Soviet S-834 Impact-Vibration Hammer: Calculations, Part II

The introduction to this series is hereThe first installment of the calculations is here.

Calculations of Main Details (Strength

Strength calculations assume that the inertial forces during impact are 150 times those of the weight.

Rotor Shaft

We checked the rotor shaft strength in the optimal mode, i.e., when the impacting force direction formed a 90° angle with the direction of the blow. To simplify calculations consider that the forces act at one point. In the vertical place the shaft is loaded with impact inertia forces from the shaft weight and parts which are located on it.

where Q1 = inertial force from eccentric weight(s) and part of the shaft ahead of the eccentric.
Q2 = inertial force from the part of the shaft under the bearing.
Q3 = inertial force from the rotor weight and the middle part of the shaft.

A diagram of the shaft assembly is shown below.


A diagram of the beam forces in the vertical plane is shown below.


A diagram of the beam forces in the horizontal plane is shown below.


The forces which act on the shaft in the horizontal plane arise from the vibrating forces of the eccentrics.

The reactions in the vertical plane are

The reactions in the horizontal plane are

The bending moment in the vertical plane in section A-A is

In section C-C it is

In section B-B it is

The bending moment in the horizontal plane in Sections A-A and C-C is

and for Section B-B

The sum of bending moments in Section A-A is

In Section B-B they are

and in Section C-C they are

The bending tension is calculated in the same way at all points.

For Section A-A

For Section B-B,

and Section C-C,

The tension in this section will be much less because the calculations do not take into account the force from the rotor shaft. Calculation of the shaft deflection will be done in Part C.

The calculations consider that the shaft is of uniform diameter, equal to 62 mm. In the vertical plane the deflection is equal to

where kg-cm
= axial inertial moment of cross-section of the shaft

E = spring modulus of shaft material = 2,000,000 kg/cm²

The deflection in the horizontal plane is equal to

The total deflection from horizontal and vertical moments is

In reality deflections will be smaller because we did not take into account the rotor forces.

Determination of Tensions in Vibrator Casing

The casing is subjected to loading tensions when the vibrator impacts on the pile cap. As the ram is located in the centre of the casing the critical sections are two perpendicular sections which are located at the planes of symmetry of the vibrator.

Let us determine the moment of resistance of the section which is shown in the drawing of bending tensions in this section, shown below.


This section is weakened by a hole for the ram but this weakness is compensated for by the local boss. So we do not take into account the hole and its boss.

The moment of inertia for the section relative to axis X-X is determined as

where = sum of inertial moments of the separate elements.
= sum of multiplication of squared distances from the mass centre of element ot the axis X-X by the area of the element.

The moment of resistance for this section is

The distance between the axes of the electric motors is mm. So the bending moment is equal to

The bending tension is equal to

Let us determine the bending tensions in the section perpendicular to the axis of the rotors. The section is shown in the drawing below.


To simplify the calculations consider the section of the casing is symmetrical and consists of two circles and two rectangles.

The inertial moment is equal to

The moment of resistance equals to

Let us now determine the bending moment considering that the load from the weight along the axis parallel to the rotor axis is distributed uniformly.


The bending tension is equal to

Spring Deflection Calculation

The maximum force for which spring deflection is required is P = 1000 kgf. The number of spring N = 2. The maximum deformation of the springs is f = 200 mm. The load for each spring is

As the springs are operating in relatively easy (not hard) conditions we can consider the permissible tension equal to 5500 kgf/cm². So the permissible tension per 1 kgf of load is equal to

The necessary spring stiffness is equal to

So we choose the spring with the following specifications:

Average Diameter

Wire Diameter

Hardness of One (1) Turn

Number of Working Turns

Npad = 14.5

Total Number of Turns

N = 21.5

Tension per 1 kgf of Load

A = 11.18

Hardness of the whole spring

So the spring we have chosen meets all of the requirements.

Determination of the Geometrical Configuration of the Eccentrics

Consider that the balanced part of the eccentrics (I and II; see diagram below) cancel each other.


So the coordinate of the center of mass of the rest of the eccentric (in the shape of a sector of a circle) is determined by the equation

The weight of the unbalanced part of the eccentric for a 1 cm thickness is equal to 1.7 kg. The eccentric moment of this eccentric is

The dynamic force of the eccentric is

The angular speed is rad/sec. The necessary eccentric moment of the eccentric is

The necessary total thickness of the eccentrics is

As during the determination of the eccentric moment it was increased a little, consider the thickness of the eccentrics equal to 80 mm.

This configuration of the eccentrics which we have come up with gives us an increase of its weight in comparison with the weight which is necessary to provide the required eccentric moment. So decreasing the moment of the rotary parts makes it easy to operate the motors.

Sizing the Bearings

The rotor shafts are mounted to spherical, double-row roller bearings No 3614 which have a coefficient of workability C = 330,000. The rotor weight Gb = 25 kgf. The eccentric weight is Gg = 28 kgf.

For the calculation of dynamic loads consider that the accelerations during impact are equal to 150 times the free weight.

As the shaft is symmetrical, each bearing is subjected to half the dynamic load

The shaft rotates at n = 950 RPM. Consider a factor of safety Kd = 1.5 and a dynamic load coefficient Kk = 1. The durability of the bearing “h” is determined as

Therefore, for 950 RPM, h = 160 hours.



Soviet S-834 Impact-Vibration Hammer: Calculations, Part I

The introduction to this series is here.

Moscow, 1963

Head of the Vibrating Machine Department L. Petrunkin
Head of Vibration Machine Construction: I. Friedman
Compiler: V. Morgailo and Krakinovskii


The impact-vibration hammer is intended for driving heavy sheet piles up to 30 cm in diameter as well as concrete piles 25 cm square up to a depth of 6 m for bridge supports and foundations.



Power N, kW


Blows per Minute Z


Revolutions per Minute


Ram mass



Force F, kgf


Determination of Velocity and Energy per Blow

Impact velocity is determined:

where = fraction of natural frequency (without limiter) to force

i = fraction of the number of revolutions to the number of impacts
R’= coefficient of velocity recovering (assume R’=0.12)

In our case




Energy of blow is determined as

Power necessary to make impacts is

Impact-Vibration Hammer Springs

So that the impact-vibration hammer operates in the optimal mode while the gap is equal to zero, the spring suspension stiffness should meet the equation

where = stiffening coefficient = 1.1 to 1.3, assume 1.2

Stiffness Distribution and Maximum Deformations
of Upper and Lower Springs

The upper springs are necessary to provide positive gaps, so their stiffness should be minimal to provide undisplaced operation the springs in the whole range of gap adjustment. Therefore

where Cb = stiffness of the upper springs
A = number of vibrations of the ram

a = maximum positive gap when the hammer is able to operate without danger of transferring into the impactless mode. When there is no limiter it is equal to the amplitude of vibrations

Assume a = 0.8.

where = coefficient which depends upon i and R’. Hammer coefficient of
velocity recovery may be increased up to R’ = 0.2. In this case = 7.1.

For calculation purposes let us assume A = 5.5. Now substitute the values into the formula

The bottom spring stiffness is then equal to

Now let us determine the maximum deformations. For upper and lower springs,

where b = negative gap. It is considered equal to “a” (maximum positive gap)

Assume .

Because of design considerations use four (4) upper and four (4) lower springs. The stiffness of one upper spring is

and the stiffness of one lower spring is

The material for the spring is “60 Sg” steel. The permissible tension in this steel is kgf/cm².

Upper Springs

Tension per kgf of load is

According to the table of S.I. Lukowsky choose the spring as follows:

The stiffness of one turn and the number of working turns is

Assume turns. For this spring,. The actual tensions in the spring are as follows:

(Units should be kgf/

and the total number of turns is

The full free height of the spring is

The distance between the support surfaces while the gap is equal to zero is

Lower Springs

According to the table the closest value A = 4.24 corresponds to the spring with dimensions

The stiffness of one turn is equal to . The number of working turns is

Assume 10 turns.

The total number of turns is

The spring height in free position equals to


Soviet S-834 Impact-Vibration Hammer: Overview

With this we begin a series of posts on the S-834 impact-vibration hammer, which the VNIIstroidormash institute in Moscow designed and produced in the early 1960’s.  With the revived interest in Soviet and Russian technology, it’s a detailed look at how Soviet equipment designers came up with an equipment configuration.  But it’s also a close-up view of how heavy machinery in general and pile driving equipment in particular is designed.

The impact-vibration hammer was a long-time interest for Soviet construction machinery institutes from 1954 to 1970.  An overview of the history of this type of equipment in the Soviet Union is here.  Since vibratory pile driving equipment was first developed in the Soviet Union, it’s also interesting to look at the entire subject; that overview is here.

The series is in three parts:

General View of the S-834 Hammer

The specifications for the S-834 are here.  What follows is an overview of the hammer itself and its general construction.  We apologise for the poor quality of the scans.

A general view of the machine. The impacting ram (1) is driven by eccentric weights and a motor within, which both lift it and force it down to impact. The hammer frame (2) receives the pile from below through a centre hole, which makes it possible for (1) to impact the pile. The motion of (1) is governed by the upper and lower springs (3). The compression on those springs is adjusted by (4), (5) and (6).
A cutaway view of the impacting ram. Basically the centre shaft (3) is driven by the electric motor (2), which in turn rotates the eccentrics (9). The force is transmitted from the eccentrics to the body (1) via the bearings (4) and the bearing housings (5). Electrical power is fed to the motor at the electrical connections (12). Once the entire assembly reached the impact point, impact force is transmitted to the pile at the ram point (10).
The ram point’s force is transmitted through the anvil (5) to a wood cushion (1), which in turn transmits the force to the pile, whose head is inserted through the tapered receptacle (2). The size of the receptacle can be adjusted with (3). The leader guides (6) are used for the leaders, which are (in typical Soviet and European fashion) behind the hammer.
Another variation of the anvil assembly.
This shows how the pile is drawn up into the leaders. The pile is attached to the bottom of the frame using a sling. This was common practice in the Soviet Union and is also done elsewhere. The alternative is to use a separate pile line. If the equipment is configured properly, this can work well.

Design Calculations for the S-834

In the posts that follow, the design calculations for the S-834 will be presented.  In looking at the work of Soviet designers, it was tempting to revise the calculations.  For one thing, although the metric system was introduced with the Russian Revolution, their implementation of the system is not really the “SI” system taught today, especially with the use of the kilogram-force.  (That’s also true with many other Continental countries such as Germany and France.)  For another, Russian technical prose can be very cryptic.

In the end, it was decided to reproduce the calculations pretty much “as they are,” with a minimum of revision.  We apologise for the inconsistent sizing of the equations.  Most of the transcription of this information was done in the 1990’s in Microsoft Word, and its conversion to HTML (for this format) in LibreOffice made the equations graphics (a good thing) but inconsistently sized the images (a bad thing.)  This is one reason why we’ve migrated to LaTex for our newer technical productions online.

As with much of the Soviet material on vibration and impact-vibration pile driving, I am indebted to VNIIstroidormash’s L.V. Erofeev for the material itself and V.A. Nifontov for its translation.

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.

Back in the Saddle at the Deep Foundations Institute


Vulcan Iron Works was involved in its industry in a number of ways other than simply selling and renting its product.  One of these was its years in the Deep Foundations Institute.  Although Vulcan was not a charter member of the organisation, it joined very shortly after its beginning and was active during the 1980’s and early 1990’s, until about a year before the merger with Cari Capital.  This webmaster was the Program Chairman for the 1992 DFI Annual Meeting in New Orleans.

So it is with pleasure that I have joined the DFI once again, continuing another tradition of the “Old Vulcan.”  My thanks to Theresa Engler, DFI’s Executive Director, who helped make this a reality.

Engineering at Vulcan

Vulcan would have never endured as long as it did without a properly engineered product, especially in the punishing environment of impact pile driving equipment. There is a great deal of technical information on this site; here we give a glimpse as to how much of it came into being. There are special sections on CAD, Finite Element Analysis, and Numerical Analysis.


Vulcan’s reciprocating steam engines weren’t only used on pile driving rigs. In the same era Vulcan was also heavily involved in building dredges. We have a complete page on the subject; we’ll concentrate here on some of the design engineering aspects of those vessels.

From the beginning of the Warrington-Vulcan product line (and presumably earlier) until the 1950’s Vulcan drawings were largely drawn in India ink on linen. They were thus durable and reproduced well, and (as is evident here) have some artistic value. Unfortunately they were hard to change, but given the static quality of Vulcan’s product line that wasn’t as big of a disadvantage as one might think.

An example of Vulcan’s vertical integration: a pile hook, along with a shackle to go with it. Today components like this (especially the shackle) are usually purchased. On the right is a thoughtful “note to blacksmith.”





Towards the end of the 1950’s Vulcan made two important changes in the way it made drawings: it went to pencil drawings and it drew them on vellum, which was preprinted with standard borders and title blocks. Additionally, after the move to Chattanooga it mandated the use of lettering templates. All of these resulted in drawings that were easy to change and had a more uniform look about them, but did not have the visual appeal of their earlier counterparts. Also, the vellum tended to fray with repeated reproduction; Vulcan’s reproduction machines used a contact process with ammonia development, which made the office stink, especially with poor ventilation. This forced frequently used drawings to have to be redrawn frequently.




Computer Aided Design (CAD)

By the mid-1980’s CAD was becoming a viable option for companies the size of Vulcan. In 1986 Vulcan purchased its first personal computer (PC) for the purpose, but the original software was unworkable, so Vulcan purchased DesignCAD and began producing drawings by computer drafting. The first hammer to be principally designed by CAD was the Vulcan 1400 vibratory hammer.


Other examples of Vulcan’s CAD output can be found in our page on the “last hammers.

Finite Element Analysis (FEA)

While Vulcan’s competitors such as HBM trumpted their use of FEA for designing hammer components, Vulcan got its start in 1977 with the analysis of the 6250 pipe cap, which was being proposed to McDermott. The analyses were conducted by Dr. William Q. Gurley at the University of Tennessee at Chattanooga, who was later involved in this effort.

Vulcan went on to analyse the 6300 pipe cap when McDermott “upsized” the hammer. Vulcan also conducted analyses on other hammer size pipe caps and piston rods as well; the former led to lightening the pipe caps considerably.


Numerical Analysis

For most of its history Vulcan used “closed form” solutions to predict the cycle behaviour of its hammers. In the early 1980’s, however, Vulcan developed the capability to analyse a hammer cycle using numerical methods and flow prediction, including valve losses. The first hammer to use this type of analysis in the actual original design of the hammer was the Single-Compound Hammer, where the complexities of the flow made such an analysis almost mandatory.

The “indicator card” developed for the S/C hammer, using an HP-85 computer, 1982. The output was actually printed on thermal tape. The HP-85, with its Basic programming and VisiCalc spreadsheet, was a useful device for hammer design and trade union negotiations alike.


Vulcan’s most ambitious numerical analysis project was the ZWAVE wave equation program. That, in turn, was a prelude to projects beyond Vulcan, namely those of the closed form solution for the wave equation and the FEA solution of the wave equation, forward and inverse.

Vulcan Iron Works: The Company

3100-Wright-BeverlyVulcan had a long and interesting history. Some of that is documented below:

Need further information? Click here to contact us.

Vulcan Patents

Vulcan was an innovator in pile driving equipment for more than a century. This history can be documented in part by the patents that were issued to Vulcan’s people, in addition to those which it acquired externally. We also include patents that were related to Vulcan either because they were issued to a Warrington or they related to a Vulcan product but were never formally assigned to the company.

We also have experience in acting as an expert witness in patent disputes; contact us if you are interested.

Patents Assigned or Licensed to Vulcan Iron Works

Formal Patent Title Application Inventor(s) Patent Number (click on nation for patent office that applies)
Note: if the patent number is hyper linked, the patent is on our site and available
United States Canada United Kingdom France Germany Australia Japan
Steam Pile Driving Hammers Original Vulcan Hammer Patent James N. Warrington 378,745

DSCN0776Vulcan #2 Hammer, S/N 463, with the “New Style Patent” number cited on the cylinder. The number is actually in error; it should read 378,745, which is of course the first patent on the list.

Caps for Piles McDermid Base, for driving wood piles directly Hugh McDermid 613,385
Pile-Drivers Sheeting Base William H. Warrington 777,459
Head-Block Head Block for Leaders H.C. Lindsly S/N 808,665
Caps for Sheet-Piling “Splined” cap for proper alignment of sheet piling cap William H. Warrington 960,746
Pile Hammers California Hammers James N. Warrington 1,019,386
Pile Hammers California Hammers James N. Warrington 1,020,467
Pile Extractors and the Like Vulcan Pile Extractor James N. Warrington 1,736,104
Power Hammer Super Vulcan Differential Acting Hammer Campbell V. Adams 2,000,908
Power Hammer Super Vulcan Differential Acting Hammer Campbell V. Adams 2,004,180
Pile Driving Hammer Internal Combustion Hammer (IC-65) Campbell V. Adams 3,013,541 958,688 1,275,714 1,484,582 422,233
Power Hammer Differential Acting Hammer Campbell V. Adams 3,096,831
Pulling Adapter Clamp for Extractor Campbell V. Adams 3,149,851
Pile Driving Apparatus Servo-pneumatic hook block, sand drain hammers Henry G. Warrington 3,171,552
Mandrel for Driving Pile Shells Expanding Mandrel Henry G. Warrington 3,329,216
Power Hammer Vari-Cycle I Campbell V. Adams and Henry G. Warrington 3,357,315 820,641 1,136,470 1,533,275 1,634,655 639,221
Hydraulic Pile Hammer Hydraulic Hammer Campbell V. Adams S/N 665,525 853,983 1,244,635 1,748,677
Boring Apparatus with Valveless Impactor Rock Drilling Henry G. Warrington and George Manning 3,444,937
Pile Driving Hammer IPH-16 (Internal Pile Hammer) Henry G. Warrington 3,454,112 908,449 1,271,544 2,010,537 1,928,621 444,891
Percussion Hammer Slide Bar Gripper Campbell V. Adams 3,455,208 868,711 1,237,246 1,572,802 1,784,044 439,225 655,568
Piling Extractor Wood Pile Puller Wayne de Witt 3,534,996
Percussion Hammer OPH-80 (Ocean Pile Hammer) Henry G. Warrington 3,547,207 921,715 1,282,615 2,022,755 1,955,300 444,896
Percussion Hammer 106 Hammer George C. Wandell 3,566,977 438,549 882,047
Percussion Hammer DGH Auto-Stop George C. Wandell 3,645,342 937,834 1,342,798 2,095,760 450,595
Free Piston Power Source Linear Vibrator John J. Kupka 3,704,651 952,774 1,362,213 2,117,070 457,844 885,621
Percussion Hammer DGH Auto-Stop George C. Wandell 3,782,483
Cushion Pot with Mechanical and Molded Joint Key Ball Cushion Ring Henry G. Warrington 3,800,888 1,408,317 900,238
Pile Driving Hammer Vulcan 3150 and 4150 Offshore Hammers Henry G. Warrington 3,815,474
Percussion Hammer Hydra-Nut Henry G. Warrington 3,938,427
Cable Tensioning Assembly Hy-Ten Cable Tensioner Henry G. Warrington 4,015,821
Retaining Assembly Ram Key Retainer John A. Lerch 4,295,752
Vibratory Hammer/Extractor Vulcan 400 and 1400 Vibratory Hammers Don C. Warrington 4,819,740 1,299,366
Method and Apparatus for Breaking Reinforced Concrete Piles and for Exposing Reinforcing Bars Concrete Pile Cutter Pulat A. Abbasov, Valentin E. Abramov, Dmitri A. Trifonov-Yakovlev, Lev V. Erofeev, Gennady S. Kuritsyn, Alexandr P. Borodachev, Victor V. Matvienko, Yuri V. Dmitrevich, Ludmila P. Lukash, Alexandr S. Petrashen, and Valery B. Petrov 4,979,489 Patent Cooperation Treaty PCT/SU/88/00114
Sea Water Pile Hammer Sea Water Hammer Don C. Warrington, Vladimir A. Nifontov, Lev V. Erofeev and Dmitri A. Trifonov-Yakovlev 5,662,175 Patent Cooperation Treaty PCT/US/96/12831

Other Patents of Interest

Formal Patent Title Application Inventor(s) United States Patent Number
Steam Pile-Drivers First pile hammer manufactured by Vulcan Thomas T. Loomis 160,781
Caldwell Cyclone Snow Plough Snow-Clearing Equipment E.P. Caldwell 405,300 and 454,109
Lubricators Lubricator for snow-clearing equipment (Caldwell Cyclone Snow Plough) George Warrington 423,580
Timing Devices for Hydrocarbon Engines Automotive George Warrington and Chester H. Warrington 1,418,996
Mandrel for Driving Pile Shells Vulcan Expanding Mandrel Clemens Hoppe 2,977,990
Muffler for Pile Driving Apparatus Decelflo Muffler Stannard M. Potter 3,981,378
Underwater Hammer with Circumferential Flow Seal Bolt Associates/Raymond International Underwater “Air Gun” Hammer George Gendron and Nelson Holland 4,098,355
Method of Driving Piles Underwater “Thin” Underwater Hammer Concept George Gendron and Lindsey Phares 4,138,199
Piling Hammer Sermec Hydraulic Hammer Brian Hays and Clive Taylor 4,802,538



Conmaco 300 hammer at Vulcan’s facility in Chattanooga, where it was manufactured (see below.)

Vulcan’s success in its product lines inspired imitators.  One of the most significant of those came from its own distributor network—Conmaco.  The relationship between the two companies was as complicated of a business as one could want, both together and apart.

Conmaco was started in 1910 as Contractors Machinery Company.  Three years later Scott Myers, whose family came to become principals in the business, became associated with the company.  In 1938 Vulcan and Conmaco signed their first distributor’s agreement in which the Kansas City, Missouri based business became Vulcan’s dealer in western Missouri and the entire state of Kansas.

While Vulcan’s future principal was off fighting World War II, the companies had their first falling out.  In the spring of 1944 the Bureau of Yards and Docks sent out a bid to Conmaco for 254 drop hammers.  Vulcan was only capable of delivering 112 of these by the end of the year, so Conmaco proposed that it produce the balance of the order itself.  In good company tradition, Vulcan’s Secretary-Treasurer, Walter Daspit, flatly turned down Conmaco’s request.  (William H. Warrington had set the precedent for this with Raymond before World War I.)  On 20 May 1944, as the Allies celebrated taking the remains of the Monte Cassino monastery, Conmaco cancelled their distributor’s agreement with Vulcan, which Daspit accepted without any attempt at settlement.

The end of World War II also brought reconciliation between Vulcan and Conmaco.  On 7 September 1945, five days after the Japanese surrendered to the Allies on the deck of the Missouri, Vulcan formally reinstated their Missouri-based dealer.

As the country’s economy shifted from wartime to peacetime, Conmaco pursued the capabilities it first explored during the war by either producing its own drop hammers or reselling those purchased at surplus.  As early as 1949 Vulcan queried Conmaco on this matter.  Nevertheless in the mid-1950’s Conmaco was generally Vulcan’s most important distributor.  In September 1957 Vulcan and Conmaco (which had formally changed its name and moved to Kansas City, Kansas) signed a new distributor agreement.

Although Conmaco continued to be an important distributor and its territory expanded to other Midwestern states, the relationship progressively deteriorated during the first half of the 1960’s.  Part of the problem was that the nature of the market was changing.  Up until that time Vulcan had an extensive network of distributors which dealt in a wide variety of construction related products.  Pile driving equipment is a specialized product in a specialized market, and “general equipment” distributors (even in some cases crane dealers, for whom pile driving equipment was a natural adjunct product) were losing interest in the product.  Things were further complicated by the emergence of a rental market for the hammers.  Vulcan’s entire marketing strategy was based on hammer sales; rentals (which made sense for infrequent users and joint ventures) cut into that.  Conmaco responded to both of these trends by concentrating on the pile driving equipment (along with other related products such as winches and cranes) and developing a rental fleet.  This last was not to Vulcan’s taste, although it was a market trend that ultimately Vulcan could do nothing about (and having as durable of a product as it had only accelerated the trend.)

Additionally, Conmaco continued to develop its own production capabilities of driving accessories, leaders and producing Vulcan hammer parts.  This basically turned Conmaco into a competitor to Vulcan.

Things came to a head in January 1965 when Vulcan cancelled its distributor’s agreement with Conmaco.  The situation was further complicated by the fact that Conmaco had hired Vulcan’s Vice President of Sales, Earle R. Evans, in February 1965.  (Evans had been involved in the coming of Prince Alexandre de Rethy of Belgium to Palm Beach the previous month.)  In spite of all of this the two companies remained in dialogue, so Vulcan continued to treat Conmaco as a distributor; however, there was no formal agreement until early the following year.

The two organisations spent the rest of the decade trying to maintain a relationship.  At one point Conmaco even had their Chicago branch at Vulcan’s old facility at 327 North Bell.  In 1967 Vulcan drew up and produced to Conmaco’s specifications the Conmaco 200 hammer, a counterpart to Vulcan’s 020.  There were many detail changes, but the most significant was that the hammer was tied together with a cable wrap system designed by Dwayne Smith, Conmaco’s Equipment Manager.

Vulcan’s General Arrangement for the Conmaco 300 hammer, similar to the 200 except for the ram weight.  This hammer became the basis for the popular Conmaco 300E5 hammer.

The cable wrap system was a mixed business.  One the one hand, the wrap system is labour-intensive and requires the hammer to have larger jaws (and thus require larger leaders) than the system that Raymond developed and Vulcan adopted for its own hammers.  On the other hand the larger jaws made it simpler to accommodate larger piling, which is especially advantageous with hammers in the 300E5 range. Additionally the smaller Conmaco hammers sported cables of kind before their Vulcan counterparts did.

Vulcan went on to produce the Conmaco 140 and 160 hammers (counterparts to Vulcan’s 014 and 016 sizes.)  But time was running out on the tempestuous relationship; the fundamental difficulties which led to the 1965 parting could not be resolved.  In November 1971 Vulcan cancelled Conmaco as a Vulcan distributor because, as Vulcan’s President H.G. Warrington noted, “the general business objectives of Vulcan and Conmaco are basically inimical.”

In spite of the apparent finality of that cancellation, things weren’t quite over just yet.  In August 1974 George Daniels of Conmaco’s Chicago branch had facilitated the test of Vulcan’s Decelflo muffler.  Vulcan reinstated Conmaco’s Chicago branch only as a Vulcan dealer in October of that year, but it was short lived: within a year Vulcan once again cancelled the agreement, which proved to be the final one between the two organisations.

Conmaco went on to produce a line of Vulcan style air/steam hammers both smaller and larger than the ones Vulcan had manufactured.  During the 1970’s and 1980’s Conmaco was a competitor to Vulcan both onshore and offshore.  Things flared up again in 1978 when Vulcan accused Conmaco of infringing upon its Vari-Cycle stroke control patent with Conmaco’s energy selector device.  Conmaco responded with a cross-licensing proposal, but Vulcan, true to form, flatly refused its request.

Although Conmaco’s product line was developed as a competitor to Vulcan, the fact that the hammers that resulted were basically Vulcan style hammers are a testament to basic durability of both the design and the hammers.  Additionally companies which service Vulcan hammers can also do the same for their Conmaco counterparts.

Additional information:

Vulcan vs. the Machinists’ Union: The Turkey Grievance

Although Vulcan’s relationship with the Machinists’ trade union at its Chattanooga facility was generally reasonable as such things go, every now and then conflict would arise outside of the triennial (usually) contract negotiations. Probably the most significant of these conflicts–and certainly the best publicised–was the “Turkey Grievance” in 1981. It’s a good way to illustrate the whole grievance and arbitration process, and in itself is an interesting piece of labour law.

The best place to begin this narrative is with the following notice, which was posted in the shop on 16 November 1981:


Years ago during a financially successful period, the Company, wishing to share such success with employees, decided to give employees a turkey at Thanksgiving and a ham at Christmas. The Company has continued this practice each year since that time.

Unfortunately, this year has not been a particularly successful year. As a consequence, we regret to inform you that this year we will not provide a turkey at Thanksgiving. We will, however, give employees a ham at Christmas.

The usual procedure in the event an employee had a complaint against Management was to hold a meeting consisting of representative of the Union’s shop committee and Management. If the matter could not be satisfactorily resolved there, the Union would file a grievance. (Sometimes the order between meeting an grievance would be reversed.)

In this case the sequence was a little different. The events were outlined by the Company’s labour attorney, Bernard J. Echlin of Vedder, Price, Kaufman and Kammholz, in a letter to Mr. Douglas D. Walldorff, Acting Regional Director of the National Labour Relations Board in Atlanta:

Before making any decision with respect to the Thanksgiving turkeys, the Company on November 16, 1982 called a meeting of the shop committee of the Union. At that meeting the Company discussed with the committee its proposed course of action. While the committee did not agree with the Company on the matter, neither did it disagree. Nor did the Committee propose any alternate course of action. Following the meeting the Company posted a notice for the information of all employees. A copy of the notice is enclosed. This notice was shown to the shop committee before it was posted. The shop committee did not object to its posting or ask that the Company delay the posting of the notice.

A day or two later a member of the shop committee asked a representative of the Company whether the Company would be willing to meet with Union Business Representative Edward Pierce to discuss the Thanksgiving turkey matter. The Company readily agreed to meet with Pierce and did meet with him. In that meeting Pierce objected to the Company’s proposed course of action but did not persuade the Company to provide employees turkeys at Thanksgiving.

Turkey-GrievanceThe Union filed the formal grievance (shown at right) four days later. It is noteworthy that Edward Pierce, Directing Business Representative for the Success Lodge 56 of the International Association of Machinists and Aerospace Workers, signed the grievance. Generally speaking a grievance would be signed by a member of the Union’s shop committee or the grievant.

The Company formally responded to the grievance on 1 December 1981 as follows:

This is in response to the grievance dated November 20, 1981 in which the Union alleges a violation of Article 1.1 and other provisions of the agreement between the Company and Union because the Company this year has not provided employees a turkey for Thanksgiving.

Section 15.1 provides that the parties’ written agreement “constitutes the entire agreement between the parties.” Thus, if there were a commitment on the part of the Company to provide employees with a turkey every Thanksgiving such commitment would have to be found in the written agreement. Neither Article 1.1 nor any other provisions of the agreement contain such a commitment.

The grievance describes the Company’s action in not providing a turkey as “unilateral action” and as having taken place “without negotiations with the Union.” Under the express provisions of Section 15.1, the Company had no duty to bargain with the Union with respect to the matter. Nonetheless, despite the provisions of Section 15.1 and long before the Thanksgiving holiday, the Company discussed its intentions with the Union Committee. Thereafter, still long before Thanksgiving, the matter was discussed at the Union Committee’s request with Union Business Agent Edward Pierce. Thus, even if there were an obligation to bargain with the Union with respect to the matter, the Company satisfied such obligation.

The grievance is without merit and is hereby denied.

The articles cited by both sides refer to the Union contract in force, the result of collective bargaining in 1980.

The Union responded by filing an unfair labour practice complaint with the National Labour Relations Board’s office in Atlanta. The NRLB responded on 22 March 1982, in part as follows:

In accordance with the National Labour Relations Board’s decision in Collyer Insulated Wire, 192 NLRB 837 (1971), and pursuant to “arbitration deferral policy under Collyer – revised guidelines, publicly issued by the General Counsel on May 10, 1972, I am declining to issue a complaint on the instant Charge based on my determination that further proceedings on the Charge should be administratively deferred for arbitration.

My reasons for deferring the charge are as follows: the issue raised by the instant charge is one that can be considered and resolved under the grievance arbitration provisions in the current labour agreement between the parties, and a grievance has, in fact, been filed. Moreover, the Charged Party has notified this office, that it is now, and for a reasonable period will be, willing to arbitrate the dispute underlying the charge in the above-captioned case, notwithstanding any contractual time limitations on the processing of grievances to arbitration.

The arbitration process was facilitated by the Federal Medication and Conciliation Service in Washington, DC. Here, the FMCS would submit a panel of names of potential arbitrators to both parties. Those submitted had met the Service’s requirements to act as an arbitrator. If one of these was suitable to both parties, the arbitration would proceed. If none of the names submitted were mutually satisfactory, the Service would submit another panel and the process would be repeated.

Arbitration-Panel-RequestThe FMCS submitted its first panel on 8 December 1981; none of the members of the panel were acceptable to both sides. Since the NRLB had elected to defer the unfair labour practice charge until arbitration was resolved, the Union initiated a request for another panel of arbitrators, which the Company co-signed and sent back to the FMCS (see right.) The FMCS submitted a panel of five to both parties on 6 March 1982. Mr. Ralph Roger Williams of Tuscaloosa, AL, proved acceptable to both parties and was selected to arbitrate this case.

Mr. Williams submitted times that he could be in Chattanooga for the arbitration meeting, and the meeting was set for 1000 8 June 1982.

At the arbitration meeting both sides presented their case, calling witnesses and cross-examining opposing witnesses. After the meeting both sides presented post-hearing briefs on the subject, which could comment on the testimony given (both positive and adversely) and present what each side felt was the applicable law in the case.

The arbitrator issued his ruling in favour of the Company on 30 July 1982. His decision can be read here. The decision came as a shock to the Union (and frankly I was surprised as well.)

Without a favourable decision from the arbitrator, the Union opted to let the unfair labour charge lapse. It lapsed so thoroughly that, on 30 January 1984 (nearly a year and a half after the arbitration ruling) the NRLB contacted Union and Company alike to see how things were moving along. The Company responded by submitting a copy of the arbitration to the NRLB. Although the Board offered the Union the option for appeal, the Union did not opt to do so, and the matter of the Thanksgiving turkeys came to a close.

The case was significant enough that the following notice appeared in the 25 October 1982 edition of U.S. News and World Report:

Turkeys given to union employees are a gratuity that can be withdrawn at any time, holds an arbitrator. Vulcan Iron Works halted its custom of giving turkeys to machinists after a poor financial year. The arbitrator finds against the union after determining that its contract did not mention the gifts and that employees had not considered them income, since they failed to declare them for tax purposes.

Although the Company regarded the matter as a significant victory, its long-term effects were decidedly mixed. With the collapse of the offshore oil industry, Vulcan’s business suffered in the years immediately following the grievance. The Union opted to pass on contract negotiations in 1983 and 1984, making a three-year contract into a five-year one. It also opted to leave the Machinists’ pension plan in 1992 for a 401(k), a significant step since trade union’s traditionally regard participation in the union’s plan as an important way of bonding the members to the union. On the other hand, decertification–even when an important union “right” was not recognised as was the case here–was never seriously entertained at Vulcan, and the Machinists continued to represent the shop employees until the plant was closed in 1998. (For some of my thoughts as to why people stick with unions even when the economic benefits are not apparent, click here.)