Vulcan 08 Hammer: Specifications and Information

The Vulcan 08 was an upgrade from the venerable #0, increasing the ram weight from 7,500 lbs. to 8,000 lbs.  The 08 experienced the upgrades of the other Warrington-Vulcan hammers, including cables and the Vari-Cycle, as shown above.  (The cables shown above used the tapered lower fittings and are to the head; later Vulcan 08 hammers have button-type lower fittings and the cables run to just above the steam chest.)

Other configurations of the 08 can be seen below.

Specifications are shown below.

The 08 in action in Norfolk, VA:

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Vulcan #1 Hammer: Specifications and Information

The #1 is, in many ways, the “flagship” of the line.  Produced from the  beginning of the Warrington-Vulcan hammers, it was and is a popular hammer.  The last Chicago general arrangement is above; other general arrangements are below.

Various versions of the specifications for the #1 are shown below.

Other information about the #1:

Vulcan #2 Hammer: Specifications and Information

The #2 was one of the earliest “Warrington-Vulcan” single-acting hammers to be produced.  The first one was S/N 6, made for the Marquette, Houghton and Ontonagon Railroad Company.  The general arrangement for the hammer is above.

2-Steam-Pile-Hammer-1887
The first extant layout of the Vulcan #2 Hammer, dated 9 February 1887. It’s probably the first extant layout of the Warrington-Vulcan hammer. Until about World War I, it was common practice for Vulcan to lay out the general arrangement and then the shop produce the hammer from just that drawing. It’s an indication of both the skill and the decision making ability of those actually producing the product, and also probably of the involvement of those doing the design work.

Specification Sheets for this hammer are as follows:

Specifications Bulletin 68
Specifications, Vulcan Bulletin 68
Specifications Bulletin 68G
Specifications, Vulcan Bulletin 68G
Specifications Bulletin 68K
Specifications, Vulcan Bulletin 68K

The #2 was very popular for a long time; however, it ran into two difficulties from the 1960’s onwards.

The first was that pile specifications were calling for a larger hammer (such as the #1) and contractors found their #2 hammers sitting in the yard with little to do.

tn2ramThe second, from Vulcan’s standpoint, is that the production costs of the #2 were not much less than the #1.  When Vulcan produced a #2 for the State of Tennessee in the early 1990’s, it simply took the #1 frame and put a 3,000 lb. ram in it.  Since the stroke was longer, more energy was available.

Other information:

The Information Vulcan Requested for Driving Accessory Purchases

Below is a sheet showing the recommendations Vulcan made to its distributors in gathering information for driving accessories.

Sales Aid Information Sheet Driving Accessories

Driving accessory orders were (and are) generally custom orders; knowing this type of information made it easier for Vulcan and its distributors to fulfil their customers’ requirements.  This sheet dates from the mid-1960’s, after the executive office moved to West Palm Beach, FL.

pedestal-driving-head
A pedestal driving head, on the floor at the Chattanooga Facility.

Note: in this era pedestal driving heads (right) were a popular item.  Casting difficulties, however, made them increasingly difficult to manufacture, and by the early 1980’s Vulcan was out of the business of pedestal driving heads.  The replacement for these is two pieces, one to mate to the hammer and one to the cap, between them a piece of (usually) pipe.

 

 

Vulcan Hammers and the Gates Formula

For many years, Vulcan included Engineering News Formula charts and data in its literature.  Vulcan dropped the EN formula out of its literature in the 1970’s, for two reasons: the wave equation was in the ascendancy, and endorsement of the EN formula was an implied endorsement of the “bearing power” of the piles they drove, an endorsement which Vulcan was justifiably reluctant to make.

Nevertheless, the use of dynamic formulae persists for smaller projects and is embedded in many specifications.  For this purpose, the FHWA favours the Modified Gates Formula, and this is discussed in the latest edition of their Design and Construction of Driven Pile Foundations.  The section on the Modified Gates Formula is reproduced below:

FHWA Gates Formula Section

Gates Formula tables can be found for many Vulcan hammers can be found at the Vulcan Foundation website.

Copies of the FHWA Design and Construction of Driven Pile Foundations can be obtained by clicking on the cover images to the right.

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 Best Way to Celebrate Your 120th Birthday is With a New Slide Bar Part

On our Engineering at Vulcan page, we posted this general arrangement of the Vulcan #2 dated 1887.

2-Steam-Pile-Hammer-1887
The first extant layout of the Vulcan #2 Hammer, dated 9 February 1887. It’s probably the first extant layout of the Warrington-Vulcan hammer. Until about World War I, it was common practice for Vulcan to lay out the general arrangement and then the shop produce much of the hammer from just that drawing. It’s an indication of both the skill and the decision making ability of those actually producing the product, and also probably of the involvement of those doing the design work.

Little did we suspect that we’d need that drawing, but then these photos from Crofton Diving of Portsmouth, VA, arrived:

The hammer in question is Vulcan S/N 116, originally sold to the Florida East Coat Railroad (not far from the West Palm Beach facility) in 1897.  The distinctive “open” slide bar design was changed about that time to what is on virtually every Warrington-Vulcan and Super-Vulcan hammer made since.  Vulcan Foundation Equipment  was able to make the spare parts Crofton required from the original detail drawings.

“Planned obsolescence” wasn’t the Vulcan way in 1897 or afterwards, which is why a 120-year old product is still driving pile and being useful to the contractor.

DSCN2960
Crofton Diving at work: the Crofton I barge driving piles at a marina, Norfolk, VA, 2009.  The pile driving rig is using swinging leaders.

ZWAVE

ZWAVE was Vulcan’s foray into the wave equation program field. It was an outgrowth of research that dated back to the late 1970’s on the magnitude of impact forces of its hammers on pile tops, so as to estimate both the loads on the equipment and the stresses on the piles. The first tangible result of this was a method and computer program based on numerical methods applied to semi-infinite pile theory; this was presented at the Offshore Technology Conference in 1987.

It became clear, however, that such a solution would not be as comprehensive as necessary, so ZWAVE was developed. Developed for MS-DOS computers, it’s “Preliminary Trial Release” (beta version) was released in 1987. The two proper releases (1.0 and 1.1) were done in 1988, after which time there was some work done the program but it had no further releases.  (The user’s manual for 1.0 can be downloaded here.)

SS88POS2

Also in 1988 was the paper describing the program, “A New Type of Wave Equation Analysis Program,” presented at the Third International Conference on the Application of Stress-Wave Theory to Piles in Ottawa, Ontario, in May 1988. This paper is available in PDF format and can be downloaded by clicking the link below.

Click here to download “A New Type of Wave Equation Analysis Program”

Unfortunately ZWAVE’s copyright status makes it impossible to make the program available for download. The paper, however, shows many of the advanced features of the program which were both referenced by later authors and included in later wave equation programs.

SS88SEGS

Abstract for “A New Type of Wave Equation Analysis Program”

This paper describes a new wave equation analysis program called ZWAVE, which is a program specifically for external combustion hammers. The program is described in detail, the discussion dealing with topics concerning the program such as 1) the numerical method the program uses to integrate the wave equation, which is different from most other wave equation programs; 2) the modelling process of both cushioned and cushionless hammers; 3) the automated generation of mass and spring values for both hammer and pile; 4) the method of dealing with plastic cushions; 5) the use of a recently developed model for computing shaft resistance during driving; 6) the computation and generation of values based on basic soil properties such as shear modulus, Poisson’s Ratio and soil density; 7) the completely interactive method of feeding data to the program; 8) the method used to compute the anticipated rebound and the energy used to plastically deform the soil; and 9) the format of the interactive input of the program and the program’s output. Sample problems for the program, along with comparison of the program results with data gathered in the field, are presented.

Vulcan Vibratory Hammers and Vibratory Technology

By World War II, Vulcan’s air/steam hammer line dominated its production and revenue stream. Of all of the attempts Vulcan made to diversify is pile hammer line after that time, probably the most successful was its line of vibratory pile hammers.

Vibratory pile driving equipment represented a major departure for Vulcan, but it also represents an interesting technology in its own right. In addition to recounting Vulcan’s experience, we have a wide variety of items on vibratory technology in general:

Need a field service manual for your Vulcan vibratory hammer? Or other information. Much of that is contained in the Vulcanhammer.info Guide to Pile Driving Equipment, information about which is here.