Hydraulic Vibratory Hammer: the Foster 4200


The Foster hammers are an interesting part of Vulcan’s last years.  The concept was to produce a vibratory hammer under a “private label” arrangement, a novelty for the pile driving equipment industry at the time (although the “Chimag” diesel hammers were a similar concept.)  In the early 1990’s Vulcan produced several vibratory hammers for the sheet pile supplier L.B. Foster, merging Foster’s own exciter technology (which was Japanese in origin) with Vulcan’s power pack improvements after parting ways with HPSI.

The purpose of this article, however, is primarily to discuss hydraulic vibratory hammers with an emphasis on the “hydraulic.”  Pile driving equipment in general and vibratory pile driving equipment in particular poses challenges for the equipment designer and manufacturer.  The emphasis here will be on the hydraulics, which are in some ways typical of mobile hydraulics and in some ways unusual.  We will use the Foster 4200–the largest of the hammers Vulcan produced–as an example.

Basics of Fluid Power

Fluid power is an important component of power transmission for many mechanical applications.  Broadly fluid power can be broken down into two parts: industrial and mobile applications.  The emphasis here will be on the latter.  Most construction machinery manufactured today has some hydraulic control or power included in the equipment.  The original vibratory hammers, designed and produced in the then Soviet Union, were electric, but with the introduction of hydraulic vibratory drivers in the 1960’s today most are hydraulic.

The concept behind fluid power is simple:

  1. A flow of fluid is pressurised.
  2. The fluid is delivered to where it is needed.
  3. The fluid does its work, and is depressurised in the process.
  4. The fluid, at or near atmospheric pressure, returns to where it is pressurised.
  5. The process is repeated.

Since the hydraulic fluid being used (petroleum-based, plant-based, or even water) is virtually incompressible, considerations such as occupy students of thermodynamics are not an issue here, although they appear in other parts of the hydraulic system.  The power being delivered at any point by a fluid power system is the product of pressure and flow, thus

HP=\frac{ PSI \times GPM}{ 1714}

where HP is the horsepower being delivered by the fluid, PSI is the gauge pressure of the fluid in pounds per square inch, and GPM is the flow rate in U.S. gallons per minute.

Both the flow and pressure of the fluid are generally delivered by pumps.  Fluid power is a little different from many pumping applications in that the pumps are almost always positive displacement, to insure the accuracy of the flow and pressure.  The pump performs the first step of the process outlined above; if the power is transmitted at the other end in a rotary way, a motor performs the third step.  Unsurprisingly hydraulic pumps and motors are basically mirror images of each other; their construction is similar (not necessarily identical) and their operation is similar.  There are several types of hydraulic pumps and motors in use; we’ll concentrate on piston motors and pumps for two reasons:

  1. These are what were used in the Foster 4200 and all of Vulcan’s high pressure machinery (more about pressure levels later.)
  2. They’re a little easier to visualise if you’re not familiar with the application.

Let’s look at a cutaway pump/motor.


The large bronze barrel in the centre of the motor has a series of holes in a radial pattern, into which the pistons (those small cylinders with the rounded left ends) are inserted.  The plate an an angle to the bronze barrel is the swash plate, which (through the other plates) either drives (motor) or is driven by (pump) the spline at the left end of the motor, which connects to whatever the motor is driving or the pump is being driven by.  The angle between the barrel and the swash plate alows the pistons to move in the barrel, much like the pistons in an automobile do.  The displacement per piston is given by

DIS = A \times L

where DIS is the displacement of each piston in cu.in., A is the cross-sectional area of the piston in sq.in., and L is the stroke of the piston in the barrel in inches.  The total displacement per revolution of the barrel is obviously

DIS_{tot} = DIS \times N

where DIS_{tot} is the total displacement in one revolution in cu.in./revolution and N is the number of pistons.

The flow of the pump or motor is given by

GPM = \frac {DIS_{tot} \times RPM}{231}

where RPM is the rotational speed of the motor is revolutions/minute.  The horsepower output of the motor (or input of the pump) was given earlier.  The torque of the motor/pump is

T = \frac{5252 \times HP}{RPM}

where T is the torque of the motor or pump in foot-pounds.

The simplest configuration of a pump or motor is a fixed-displacement.  In this the angle between the barrel and the drive spline (or keyway) is fixed and thus the stroke L is fixed.  This precise flow is an important aspect of fluid power systems, as it regulates the speed of the load, which can be rotational or translational.

However, there are good reasons for varying the displacement–and thus the flow output–of a hydraulic pump.  The pump shown above has a mechanism on the right which varies the position of the right end of the barrel, and thus the displacement of the pistons.  Although it’s possible to vary this control manually, the mechanism shown does so by measuring the pressure of the hydraulic fluid, and thus is referred to as pressure-compensated.  We’ll come back to why that’s important later.

One further thing that complicates the equations above is leakage around the pistons.  Because of the speed of their movement, most piston pumps and motors–and for that matter most hydraulic motors, pumps and valves–do not use seals in them, but tightly fit the pistons or valve spools to the barrel or valve body and allow some leakage.  Although this may seem inefficient, the small oil flow allows for lubrication of the components and cooling of the pumps and motors without the drag that seals would induce.  This reduces the flow available for power transmission, and the ratio of the ideal flow to the flow actually available after leakage is referred to as volumetric efficiency.  This can vary for a number of reasons, but for piston motors and pumps such as we are dealing with here a volumetric efficiency of 90% is a good estimate.

Overview of the Foster 4200

A general overview of vibratory hammers, their theory and application, is given in the monograph Vibratory and Impact-Vibration Pile Driving Equipment.  Without going into the detail of that monograph, the following illustrates the basic components of the system.


The exciter does the work of the system by inducing a vertical, sinusoidal force in the pile, which sinks by its own weight.  The eccentrics producing this force are turned by a motor, which is driven by a pump on the power pack through the hoses.  The pump in turn is driven by a diesel engine, which is the prime mover of the system.  There’s also a clamp to hold the driver to the pile; in reality there are two hydraulic systems, one for the motor and one for the clamp.  This allows us to illustrate hydraulic systems with both rotational output (the motor) and translational output through a cylinder (the clamp.)

The exciter for the 4200 is shown below.  If you’re interested in details on this unit, including the specifications, you can download the Foster 4200 Field Service Manual, First and Second Units.


The 4200 has two motors to drive its eight (8) eccentrics, but in this case only one clamp for sheeting.  A hose layout for the exciter is shown below.

The hose layout for the 4200 exciter. There were five hoses going back to the power pack: the supply hose (94), return hose (93), case drain (2), and the two clamp hoses (15).

The supply hose is the hose which carries the pressurised oil to the motors.  The return hose carries the depressurised oil back to the power pack.  The case drain hose returns the leakage from the motor (see earlier comments on volumetric efficiency) back to the power pack.  In some cases the return hose can be used as a case drain hose, but in this application there is too much back pressure.  The two clamp hoses are bi-directional, as we will discuss shortly.

foster pp outside
An outside shot of the power pack. The control panel is for an electric-over-hydraulic system which Vulcan began using when it built its own power packs. The remote control pendant allows the unit to be operated from the crane or wherever is most convenient. The hoses come from the exciter and can be 50′-150′ long for this size of unit. The quick disconnects allow easy connection and disconnection of the hoses from the power pack, but should be avoided when possible because they contribute to frictional losses in the hydraulic fluid.

Until 1991 Vulcan purchased its power packs from HPSI, which used an air-over-hydraulic control system.  In going to electric-over-hydraulic, Vulcan was more in line with its competition such as ICE and APE.

foster pp inside
Looking through the doors of the power pack. The diesel engine is the prime mover of the system. The pump drive transmits torque to the “pump stack,” which includes the pressure-compensated pump for the motor and the clamp pump. The reservoir stores and cools the oil between cycles in the hydraulic system.

Older hydraulic systems used several types of transmissions between engine and pumps, including power-take-off (PTO) drives.  Vulcan used a direct drive which was basically a thin circular plate which transmitted torque from the engine to the pump shaft.  This seriously reduced mechanical inefficiencies in the system.

The hydraulic schematic of the system. Most of the fluid is stored in the reservoir (1) until drawn into the main pump (6) or the clamp pump (6e). The fluid driving the motors runs through the supply quick disconnects (12) to the hammer, and returns through the return quick disconnects (13). The oil then goes through the oil cooler (15) before returning to the reservoir (1). The clamp pressure, alternating between the two hoses (10) and (10a), is controlled by a directional control valve (7). The case drain line (19) goes directly to the reservoir (1) to minimise back pressure.

We can thus see the essential components of the hydraulic system:

  1. Prime mover (4).
  2. Pump to pressurise the oil (6,7)
  3. Equipment to do the work and depressurise the soil (motors, clamp)
  4. Cooling and oil storage (1,15) after oil is returned to the power system.

Pressure in Hydraulic Systems

Fluid power systems are designed to deliver an oil flow at a given flow rate.  So what pressure does this come at?  This is a key point in fluid power systems: pressure is determined, without any other restrictions, by the load itself.  This is true whether the load is rotational or translational, and in this system we have both.

Rotational (Motor) Loads

The main pump (6) is configured to, at maximum displacement, deliver a given flow of oil.  This in turn determines the rotational speed of the motor.  The power put into the system–a product of the pressure and the flow–varies with the load of the system, and that in turn varies with the pressure.  So are there limits to that pressure?

The answer is obviously yes.  Hydraulic systems, broadly, divide themselves into two pressure ranges: high and low.  Low pressure systems generally have their pressure limited to about 2500-3000 psi, and a wide range of hydraulic systems operate in this range.  High pressure systems run around 5000 psi and beyond.  Each system has its own standard of components.  The Foster 4200, in common with Vulcan’s 2300 and 4600 vibratory drivers, was a high pressure unit.

So how do we limit the pressure?  In “classic” systems such as the HPSI power packs Vulcan started with, there was a fixed displacement pump and a relief valve (11).  When the pressure reached the system maximum, the hammer would slow down and the excess oil would “dump over relief.”  This worked, but it generates a lot of heat, which means that the system’s thermal protection shuts it down in short order.

On this hammer, Vulcan used a pressure-compensated, “load-sense” system, which required a special relief valve which, in turn, informed the pump that maximum pressure had been achieved.  The pump would reduce its displacement as described earlier until the flow and pressure were balanced.  There was still a relief “just in case” for safety purposes, but much of the dumping over relief was eliminated.  This also enabled the operator to slow the hammer down without slowing the engine down, allowing the engine to operate in its best torque range.

Translational (Cylinder/Clamp) Loads

Before we describe the hydraulics, let’s look at what the clamp does.  Some types of hydraulic clamps are shown below (from this article.)

Figure 54. Types of clamps that grip the pile using friction. a) clamp for steel pipe and reinforced concrete shells, b) direct-action clamp using a hydraulic cylinder and accumulator for sheeting and similar piles; c) lever-type clamp for piles and pipes.

Fun fact: Foster started this project using the clamp type (c), which they inherited from the Japanese.  This clamp was common when cylinders were not large and the lever action was needed to obtain the force, but these clamps were heavy.  Vulcan convinced them to convert to type (b), albeit without the accumulator, which had been a U.S. industry standard for a long time.  Vulcan also convinced Foster to adapt an integral clamp cylinder as opposed to the bolt-on types which are still common on American vibratory equipment.

The clamp is simple: the clamp in pressure is applied to the large (head) end of the clamp and the jaws close on the pile, clamping the pile and stopping the cylinder.  When driving is finished, pressure is applied to the small (rod) end of the clamp and the jaws open.  The speed of the movement either way is computed as follows:

V = \frac{231 \times GPM}{A}

where V is the velocity of the clamp in in/min and A is the area (head or rod) under pressure, sq.in.  Moving cylinders is a common hydraulic operation, one (with the right control system) can be done with great precision.  In this case positional precision wasn’t a big deal; the clamp cylinder just needs to go until the pile stops it.

So what happens then?  The system “deadheads,” and, with a fixed displacement pump, most everything goes over relief, with the usual heating problems.  There are ways of getting around that but here again pressure-compensated pumps are the best answer, going to nearly zero flow when the clamp deadheads.

Another fun fact: In adopting Vulcan’s clamp design, Foster also incorporated Vulcan’s safety check valve in the clamp, which continued pressurisation in the event the clamp hose was cut.  Foster also incorporated Vulcan’s interlock that made it impossible to start the hammer until the exciter was clamped to the pile.  That’s one reason why a separate hydraulic system is needed for the clamp; another is that full clamp force and pressure is guaranteed, whereas with the pump for the motor the pressure varies.


We’ve covered a lot of ground here.  Hopefully you’ve gotten a better idea of what hydraulic systems in general–using a real-life application–are all about.  Vulcan’s effort with Foster’s vibratory hammers was a unique one, and probably resulted in the best vibratory driver (if not the most economical to produce) Vulcan ever made.


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/sq.cm.)

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.

Reciprocating Vibratory and Impact-Vibration Hammers

Virtually all vibratory pile drivers use rotating weights to produce the alternating force that mobilises and fluidises the soil. But it’s necessary to use the weights in pairs to cancel out the horizontal forces that result. What if the force could be produced using a reciprocating weight, thus eliminating the horizontal cancellation requirement? In the 1970’s Vulcan investigated this problem. We discuss here the two solutions it considered: the linear vibrator and the hydroacoustic driver.

Linear Vibrator

The linear vibrator was the brainchild of John J. Kupka, an Austrian immigrant who had done work for MKT on their “C” series hammers and the Horn Construction “HC” hammers that Vulcan had produced. Kupka designed the linear vibrator for Vulcan and it was tested at Vulcan’s West Palm Beach fabricating facility in March 1971.

The Linear Vibrator, from its U.S. Patent (3,704,651) drawing. The hammer worked by applying air through the inlet 70, which was admitted through slots 56b in the ram 56. The slots were set up so that they only admitted air to one side at a time and for a short length; the rest of the ram travel upward or downward past the slots expanded the trapped air and extracted the energy from it in that way. The two air cushions at the ram ends 56a additionally bounced the ram between its two extremes and provided for transmission of the alternating force to the body, and through the primitive clamp at the bottom to the pile.

The total assembly above was 67 5/8″ long; the small and large diameters of the piston 56 are 4 1/2″ and 7 1/2″ respectively. The outer diameter of the housing 18 was 11 1/2″.

Linear Vibrator-25
Sample Output from the Linear Vibrator’s test at the Special Products Division facility.

The results of Vulcan’s March 1971 tests by Continental Testing Laboratories can be found here. The device worked as designed but Vulcan never pursued the technology. The primitive clamp has been noted; the suspension 40 at the top, with its Belleville washers, was already being made obsolete (along with the steel coil spring suspensions of earlier vibratories) by the rubber springs Vulcan was to use in the next decade. A more serious problem was the magnitude of the force. The peak force put out by the hammer was just shy of 5 U.S. Tons. The 400, the smallest rotating vibratory Vulcan put out, had a force of 17 tons, and the “small” 1150 (which competed with machines such as the MKT V-5 and ICE 216) 42 tons.

The basic problem with the force lay in the use of compressed air and the piston ring sealing technology used in the machine. Although Vulcan could have easily produced such a machine with the technology it used for the air/steam hammers, upscaling it to compete with vibratory hammers even in the early 1970’s would have resulted in a fairly large machine. The concept of a ram/valve with expansive use of the compressed air resurfaced in the Single-Compound Hammer which Vulcan developed a decade later.

One way of getting around the size and pressure problem would have been to make the device operate with hydraulic fluid and the pressures that went with that. In the late 1970’s Vulcan toyed with that idea but it was also presented with another concept, namely the hydroacoustic driver.

Hydroacoustic Driver

In 1974 the Naval Civil Engineering Laboratory issued its Technical Note N-1362, “Evaluation of a Hydroacoustic Rapid-Impacting Pile Driver”by Dr. Carter J. Ward. This report described the operation and tests on a new concept in hydraulic pile driving. Abstract for the report is as follows:

Tests to evaluate the driving capabilities of the rapid-impacting hydroacoustic pile driver on various types and sizes of vertical piles and horizontal batter piles are descjbed and discussed. The functional and operational characteristics of the driver are described, test results and output analysis are presented, and the hydroacoustic driver is compared operationally and economically with the vibratory driver and conventional diesel pile hammer.

Hydroacoustic driver, as tested by NCEL. The device, although compared with a sinusoidal vibrator, is actually a reciprocating version of an impact-vibration hammer in that the lower end of the hammer/valve (piston) impacts an anvil rather than transmitting force alternately. Like the impact-vibration hammer, it was capable of high blow rates, if not a great deal of energy for each blow.

The concept was presented to Vulcan in 1977-8; however, the company was going through a generational change, this compounded by the demands of the existing air/steam line (the 6300 was being designed and produced at the time.) To apply this to conventional pile driving would require some conceptual changes in the transmission of energy to the pile (i.e., lower energy per blow with higher blow rate vs. high energy per blow and low rate.) Additionally the lower energy per blow may or may not be able to move the pile past the quake point of the soil and induce plastic deformation of the soil, essential in impact pile driving. The hydroacoustic driver nevertheless remains an interesting concept for generating impact, if not pure sinusoidal vibration.

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.

Vulcan High-Frequency Vibratory Hammers

The mid-1980’s were lean years at Vulcan. The offshore market was still down, the aftermath of the collapse of oil prices earlier in the decade. Vulcan’s own diesel program had to be stopped, plagued by design and manufacturing problems and an overvalued US Dollar. The vibratory hammer program was going reasonably well but the market was competitive. Vulcan had reached the point where it had effectively closed its own manufacturing facility and farmed out what was left.

It was in this gloomy situation that Vulcan designed and produced one of the most innovative products it had ever produced, the 400 vibratory hammer, the first of Vulcan’s high-frequency machines.

High frequency (~2400 RPM, not to be confused with the ~7200 RPM resonant machines) vibratory drivers had been produced in Europe. Depending upon the soil conditions and configuration of the pile, the vibrations used to drive or extract the pile can also be transmitted to neighbouring structures. Since European contractors drove piles more frequently in close quarters with sensitive structures than their American counterparts, European vibratory manufacturers produced high frequency machines first. Their higher frequency, combined with lower amplitude for the dynamic force, reduce the transmitted vibrations through most soils.

Vulcan’s rationale for a high frequency machine, however, was somewhat different. The first impetus for the 400 was the development of aluminium sheet piling, which made development of a driver smaller than the 1150 attractive. MKT had already developed a medium-frequency small machine (the V-2) to drive aluminium sheet piling, but the machine a) weighed over a US ton and b) had a clamp suited to steel piling, which mangled the heads of aluminium sheets.

What was needed was a lighter machine whose clamp was easier on the pile. Vulcan’s interpretation of the theoretical data led it to believe that a high frequency machine would drive the piles (which was certainly the case with the lighter sheeting sections.) The result was the first 400 vibratory hammer, designed and built in the summer of 1987.

The 400 had several innovative features:

  • A one piece gear-eccentric, machined out of plate with the eccentric weight burned out. The gear teeth were a much smaller pitch than their medium frequency counterparts, a feature replicated on the “A” series machines four years later. The small pitch ran more quietly an dispensed with the need for surface hardening.
  • A clamp that was burned out of plate. The cylinder bolted to it used the flat end of the rod as the movable jaw. This only left a shallow round dent in the sheeting when clamped.
  • The “U” configuration which wrapped around the exciter case and transmitted the force from the crane to the pile during extraction. This and other features were subject to U.S. Patent 4,819,740. (This patent has been a nuisance to Vulcan’s competitors for long time, cited in several patents from inventors at HPSI, APE, J&M, ICE and MGF.)
  • It was the first Vulcan pile driving machine to completely dispense with castings.

The result was a machine that weighed only 1100 lbs.–half of the MKT V-2–and still drove the piles successfully.



Vulcan’s Medium Frequency Vibratory Hammers

In 1984 Vulcan re-entered the vibratory hammer market with the introduction of the 1150 vibratory hammer. This hammer made its debut on a project in Bangor, Maine for Cianbro Construction. More suited for the American market and adequately powered, these machines were far more successful than the Vulcor hammers had been.

The technology used was pretty typical for vibratory hammers of the era, including the large-pitch teeth gears bolted to cast steel eccentrics, 355 mm (14″) throat width for American-style sheeting installation, Volvo hydraulic piston motors (for the high pressure units; vane style motors were used on the low pressure 1150,) and a clamp with an industrial style cylinder bolted on to push the movable jaw into the fixed jaw. Both jaws had two parallel sets of teeth with a gap in between to accommodate the interlocks on the sheet piles, which enabled the hammer to drive two sheets at a time.

Vulcan produced three sizes of medium frequency hammers, the 1150, 2300 and 4600. The size designated the eccentric moment of the hammer in inch-pounds. All of the hammers rotated at 1600 RPM.

Vulcan used the HPSI power pack for its vibratory hammer throughout the 1980’s. (One of these is shown on the flatbed trailer in the 4600 photo below.) This power pack was simple and reliable, using air controls (as opposed to the electric controls used by competitors such as ICE and later APE.

Note: if you’re looking for service and other technical information on Vulcan vibratory hammers, take a look at the Vulcanhammer.info Guide to Pile Driving Equipment.

Below: a 2300 on the job driving h-beams in Portsmouth, Virginia, in 1990. The contractor was Tidewater Construction. A diesel hammer can be heard driving piles in the background for part of the video.

Below: the 2300L extracting soldier beams in Atlanta, Georgia, in December 1990. The fact that these machines can both drive and extract piling without modification is part of their appeal.

Below: a video of the installation of bearings in the 2300L, and a little “tour” of PACO’s yard.

The “A” Series Vibratories

In 1991 Vulcan introduced the “A” series of hammers (1150A, 2300A and 4600A) series of hammers. The biggest changes were a) the abandonment of the Morse shear fenders and b) the complete reconfiguration of the gear and eccentric design, inspired by information obtained from the Soviets. The first “A” series hammer was a 2300A, first used on a job by Agate Construction in New Jersey.

Vulcan also began to manufacture its own power packs, where it was able to make many technological advances.

Foster Units

One of Vulcan’s more interesting ventures in the 1990’s was the private label manufacture of a line of vibratory hammers for L.B. Foster in Pittsburgh. The first hammer to be produced was a replica of Foster’s existing 1800 unit, but it became apparent that this unit was very expensive to produce. Vulcan then designed a line of medium frequency vibratory hammers, the 1050, 2100 and 4200 hammers. With the combination of Vulcan’s and Foster’s experience in vibratory hammer design and manufacture, this was the best line of medium-frequency vibratory hammers that Vulcan ever produced.

Some general arrangements of the Foster hammers are here.

After the Acquisition

After it was acquired by Cari Capital, the company continued to support the line; however, it was left behind when Vulcan Foundation Equipment acquired the air/steam hammer line in 2001. It was ultimately sold at auction the following year.  Current service and support for these units is furnished by Pile Hammer Equipment.

Uraga/Vulcor Vibratory Hammer

Vulcan’s first venture into the vibratory market took place in the 1960’s with the introduction of the Uraga electric vibratory hammer from Japan, which Vulcan marketed as the Vulcor Vibratory Hammer.

Vibratory pile driving technology had been developed in the Soviet Union. One of the first countries to pick up the technology was Japan. With its volcanic soils, it is an ideal place for a vibratory hammer to be used.

Most early Japanese vibratory hammers (which are described some here) followed the Soviet pattern of electric motor(s) driving eccentrics through a chain drive system. (An example of this kind of design is shown here.) This unimaginative application of the technology prompted one Soviet trade official to describe the Japanese as “not very good students.”

The Uraga/Vulcor machine was a departure in that Uraga reversed the rotor and stator on the electric motors and positioned one motor inside of each eccentric. This resulted in a vibrator with a more direct drive than has been seen before or since, making for an efficient construction and operation.

Uraga VHD-1 model, with only one “stack” of eccentrics.

Unfortunately the width of the machine clashed with the normal American practice of setting the sheets before driving, which requires either that the vibratory hammer be narrow enough (less than 355 mm) at the throat or use an extension (which adds to both the vibrating mass and hanging weight of the hammer.) Some Uraga machines also suffered from misalignment of the eccentric bearings, a function in part of the “modular” construction of the machines (to increase the number of eccentrics, it was simply necessary to add another “stack” to the unit.

All of these difficuties, combined with American contractors’ aversion to electrics on the job, put the Vulcor at a disdavantage to other vibratories coming into the U.S. By the time Vulcan moved to West Palm Beach, the Vulcor programme was pretty much over and it would be another twenty years before Vulcan would attempt a vibratory hammer again.

The Uraga VHD-2 (with two eccentric stacks) at a power plant project in California. The hydraulic clamp, although primitive by modern standards, was an advance over the “lever-style” clamps use by many other Soviet and Japanese units. Even Foster was still using lever-type clamps in their units in the early 1990’s.

More on the Uraga/Vulcor Hammer:

Vertical Drains: Sand and Wick

One of the more interesting applications of vibratory hammers is the installation of wick drains, a type of vertical drain. But vertical drains are of interest far beyond vibratory technology.


Soils are a composite of solid soil particles, water and air. When soils reside below the water table (phreatic surface,) there is no air, and the soils are referred to as saturated. The soils dealt with in vertical drainage are generally saturated.

Soil particle sizes vary, and with that variation come many of the variations in soil properties. Soils with large particles (sand and gravels) are referred to as cohesionless soils. Soils with smaller particles are usually silt or clay soils and are referred to as cohesive soils.

In either case, water not only fills the space (voids) between the soil particles, but it is capable of flowing through the soils as well. Flow in rivers and streams is due to the fact that the water is flowing “downhill” due to gravity, and the same phenomenon can take place in the soil voids. The property of soils relating to their allowance of water flow in the voids is referred to as permeability. As a general rule, the smaller the particle size, the lower the permeability of the soils.

One action that can result in water flow in a soil is the placement of a new load on top of the soil, which in turn exerts downward pressure on the soil. Unless the soil particles are in their most compact arrangement (which is unlikely,) water will be forced out of the soil voids under the new load. If this water is forced out, the structure on top will settle, sometimes significantly.

In the case of cohesionless soils, the large particle size enables relatively rapid water flow out from under the load, and the settlement can be very rapid. But if the soil is cohesive with small particles, the water movement (and thus the settlement of the structure) can be very slow, sometimes months or years. Structures built on top of this kind of soil can be fine to start with but over time settle significantly, creating serious structural damage and requiring expensive repair or demolition of the structure.

Although there are several ways of dealing with the problem, one of them is to drain water out of the cohesive soils before placing a structure on top of them, thus getting the settlement out of the way and enabling a stable structure to be built. The method used to do this is referred to as vertical drainage, and specifically two types of vertical drains and their installation will be described here: sand drains and wick drains.

Sand Drains

Theory and Methods of Sand Drains
A Vulcan 140C hammer installing sand drains.

A sand drain is basically a hole drilled in a cohesive soil and filled with sand. Since the sand has larger particle size, its permeability is much higher, thus water will flow through it much more easily. As shown above, an array (it’s actually a two-dimensional array) of sand drains is installed, and a load is applied on top of the drains. The load shown above is an embankment, such as is used on a highway, and an additional, or surcharge, load is used to speed up the drainage process. The excess water is collected at the top and directed away from the jobsite.

The tricky part comes in getting the sand drains in the ground. The obvious solution is to simply drill the holes and fill them with sand, but if the soil is soft (which is frequently the case,) the holes will collapse. Although sand drains were first used in California in the 1930’s, the sand drain projects that were of special interest to pile hammer equipment manufacturers took place in the late 1950’s and early 1960’s during the development of the Meadowlands area of northern New Jersey.

Both Vulcan and McKiernan-Terry (MKT) developed equipment to install sand drains. Vulcan developed a special series of differential-acting hammers referred to as sand drain hammers. The major difference between the sand drain and conventional hammers was in the cylinder head at the top of the hammer. Instead of using sheaves for the hammer lifting cables, a bar which interfaced with a retractable hook was used. Vulcan developed a special hook block which was patented (U.S. Patent 3,171,552.)

Basically, the hammer would first drive a mandrel (a piece of pipe) into the ground. After this, the sand drain door (the large piece just below the hammer) would be opened, and sand would be dumped into the mandrel. Compressed air is then applied to the sand, the hammer is hooked to the crane with the hook block, and the mandrel is pulled out of the ground, leaving the sand column in the earth to do its job.

A detailed description of the process is given in MKT Bulletin 71 (why Vulcan didn’t develop a piece of literature like this is beyond me.) Below is another view of a Vulcan hammer over a sand drain dump tube.

Another view of a Vulcan hammer over a sand drain dump tube.

Wick Drains

A cursory examination of the procedure for sand drains shows that the procedure is fairly involved. It invites simplification, at least for some applications. A popular simplification is that of wick drains.

A wick drain is just what the name implies: a geosynthetic “rope,” usually about 100 mm wide and 5mm thick, which acts as a high-permeability conduit for water to flow out of the soil and to the surface, in the same manner as takes places with sand drains. As is the case with sand drains, they are installed as an array, generally in 3 metre spacings.

Candle makers have the luxury of melting the medium into which their wicks are places. Since things aren’t so simple for the contractor, he or she has to use a mandrel to insert the wicks. The simplest way to do this is to push the mandrel/wick combination into the ground, but some soils are too stiff for this, so the mandrel is frequently vibrated.

Vulcan vibratory hammers have been used in some cases to install wick drains. Since many drains are installed, this is a fairly demanding application for a vibratory hammer, but it is another example of the versatility of vibratory pile drivers.

Vulcan 4600 sporting bias weights and a caisson beam installing a pipe mandrel for a wick drain project, 1997. Vibratory hammers are best in cohesionless soils; since vertical drains are used in cohesive ones, this places yet another demand on the performance of the machine.

Further Information

A much more detailed explanation of the theory and installation of vertical drains–and especially of wick drains–can be found in the FHWA document RD/86/186, Prefabricated Vertical Drains, August 1986, which can be found by clicking here.