26.5 27.8 1.49 3. Base 17.3 18.1 1.20 4. Columns 14.3 15.0 1.10 5. Valve Chest 3.4 3.6 0.54 6. Sheave 3.3 3.5 0.53 7. Piston Rod 3.0 3.1 0.50 100.0 104.8 B. Aerodynamic 1. Exhaust 2.74 0.47 2. Auxiliary Exhaust 4.74 0.61

#### b. Other tests

Other observations of interest are:

1. The trip noise shows up before and after the Exhaust.
2. The Bottom Rear Cylinder radiates a similar shaped bulge after impact as the Auxiliary Exhaust.
3. An unexplained increasing level just prior to Impact as seen principally at the Exhaust which disappears in the Damped Ram, Test 10 – possibly piston rod seal leakage.
4. Exhaust Port seal leakage clearly disappears in Test 10.
5. The Nylon Slide Bar allows a possibly puzzling increase in its impact noise over the steel bar. Subjectively, the steel bar has a ringing that the Nylon Slide Bar lacks. The data, however, do not bear this out. Though the presumed Exhaust leak between the Trip Impacts disappears, it does so when the cylinder is wrapped in Test 9 instead of Test 10 when the leak was fixed. The inconsistency of this signature may well be due to the flexibility of the Nylon. particularly after the Cylinder wrapping.
6. One of the most dramatic changes in the Impact noise occurs when the Peak level drops from 134 dB to 123 dB at the Top Cylinder after it was wrapped.
7. Most of the change in the Peak Impact level was very gradual and actually negative at times. It shows the difficulties attendant to a noise reduction program of this type. When the overall reduction is dependent on a lot of small changes, much control must be exercised over the test technique and accuracy.
8. It is quite obvious in studying the 25 and 50′ signature that much of the significant source noise sort of “washes out” as one gets further from the hammer. This is particularly true of high frequencies and one should be cautious in expecting this to happen with low frequencies. If the low frequencies are high enough in level, they will still control the measurement even though the hammer may seem a lot quieter.
9. By similar observation and calculation. one comes to the conclusion that the principal sources are listed below in approximate order of importance.
1. Ram
2. Cylinder
3. Exhaust
4. Auxiliary Exhaust
5. Trip
6. Base
7. Columns
8. Valve Chest
9. Sheave
10. Piston Rod

### 2. Statistical Distribution Analysis

Though the Peak Impact Level dominates the undetailed information of the enforcer’s meter, it is not a significant measure of the quietness quotient. For example. the difference in Peak Impact Levels between Tests 1 and 10 is only 1 dB and 4 dB at 25′ and 50′, respectively. Obviously, some other measure is required to evaluate the substantial reduction in Loudness observed by the listener. Statistical Distribution Analysis helps fill this need since it measures the entire signature every 0.02 seconds.

#### a. Linear Probability Graphs

From the Probability Graphs. data has been summarized in Tables II, III and IV showing the differences between incremental changes in hammer configuration an well as the overall. These overall differences have been plotted as Probability Graph Noise Reductions, GR73-447 thru 453. showing the maximum reduction from Test 1 to 10 as 32 dB at L20 and the Auxiliary Exhaust Position. The maximum source noise reductions are tabulated below:

 Mike Percentile dB Reduction 1 – Base L50-20 16 2 – Ram L50 18 3 – Exhaust L30 25 4 – Trip L40 21 5 – Top Cyl. L30 27 6 – Bot. Cyl. L40 27 7 – Aux. Ex. L20 32

#### b. Histograms

Comparative Noise Reductions are shown at the L40, L30, L20 and L3 Percentiles for each change in hammer configuration at 50′ East. The pattern clearly speaks well for the Damped Ram, though we must caution that the Exhaust leak was also fixed during this test. Still it shows as the best at most of the Percentiles and we have noted before that only a small percentage of the damping plate was actually constraining the EAR to the Ram. Even so, a substantial reduction was obtained. This, coupled with the fact that the Ram has the largest Radiating Area, makes it the number one control measure to be applied in the hammer redesign.

In Graph GR73-454, the overall differences are given for each test change while Graph GR73-455 more clearly shows the contribution of each change by itself.

#### c. Octave Probability

From these graphs. the data in Table V has been derived which in turn has lead to the Statistical Spectra Summary graphs.

#### d. Statistical Spectra Summary Graph

Finally, the “proof of the pudding” is in the actual reductions achieved at likely observer distances. Graphs 472 thru 475 show the overall reduction in noise at a distance of 50′ for the L40, L30, L20 and L3 Percentiles. Levels of Noise Reduction for the annoying frequencies of 500 Hz and above ranged from 14 – 20 dB for all except the brief Percentile of L3 where the Noise Reduction ranged from 4 to 12 dB.

With the above evidence. it is proven that a concerted design and evaluation effort will pay substantial dividends in quieter pile hammers. Recommendations for appropriate control measures are given in Volume I of this report.

## INSTRUMENTATION LIST

### Field

• Microphones
B&K 4135 S/N 125108
B&K 413S S/N 125107
AKG C60 S/N 382
GR 1560-P40
• Power Supplies
B&K 280L S/N 144144, 144736
AKG N60EA S/N 148
GR 1560-4100
CR Octave Band Analyser 1558AP, S/N 203
• Attenuators
UAC S/N’s 60, 61, 62
• Recorders
Ampex 351 S/N 4042057
Preamp Channel A S/N 27688
Preamp Channel S S/N 28677
Ampex 351 S/N 0140176
Preamp Channel A S/N 34053
Preamp Channel B S/N 29042
Sanborn Strip Chart Recorder 954B-1OO
Log Preamps 350-1400
Channel A S/N 922, Channel B S/N 928, Channel C S/N 867, Channel D S/N 625
• Other Equipment
Hewlett Packard Scope 120B S/N 601-06957
B&K Octave Filter Set
KLH Speaker Model 6 S/N 35950
B&K Microphone cable set
Taylor weather Station

### LABORATORY

• Ampex Playback Deck 351A S/N 4840140
• Marantz Power Amp S/N 5042
• B&K 1612 1/3 Octave Band filter
• B&K Octave Filters
• Sanborn Strip Chart Recorder 954B-100
Log Preamps 350-1400
Channel A S/N 922, Channel B S/N 928, Channel C S/N 867, Channel D S/N 625
• Dumont Scope 401B S/N 160
• GR Octave Band Analyser 1558AP S/N 203
GR Impact Noise Analyser 1556A S/N 767
• Dynaco Monitor Preamp
• CM labs Power Amp S/N 0266
• KLH Speakers Model 6 S/N 35950, 35839

Leaders is the generic term for the guide which allows the pile hammer to be positioned on top of the pile and then started to drive the pile to its desired head elevation. Although not a prominent part of Vulcan product line, the company did produce leaders of many kinds. This page is also intended to give an overview of pile hammer leaders in general.

Leaders can be broadly categorised in two ways: the method by which they interface with the hammer and the way by which they are connected to (and guided by) the crane.

## Hammer Interface

Pile hammer leaders basically guide the hammer in one of two ways: from the side or from the back.

Vulcan hammers–in common with most American pile hammers–were guided from both sides of the hammer, with both hammer and pile between the guides. Leaders such as this are referred to as “U-type” leaders. Most of the Vulcan hammers shown on this site are riding in U-type leaders, both onshore and offshore.

European practice prefers to guide the hammer from the back. This is generally referred to as a “spud” type leader. The spuds have both a structure for stiffness and rails to interface with the hammer. These rails can be of several types, including round rails (Delmag, Nilens) or two channels back to back (Russian.)

## Crane Interface

There are several methods of interfacing the leaders with the crane, depending upon the nature of the job and the preference of the contractor:

The most common type of leaders, these are simply suspended from the crane (at some distance.) For most plumb pile applications, they are suitable.

Underhung leaders are attached to the boom point, either rigidly or through a cable.  Underhung leaders at one time were popular, but have been largely displaced by fixed leaders.

Fixed leaders are connected to the crane at two points: the boom point, generally using a swivelling connector, and at the base of a crane using a spotter.  This setup allows the maximum manoeverability of the pile hammer, which is especially important for complex batter piles.  It is also generally more efficient in moving from one pile to the other.  The concepts and examples of the various features of fixed leaders are shown in the photographs below.

These leaders are a shortened version of swinging leaders. They are intended for use with a template, which both holds the pile in place and sets the batter angle of the pile as well. Virtually all of Vulcan hammers used offshore were run in these leaders.

## Vulcan Expanding Mandrel, and Mandrel Driven Shell Piles

From a corporate standpoint, the Vulcan Expanding Mandrel (right) was one of its more forgettable products. But the application of mandrel-driven shell piles is an important one in deep foundations, as it represents, from a design standpoint, an interesting combination of driven and drilled piles.

### Pile Shells and Mandrels

Driven piles can be divided into two types: low displacement piles (such as H-beams and to some extent open-ended pipe piles) and high displacement piles (wood piles, concrete piles and closed-ended pipe piles.) Which one you use depends upon the application and the geotechinical environment you’re working in.

Advocates of drilled piles, such as drilled shafts and auger-cast piles, note that, once you’ve drilled the hole into the ground (or while you’re drilling,) you can fill the hole with reinforcing steel and concrete and have a deep foundation. The main weakness to that approach is that, in many soils, the soil will either completely collapse into the hole or contaminate the concrete during the pour, thus compromising the integrity of the foundation.

Closed-ended pipe piles eliminate this by providing a barrier between the concrete (which the pile is filled with after driving) and the soil, both along the shaft and at the toe of the pile. They also provide whatever advantage there is in displacing the soil during driving. However, to prevent collapse during driving, a minimum wall thickness is required for driving which is beyond what is necessary for the structural integrity of the foundation.

So what if the wall thickness could be reduced, saving the expense of steel unnecessary to the foundation? The answer to that “what if” is mandrel-driven piles. By using a heavy mandrel which is inserted into the pile before installation and driving the mandrel, the impact force is transmitted along the pile shaft, thus reducing the driving stresses the pile experiences (as opposed to transmitted all of them through the pile head.) Getting a mandrel that can survive the rigours of driving has been one of the greatest challenges of driven piles.

The most elaborate system of shell piles (as opposed to regular pipe piles) and mandrels developed was the Raymond Step-Taper system, with its matched system of pile configurations and mandrel cores. This was used successfully for many years, but is a very specialised operation. (Another solution to driving Step-Taper Piles was a hammer which actually was inserted into the pile, as described in the Vulcanhammer.info Guide to Pile Driving Equipment.)

A more generic solution was to use corrugated steel pile, similar to the corrugated pipe used in storm water drainage. Thin walled, reasonably rugged in handing and economical, corrugated shell pile is probably the most common type of mandrel-driven pile in use. The shell is lifted into place in the leaders, the mandrel is inserted into the pile and locked into place, the mandrel and pile are driven into the ground, the mandrel is removed and the shell pile filled with concrete. To prevent soil from plugging the shell, a boot is frequently used at the pile toe. Shell piles can also be installed on a batter (as shown above,) something that drilled piles have serious difficulty with.

An interesting side note is that it is necessary in many cases to cut off portions of the shell for proper pile length. In Third World countries, the cut-offs find their way into use as culverts in poor sections where the authorities have not seen fit to provide proper drainage.

More details on this can be found in Vulcan Bulletin 90, which you can download by clicking on the cover image to the right. This also shows a cross-section of Vulcan’s mandrel as well.

### Vulcan Expanding Mandrel

Vulcan’s Expanding Mandrel was designed by Clemens Hoppe of Hercules Concrete Pile, patented under U.S. Patent 2,977,770, and licensed to Vulcan. It was produced in two sizes, 12″ and 14″, which corresponded to the two sizes of shells it was intended to mate with. The driving head at the top of the mandrel was configured to mate with either a Vulcan #1 or Vulcan #0 series of hammers.

The mandrel used a system of cams to expand the mating surface of the mandrel to the corrugations of the shell. A manual cam lever at the top of the mandrel is used to expand the mandrel once it has been inserted into the shell. Once the pile is driven, in theory the lever could be used to radially collapse the mandrel and permit its extraction from the driven shell.

“In theory” is important because the Vulcan mandrel, in common with just about every other mandrel in use, was prone to become jammed due to the impact of driving. This brings up one aspect of mandrel use that is well known amongst those who do it: mandrel driving is some of the most difficult driving a pile driver can experience (offshore pile driving is probably the most difficult of all.) To start with, the mandrel is by definition a “high impedance” transmitter of force, which means that it presents a high apparent resistance to the hammer.

Beyond that, when the mandrel jams, the most common method of freeing it up is a delightful procedure called “bumping out,” where a sling is wound around the mandrel head to a beam above the ram. The ram is sent upward to impact the beam; like a pile extractor, the sling transmits the upward impact to the mandrel head and (hopefully) loosens the mandrel to allow its extraction. Raymond superintendents were especially adept at this, which helped earn them a reputation as hard on the equipment. It also inspired Raymond to develop the full-length column rods and later cables to hold its hammers together, something which Vulcan belatedly adopted.

The Vulcan Expanding Mandrel, manufactured primarily at the West Palm Beach facility, was mildly successful, but never dominant in the shell pile market. Although Vulcan’s distribution system was part of the problem, the Vulcan mandrel lacked the durability of the more popular Rusché and Guild Mandrels (something that the latter’s inventor, Charlie Guild, wasn’t shy about reiterating!) Both Guild and Fred Rusché were serious Christians; perhaps their experience in the shell pile installation business was an impetus for that!

The 1970’s saw the end of the production of the Vulcan Expanding Mandrel.

## Vulcanaire Supertherm, and the Airmizer Hammers

Energy conservation is an important consideration today in a world where the competition for energy sources is intensified by rising demand. But making best use of fuel isn’t new, and both the Vulcanaire Supertherm and the Airmizer hammers were Vulcan’s contribution to energy savings.

Both of these products were the original idea of Moses Hornstein, the owner of Horn Construction in Merrick, NY.  He was evidently focused on saving fuel and energy in the operation of his hammers.

### Vulcanaire Supertherm

The Supertherm was first demonstrated in late 1964. The Supertherm was simple in concept. As described in the unit’s field service manual (which has more information on the unit:)

The installation of the VULCANAIRE SUPERTHERM is intended to raise the volume of air produced by compressor through the expansion of air by the use of heat. To achieve this; exhaust gases, which are normally wasted, are diverted through the use of a diverter valve and transferred through a heat exchanger through which also passes air from the receiver on the compressor. The temperature of the air is maintained within certain limits by the use of an automatic control device to produce the greater volume of air to be used on equipment at an elevated temperature.

The objective for the contractor was to use a smaller compressor to power the same size of pile hammer.

Production of the unit was performed at the Special Products Division in West Palm Beach, as the Chattanooga facility lacked the fabrication capabilities necessary to produce the unit.

Although the unit worked as intended and performed well, the simple concept didn’t translate into a simple design. With numerous parts and complex fabrication and assembly, the unit was uneconomical to produce and difficult to install. Air compressor manufacturers learned how to use hot exhaust gases in other ways to improve the energy efficiency of their products.

By the late 1970’s, even with the elevated energy prices of the era, production of the Vulcanaire Supertherm had gone cold.

### Airmizer Hammer

Hornstein was not content with increasing the efficiency of his compressors: he and Vulcan commissioned the Austrian engineer John J. Kupka to develop the “Airmizer” hammers.  These were compound hammers similar to MKT’s “C” hammers, and used a similar cycle that James N. Warrington used with the California hammers.  Some photos of this hammer are shown below.

Expensive to build and complex in construction, the Airmizer hammers were less successful than the Supertherm, and the remaining inventory was scrapped in the late 1970’s.

## Vulcan 106: the “Switch-Hitter”

Note: the field service manual for the 106 can be found in the vulcanhammer.info Guide Volume 1.

Creating excitement in a “need-driven” type of equipment like pile driving equipment isn’t easy, especially one with as long of a history as Vulcan’s. Vulcan tried to do just that with the 106 hammer, a hammer which both technically and from a marketing standpoint came in with a great deal of promise but never quite lived up to it.

The introduction of the Vari-Cycle in the late 1960’s made it possible to change the energy of a Vulcan hammer without having to change the operating pressure.  Installing the Vari-Cycle in the #1 and #0 series hammers was problematic due to leader clearance issues, but another issue was that many specifications required a certain ram weight. Additionally, the desire was there to make some improvements to the design which, although successful, was certainly not perfect.

The result of this was the 106 hammer, the “Switch-Hitter,” complete with the baseball theme as shown in the literature cover at the right. Although it was certainly possible to change the ram in a #1 hammer to an 06, the idea here was to allow this to be done without disassembling the hammer.

Other than the removable weights, the biggest objective for this hammer was to remove the keys, which are the most persistent maintenance item on a Vulcan hammer.  That included the ram keys (the “Octo-Conic” system worked, but not the way it was designed,) the slide bar key, and the upper column keys, which were replaced with the column nuts.  The lower column keys were left on the hammer.  The 106 saw the introduction of the valve detent, designed to reduce valve flutter and the use of the ubiquitous door springs on the valve.

The 106 was introduced in 1971.  Vulcan was issued a patent on these innovations (U.S. Patent 3,566,977) but, as was all too often case with pile driving equipment, the innovations didn’t pan out as expected.

The biggest problem was in the removable weights.  Once installed, these would “belly out” and freeze in the space provided, and the “Switch-Hitter” would no longer switch.  So the central purpose of the hammer was defeated.

Beyond that, the column nut/key combination was soon to be overshadowed by the installation of tie cables on Vulcan hammers.  Already standard on Vulcan offshore hammers and Raymond hammers, the “cable through the column” arrangement was superior to the set-up on the 106.

The only innovation to be propagated from the 106 to the other hammers was the valve detent, which, although helpful, did not overcome thing such as hammer icing and the use of motor oil.

## Proposed Hammers During the 1960’s and 1970’s

We look at the hammers that began the change in Vulcan’s product direction during the late 1920’s and early 1930’s, and we also document the “last hammers” of the 1990’s. Here we look at those hammers which were proposed during the 1960’s and 1970’s that were never built.