Vulcan’s Blow Count Specifications

The durability and longevity of Vulcan pile hammer is something that is seldom replicated in just about any other manufactured product.  Since pile driving is self-destructive on the equipment, this is a remarkable achievement, but it should be tempered by the fact that it’s possible to render a Vulcan hammer inoperable by the way it’s used.  There are many things that can make this happen–inadequate or nonexistent hammer cushion material or lubrication to mention two–but the one thing that Vulcan decided to include in its warranty was the blow count specification.

Recording the blow count–the number of hammer blows per inch, foot or metre of pile advance–is virtually universal on pile driving jobs.  The dynamic formulae basically translated blow count into pile capacity.  While anyone familiar with pile dynamics understands that blow count is a crude measure of the response of a pile to impact, including a blow count specification is a good first measure of both the advance of the hammer and also how much energy is being returned to the hammer, which is a case of hammer damage.

BLOWCNT
Blow count-resistance graph, developed by Vulcan in the early 1990’s as part of its effort to educate state and federal agencies on the basics of pile driving. As the blow count increases, the amount of SRD (soil resistance to driving) increases, but at a progressively slower rate. This indicates that simply increasing the blow count is a “diminishing returns” proposition, destructive for hammer and pile alike.

High blow counts indicate that more and more of the energy was going back into the hammer rather than into the pile, thus increasing the danger of hammer damage.  They also indicate that pile top stresses increase with higher blow counts, as the movement of the pile to mitigate the maximum impact force decreases.  Thus high blow counts just to get the pile to tip elevation without considering changes in hammer or basic drivability considerations is a losing proposition.

Starting in the late 1970’s, Vulcan voided the warranty on its hammers if the blow count exceeded 120 blows/foot.  It’s interesting to note that Vulcan never made its specification in blows/inch.  This was true for its onshore hammers; however, for its offshore hammers it was forced by circumstance to increase the hammer refusal criterion as follows:

BLOW COUNT SPECIFICATION 1

Vulcan hammers are designed to withstand a continuous driving resistance of 120 blows/foot (400 blows/meter). In addition to this, Vulcan hammers will withstand refusal driving resistance of 300 blows/foot (1000 blows/meter) for five (5) consecutive feet (1500mm) of penetration. Any resistances experienced in excess of these are beyond rated capacity and will void the warranty. This definition is not an exclusive definition of excess of rated capacity and other criteria may apply.

1 Specification applies to all Vulcan offshore hammers, not just those listed in this catalog.

This was drawn from the API RP 2A specification, which was discussed relative to pile stick-up.  An elevated refusal blow count specification was justified by two things.  First, the offshore hammers were more robustly built than the Warrington-Vulcan hammers which made the company famous, as they were derived from the Super-Vulcan hammers.  Second, the remoteness of offshore job sites made high blow counts a necessity, as bringing a larger hammer to the job was frequently impractical.  (Improved methods of drivability predictability lessened the possibility of this happening.)

Blow count limiting warranty specifications are not an absolute method to prevent hammer abuse, but they’re a good start, and Vulcan used them to the advantage of itself, its end users and the owners of the projects where Vulcan hammers were used.

Advertisements

Driving Piles with Stub Leaders and a Template

The best known setup for pile driving equipment is a crane and a set of full-length (of the pile and hammer) leaders, attached to the crane in a variety of ways.  But another alternative is to use a “stub” leader, i.e., one that is very short, and a template to align, position and guide the pile.  This is traditionally associated with steel piling, so we’ll look at this first.

For these hammers the platform itself is the template, the piles are driven from the top through the legs.  Most conventional platforms had angled legs so the hammers almost invariably drove on a batter, which gave rise to the “stick-up” problem, more about that below.

But using stub leaders and a template isn’t restricted to steel piles; it has also been done on concrete piles, as can be seen below.

From a contractor’s standpoint, handing hammers in stub leaders requires a considerable level of skill from the crane operator, but the weight savings and ability to handle the hammer in difficult situations makes the use of stub leaders, when possible, a very attractive option.

Engineering Aspects of Stub Leaders

From the photos above, you can see that piles can be driven with stub leaders either plumb or on a batter.  Plumb piles are not much different with stub leaders than with conventional leaders: the key is to have the hammer straight and square on the pile, which means that the leader setup should be balanced to hang straight and side forces on the hammer be avoided.

With batter piles, since the offshore industry used them (and still does) intensively, the most complete specification for such piles is the American Petroleum Institute’s RP2A specification.  With stub leaders the pile basically supports the hammer during driving, and the hammer in turn loads the pile with both the impact loads and the static load of the hammer assembly, which in turn acts both parallel and perpendicular to the axis of the pile.  Basically there are two important engineering aspects to configuring driving batter piles with stub leaders on a template:

  1. Column buckling due to the weight of the hammer acting on the axis of the pile.
  2. Beam loading of the hammer due to the component of the weight which acts perpendicular to the axis of the pile.  This creates a cantilever beam with a maximum bending moment at the template.  Obviously the weight of the hammer assembly (along with the weight of the pile) will induce bending stresses.  These stresses are both tensile and compressive, and both are important to the structural integrity of the pile during driving.  The template must also be designed to handle the loads and moments on its structure.

With steel piling, the combined weights of hammer assembly and pile limit the permissible length of the “stick-up” of the pile.  Steel piles are easily spliced and added on to, so piles which are much longer than the stick-up can be drive.  (Piles which are much longer than practical lengths of conventional leaders can be driven as well.)  With concrete piles, these can be splices but there is less flexibility and less resistance to bending moment with splices, which limit the possibilities of driving these with stub leads on a batter.  (The ability or lack thereof of concrete to withstand bending stresses also complicates the situation.)

One more important point: the weight of the hammer assembly cannot generally be assumed to be at the pile head, but above it.  That’s why the center of gravity information is so important for offshore driving, which led to Vulcan tips such as this.

Stub leaders combined with templates is an attractive option for driving piles, but proper engineering and construction procedures must be followed for successful results.

 

Hong Kong and the Straits of Hormuz: It’s Amazing It Took This Long

Although Vulcan exported its pile driving equipment from the start, it was it’s foray into the offshore oil business that gave Vulcan a truly international perspective.  That perspective put some of the world’s “hot spots” into its field of interest, and two of them are very active these days: Hong Kong and the Straits of Hormuz.


Most of Vulcan’s activity in East Asia was in South East Asia; thus, its main “centre of focus” for its equipment and travelling personnel was Singapore.  With Vulcan’s sale of its first pile hammer package to the Petroleum Corporation of the People’s Republic of China in 1981, Hong Kong became of interest.  At the time China was a very closed country; Hong Kong acted as a window to the world, although from a commercial standpoint Vulcan didn’t use it that way.

The UK’s decision to return the entire colony to the People’s Republic when the lease on the “New Territories” (the area of Hong Kong north of Kowloon and excluding that and Hong Kong Island) expired in 1997 was formalised in 1984; however, rumours swirled about a handover years before.  The attitude of Vulcan’s business associates towards such a reintegration was bluntly summarised by one of them: “They’ll screw it up.”  The contrast between the state socialism of the People’s Republic and the free-wheeling capitalism of Hong Kong was pretty stark, and it was hard to imagine that the former would allow the latter to go on in the same way for any length of time.

Up to now the PRC has surprised many people with the relatively light hand they’ve actually had on Hong Kong.  Some of that was the desire of the PRC to have Hong Kong be a “model province” for the “capitalist roaders” in the rest of the country, an incentive for economic development.  Another factor was to make the reintegration of the greatest “wayward” region–Taiwan–more attractive to those on the island.  Still another was the PRC’s desire to maintain Hong Kong as an economic powerhouse and thus contribute to the country’s overall prosperity.

Such desires have butted up against two things: the linking of Hong Kong’s people of free expression to economic freedom, something the mainland has avoided, and recent changes in the Chinese leadership.  Now the latent conflict of the two is out in the open.  The Chinese leadership will have to tread carefully; if they don’t, they could fulfil my business associate’s prophecy and China will be the worse for it.


The Straits of Hormuz has been the central “choke point” of world oil shipments for many years.  The Persian Gulf is ringed by oil-rich nations and 20% of the world’s oil supply passes through it.  That vulnerability really came into public consciousness with the Yom Kippur War and the first “oil crisis” of 1973.  It wouldn’t take much to mine or otherwise sabotage the Straits of Hormuz, which increased the Western military interest in the place.

The countries that ring the Gulf have been aware of this vulnerability for a long time.  Saudi Arabia built its Yanbu oil terminal on the Red Sea in an attempt to provide an alternative to the Straits.  Vulcan’s first contact with and sale to the Korean contractor Hyundai was due to the fact that they were contracted to build this terminal and need pile driving equipment to accomplish it.  On the other side Iran was looking to build a major port at Chabahar on the Indian Ocean using Vulcan’s long-time customer Brown and Root, but the 1979 Revolution stopped that effort.  (The Islamic Republic built a port there, currently operated by India.)

With the Sunni-Shia divide and the ill-conceived war in Iraq (which deprived the two sides of a buffer) the Straits had opponents on both sides, and it was only a matter of time before it would become a hot spot once again.

The amazing thing in both these situations is not that they’re points of conflict, the amazing thing is that it has taken as long as it has to reach the current situation.

The Pile Buck Ads 1: Vulcan 3100 Assembled — vulcanhammer.net

This site has never had an “advertising budget” but in the last decade the publisher Pile Buck gave it the opportunity to advertise itself in its books Sheet Pile Design by Pile Buck and Pile Driving by Pile Buck. There were five in the series, and this is the first, using the assembly of the […]

via The Pile Buck Ads 1: Vulcan 3100 Assembled — vulcanhammer.net

From “Deal Yourself a Winner” to “A Pile Driver Talks About God”

The ad above is another Offshore Technology Conference ad from the early 1970’s.  It was aimed at its industry: the oilfield was well endowed with hard-drinking, card-playing people, a simple fact that doesn’t fit into some peoples’ idealisation of the past.  The onshore construction industry wasn’t much different, although the higher risks–and rewards–of the oilfield made everything more intense.

Contrast this with the following forty years later from Rusty Signor, then President of the Pile Driving Contractors Association and President of TX Pile, LLC, in an issue of Pile Driver:

In my last message, I ended with a different, more positive view on the news in our current world situation. This time, I am going to do another first: a book review. The book is Seven Men and The Secrets of Their Greatness by Eric Metaxas.

Certainly advice on engineering techniques, safety practices and legal tips are very important for our pile driving business; however, personal character development is also something to consider for most. You may or may not know of all the seven men in this book, but the ones you thought you knew are viewed from a very different standpoint than how you probably learned about them in school. The book focuses on their complete reliance on their spiritual calling. Since this is not a government publication, I can use the word God.

For instance, everyone knows about George Washington and the story of the cherry tree. However, did you know that he was a deeply religious man and that he relied on his faith in helping him make  decisions? He prayed on his knees several times a day with a Bible before him. Washington believed that God had a special purpose for his life and that providence saved him from being killed. In one battle alone, three horses were shot out from under him and he had bullet holes through his hat and clothing. He empowered his men with God-filled inspiration and they would follow him anywhere. I bet you never read that in grade school.

Another man mentioned is Jackie Robinson, who broke the color barrier in Major League Baseball. I recently watched the movie about his story, 42. Again, the movie didn’t really focus on Robinson’s critical reliance on his faith in God to be able put up with and finally put down all the Jim Crow nonsense. He had extraordinary athletic talent in basketball, football, baseball, tennis and track and field. Robinson also had a tendency for anger explosions dealing with racial injustices. His mother and preacher led to a deeper faith that controlled his anger and justice allowed would him only to be see won that with the restraint path to and love. The manager for the Brooklyn Dodgers was an extremely religious person who was looking for this sort of man: someone talented in baseball, but who also had a strong, Bible-based character. Everyone knows the rest of the story, but generally not the one centered on God.

In the business world, sometimes we get too caught up in our challenges with competition, problems with equipment, governmental codes, etc. We just need to stop and look up like these men did – to result in your success and happiness.

Vulcan 3100 Hammer: Specifications and Information

Like the 060 and even more the 040, the 3100 was a major step up for the company.  Even though it became the “gateway” to the company’s largest hammers, itself it was a dead end offshore for reasons that weren’t fully appreciated at the time, at least not by Vulcan or some of its end users.

The specifications:

The first 3100 was built for McDermott.  Even though the 560 had been introduced earlier and was lighter for the same energy, McDermott felt that the traditional “heavy ram-low striking velocity” approach was better, and also had the crane capacity to handle this size of hammer.  The hammer was ordered in the fall of 1973.

The road to completing the hammer was a rough one.  That fall was the occasion of the first oil shock, which was great news and bad news at the same time.  It was great news because the oil price spikes made the oil industry very active during that decade and early into the next one.  It was bad news because the demands on the supply chain of foundries and forge shops, coupled with the energy shortages that resulted from the oil shock itself, made lead times immensely long.  And, of course, patterns had to be built for all of the major castings.

The hammer was finally completed on 11 June 1975, but there was another twist: it was assembled on the deck of McDermott’s Derrick Barge 8 in Bayou Boeuf, Louisiana.  Vulcan traditionally preferred to ship their hammers assembled, but freight and delivery issues forced this method.  It was successful, not only making it simpler to ship the heavy hammer parts in pieces, but also to familiarize the end user’s personnel with the hammer itself.  By the 1990’s it became the standard method of delivery for hammers going to the Gulf of Mexico.

In spite of its difficult production road, the 3100 was successful from the beginning, with fewer of the “growing pains” that some of the earlier hammers had experienced.

As was the case with the 040, Vulcan used the hammer for advertising purposes, both then and many years later.

The general assembly is below (the hammer was so large, it required a two-sheet drawing.)

In spite of its success the 3100’s main claim to fame was to be the basis for the 5100.  Why was this so?

The first was obvious: the 560, virtually the same energy, was lighter and more economical to produce and operate.  The second was that, with offshore high-impedance steel piling, the higher impact velocity, problematic with concrete and wood piles, was actually preferable, albeit harder on the hammer.  The hammer never went much past its origin, in spite of the celebration that surrounded its inception.

Vulcan 040 and 340 Hammers: Specifications and Information

Vulcan’s personnel brought back many colourful stories from the field.  One of those came from Jesse Perry, Vulcan’s senior field service representative.  Offshore pile driving is a brutal, unforgiving business; offshore piles are tip elevation piles, and the expediency of “beating the pile to death” to get done in the high hourly barge rates was hard on hammers, especially those new in the product line.  One of those end users vented his frustration on Jesse, who responded by throwing his wallet on the table and telling the customer that he’d bet its contents that the hammer would work.

I never knew that Jesse ever lost his wallet in that way.

In a sense, however, Vulcan itself “threw its wallet on the table” with the 040 and 060 hammers; the 040, more than any other hammer, brought it in to the “big leagues” of offshore pile driving and, through its growing pains, made Vulcan the “stamp of quality offshore everywhere.

First, the basics: the 040 specifications.

The first 040 was sold to Ingram in August 1965; below are some photos from their barge.

Many other offshore construction concerns joined Ingram in using the 040, including McDermott, Dragados, DeLong, Santa Fe, Movible Offshore (soon Teledyne Movible Offshore,) Fluor, Brown & Root, AGIP, Creole Petroleum (now PDVSA,) and Humble Oil.

Offshore wasn’t the only place where the 040 could be found.  One of the most significant projects it was involved with was the long I-10 bridge across the Atchafalaya from Lafayette to Breaux Bridge, LA, built in 1969.

The 040 underwent many changes as it went along; early 040’s have many versions, as is evidenced by the general assemblies below.

Being the seminal hammer that it was, the 040 was useful for advertising, a usefulness that went past the Vulcan Iron Works itself.

340 Hammer

In 1972, with the introduction of the 560, Vulcan decided to rename the 040 the 340 hammer.  Vulcan also made some other important changes, such as moving to an iron (as opposed to a steel) ram.  The first 340 was delivered to McDermott in early 1973.  Specifications, a general arrangement and a photo are shown below.  It turned out to be the last hammer the Vulcan Iron Works produced, sold to PDVSA in 2000.

 

The Stamp of Quality…Everywhere

Vulcan frequently produced an ad for the Offshore Technology Conference.  Probably the best one was the “Stamp Ad.”  The “stamp of quality” theme had appeared in Vulcan’s literature for many years before that, but Vulcan’s graphic artist Carol Carr took it to a new level with this one in the early 1970’s.  It was unusual in many respects; it was in colour (colour wouldn’t become standard in Vulcan literature until late in the decade) and it was an 11″ x 17″ fold-out.

Snippets of Carol’s artwork have been on this site since its beginning in 2007, such as the masthead below:

It’s also in the current masthead as well.

The back of this ad is here:

Stamp-Ad-Back

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.

Daniel-Orifice-Diagram

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.

A33
A valve setting instruction from 1920. Note the “cast in place” oiler on the early hammers.

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.

p2-vs-CFM-Plot

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”:

Configuration A Valve Orientation
Configuration “A” is the valve orientation which allows the inlet fluid to pass around the “back side” of the valve and into the cylinder. For single-acting hammers, this is the lower side of the piston, and takes place during the upstroke. For differential-acting hammers, this is the upper side of the piston, and takes place during the top part of the upstroke and during the downstroke.
Configuration B Valve Position
Configuration “B” is the valve orientation which allows the inlet fluid to pass through the “slot” in the valve and out of the cylinder into the atmosphere. For single-acting hammers, this is the lower side of the piston, and takes place during the top of the upstroke and during the downstroke. For differential-acting hammers, this is the upper side of the piston, and takes place during the early part of the upstroke.
Configuration A Test Setup
The test setup for Configuration A.
Configuration B Test Setup
The test setup for Configuration B.
General Arrangement of Valve and Test Apparatus
Test Arrangement for Configuration A
Test Arrangement for Configuration B

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.