Harwood
Engineering Company, Inc.
Hydraulically Driven Pumps
The future need for higher pressure with consequent demand on equipment
will make these pumps increasingly important — December, 1957
by Donald H. Newhall
Excessive torque, side thrust, and friction make crank-driven pumps cumbersome
and expensive at high pressures. But in hydraulically driven pumps, these
effects are eliminated or minimized and higher pressures can be attained.
Also, longer strokes are characteristic; therefore, rates of stress buildup
are reduced in such areas as check valves and packing, which results in
less maintenance and longer life. Without rotational inertia, controllability
is increased which in turn increases safety.
Essentially continuous flow and high efficiencies can be obtained with
a combination of double acting pumps.
Development of the hydraulic drive for high pressure piston pumps has
been a key factor in attaining high pressures now found in commercially
feasible systems. Compared to other forms of reciprocating pumps widely
used at lower pressures, this pump has received little attention in technical
literature. It is often called an intensifier and not only handles larger
differences between intake and discharge pressures, but makes certain design
problems which are multiplied at high pressures.
This hydraulically driven pump is useful at pressures as low as 5000
pounds per square inch — far lower than originally planned — and up to
200,000 pounds. It has high efficiency and considering its large capacity,
it is physically small. Nature of the hydraulic drive itself makes the
pump easily adaptable to automatic control, and its controllability simplifies
the problems ofsafety and maintenance. It will pump both liquids and gases.
Because of its flexibility, it is constantly being adapted to new applications
in chemical and physical processing.
Usually, high pressures of about 200,000 pounds are used for research,
but they also have a few important industrial applications — e.g., in autofrettage(12)
or cold-working of cannon(7) and other high pressure equipment.
Another important use is in strengthening materials used under intermediate
pressures of about 35,000 pounds which are now common in certain chemical
processes. High pressure also makes it possible to test pressure-carrying
parts to destruction for determining actual safety factors. The increasing
industrial use of apparatus for high pressure suggests that high pressure
technology will have wide commercial use.
All high pressure pumps now used are reciprocating piston pumps. Others
employing such mechanisms as gears, vanes, and impellers, cannot operate
at the high stress levels or handle leakage resulting from mechanical deformation
and pressure differences characteristic of high pressure systems. Reciprocating
pumps of all designs have four problems in common.
Cycling Stresses
If pressure does not exceed 20,000 pounds per square inch, almost any design
may be portioned so that fatigue and high stress will not be a major problem.
At higher pressures, more careful stress analysis is needed if safety and
long life are to be achieved. This limits the choice of materials of construction,
and it requires sound metallurgical practice.
Reciprocating pumps impose a stress cycle with each piston stroke. To
pump a given quantity of fluid, the number of strokes required and, therefore,
the number of stress cycles imposed, is inversely proportional to piston
displacement. The probability of fatigue failure is a function of the number
of cycles of applied stress so that, other things being equal, the larger
the displacement the better.
Check Valves
A pump is no more reliable than its check valves. Two of these are required
for each plunger, and for every cycle of stroking, each check valve must
function once (open and close) with full range of applied stress between
zero and maximum. The higher the rate of stroking, the faster the check
valves must seat, and hence, the greater the impact or battering of the
seat. This can be serious indeed at high pressure. Large displacement again,
therefore, lengthens valve life.
Furthermore, in many pumps the cylinder block house not only plungers,
but also ports for intake and discharge, as well as check valves. The system
of holes and cross-drillings necessary in this design creates complicated
patterns of combined stress that invite early and unpredictable fatigue
failure. This design requires a major disassembly to reseat or service
the check valves.
Packings
Regardless of the pump form, leakage across packings is a critical problem
in design and maintenance, particularly with high pressure equipment. The
speed of rubbing and the large forces present at high pressure cause friction
which in turn produces heat. This is another reason for preferring large
pistons and long strokes.
Different types of packings are available for use at high pressure.
Many in common use have been described(2, 5, 14), but the packing
design has to be chosen for the particular problem. So far, a universal
packing to meet all conditions has not been found. Some of f the useful
packings can fail from time alternating flexure resulting from frictional
forces as the ram moves back and forth. Again, larger displacements and
fewer strokes are advantageous.
Maintenance
Virtues of pump designs which require infrequent and brief downtime, are
too well appreciated by experienced chemical engineers to dwell upon at
length. In a good design, ease of repair and replacement are reflected
sharply in the economics of plant operation.
The three previously discussed design problems are also maintenance
problems, and in any good pump account for most of the downtime. But the
finest equipment, if abused, will require other attention. A system assembled
without adequate filtering, or without first cleaning components, will
certainly give trouble, regardless of quality. Besides adequate filtering,
the equipment should be arranged for ready accessibility when cleaning
and changing filters. Obviously, maintenance which is dependent on the
number of strokes will be less when large pistons of long strokes are used.
Limitations of the Crank Drive
In all four considerations previously mentioned, the fewer piston strokes
per unit volume pumped, the longer the life, the less the hazard, and the
smaller the maintenance problem. Therefore, pistons should be large and
stroke should be as along as consistent with good engineering practice
and economy.
In mechanical crank-type pumps, cost soon becomes prohibitive as diameter
and stroke of pistons increase. Force on the piston, and hence the torque
on the crank, increases as the square of the diameter; and the torque increases
again proportionally with the increase in stroke. The side thrust at the
cross-heads and journals rapidly becomes a serious design problem. For
increased capacity, crank-type pumps can be driven faster, but increased
speed quickly leads to these design difficulties. All crank devices
are necessarily subject to these fundamental limitations.
Hydraulically Driven Pumps
In the hydraulically driven pump, large pistons of long stroke can be used
to minimize these four critical design problems. It has a simple, direct,
in-line drive with no cranks, eccentrics, or crosshead bearings to add
friction and wear. Working on the intensifier principle, liquid pressure
is applied to a piston of comparatively large area, which drives a piston
of smaller area. The ratio of pressures at the two pistons is inversely
proportional to their areas.
Single-Acting Pump
The simplest practical system of this type is a single-acting hydraulically
driven pump (Figure 1). the drive portion is shown
with Joint Industrial Conference(8) symbols, but for simplicity,
the intensifier part is shown without constructional detail. The intensifier
has two main parts — the low pressure, double-acting cylinder and the high
pressure cylinder. The pistons in these cylinders are mechanically attached
— the low pressure piston pushes the high pressure piston in a delivery
stroke and pulls it in the return stroke. When low pressure is applied
to the bottom face of the low pressure piston, oil, put on the top side
by the previous return stroke, is exhausted at atmospheric pressure through
the directional valve to the reservoir. At the end of the delivery stroke,
the directional valve reverses the low pressure flow to the top of the
low pressure piston, exhausting the oil on the bottom side of the to the
reservoir at atmospheric pressure. During the return stroke, the high-pressure
process discharge check valve is closed and process fluid at the intake
pressure fills the high pressure cylinder, readying it for the next delivery
stroke. This is a convenient arrangement when only a few strokes are to
be used at one time. An electrically operated, oil-pivoted, directional
valve can be used in the place of that manually operated (Figure
1) to provide automatic reciprocation of the single-acting pump, if
desired. The flow is still intermittent, with time lost in pumping down,
or returning the pistons to the bottom of the stroke.
The intensifier portion is shown in Figure 2
with a yoke separating the low and high pressure cylinders. Here, a low
pressure ram separates the low pressure and high pressure pistons. This
ram is larger in diameter than the higher pressure piston, but considerably
smaller than the low pressure piston. With this construction, the high
pressure piston is never exposed to oil in the low pressure cylinder and
thus hydraulic oil cannot contaminate the high pressure process fluid.
Double-Acting Pump
A continuous, uniform flow may be attained with the double-acting arrangement
(Figure 3). While one oft he double-acting pistons
is delivering, the other cylinder is being charged during its intake stroke.
This apparatus is a neat, compact assembly (Figure 4,
A and B). While these units could be made to stroke with a manual four-way
directional valve, they are almost invariably assembled for fully automatic
operation. Meeting requirements for explosion proofing is conventional
and the application of remote control is simple.
Design Problems in the Hydraulic Drive
In hydraulically driven reciprocating pumps like those discussed previously,
the four principal design problems, common to all reciprocating piston
pumps, are made easier. Moreover, these pumps are useful and economical
at far lower pressures than originally planned.
Cyclical Stress
The hydraulic drive makes possible a symmetrical construction that minimizes
stress concentrations that become so critical at high pressure. Since the
high pressure heads of the intensifier are cylindrical and free of geometric
discontinuities, the only stress concentration remaining to cause concern
is that characteristic of thick, hollow cylinders subjected to internal
pressure. This stress system is classical(9) in the theory of
elasticity, and considerable information has been published concerning
various solutions(10, 12). Ordinance engineers for many years
have worked on the same stress problems in guns which deal repeatedly with
peak pressures of 38,000 pounds and higher. The techniques of strengthening
tubes by autofrettage in monobloc cylinders and of compounding tubes by
shrinkage techniques originated in ordnance development. Both these techniques,
either singly or in combination, are used in manufacturing the high pressure
ends of the intensifiers illustrated. Designs illustrated are based on
additional data and on design practices somewhat beyond those reached in
the ordnance field; pressures handled with this equipment are more than
four times the peak pressure of cannon.
The favorably oriented stresses developed by these techniques extend
endurance limits considerably beyond that anticipated with conventional
design. The life of a pressure cylinder, normally limited by fatigue, may
easily be trebled by favorably oriented residual stress. For pressures
as high as 60,000 pounds, apparatus can have practically infinite life;
this may be true even up to 80,000 pounds.
Over a 9-year period the Harwood Engineering Company has been required
to replace only three high pressure heads, two of which were made from
the same heat of steel, and failed in the same way after only a few pressure
cycles. Failure occurred at a localized metallurgical imperfections that
were not revealed at inspection. The third failure resulted from hydrogen
embrittlement when a pump made of alloy steel was used to pump hydrogen
gas. Even though these pumps are in regular use, those operating at extremely
high pressure have experienced only a few thousand cycles. Many of the
pumps designed for 60,000 pounds and lower have had hundreds of thousands
of stress cycles without failure. In fact, several of the pumps designed
for 60,000 pounds have been used continuously for many months in fatigue
studies of other apparatus where the frequencies of application have ranged
from 20 to 50 cycles per minute.
Check Valves
In addition to lessening wear on the check valves by reducing the number
of strokes per unit volume pumped, this hydraulic-drive system eliminates
the stress concentration resulting from cross-drillings for pressure ports
and check valves by the simple device of using external check valves connected
by piping to the high pressure head. Large-diamter pistons and long strokes
make the effect of lost volume (clearance volume) in the short piping to
the check valves even smaller than normally encountered with other pumps.
Ratio of lost volume to total volume of the cylinder — i.e., volume between
the piston and the check valve — determines volumetric efficiency. In any
case, with liquid pumps, volumetric efficiency is usually not a serious
concern. Gases at high pressure tend to behave like liquids; hence, the
problem of volumetric efficiency is not so acute in the final compression
stage as it is in the first stage where the wasted work of compression
is relatively more important.
Packings
Life of a packing is lengthened by lower piston speed, fewer reversals
of direction, as well as the tendency of liquids to be lubricating at high
pressure. Water at room temperature and high pressure is a surprisingly
good lubricant, but at pressures exceeding 135,000 pounds per square inch,
it is likely to freeze(3). Some high pressure packings will
last a fair amount of time with essentially no lubricant in the conventional
sense. Wherever possible, it is, of course, desirable to lubricate packings,
but frequently the task is nearly impossible because of the severe requirements
of processes that cannot tolerate contamination.
Maintenance
Figure 4, B illustrates the ease with which a typical
hydraulically driven pump can be maintained. The check valves, being external,
can be replaced within minutes and then repaired at convenience. The replacement
of high pressure packings and closures is accomplished without a major
disassembly and with simple hand tools. On the low pressure side, the oil
is kept free of foreign particles and abrasive matter by filters. Reasonable
attention to general shop tidiness reduces to an infrequent routine the
necessity for filter and oil changes. It is extremely rare that low pressure
components require service.
General Design Considerations
By using a maximum 2000 pounds per square inch as the driving pressure,
several advantages accrue. Hydraulic oil can be used for the pressure liquid
and the low pressure piston can be packed with piston rings as in automobile
and Diesel engine pistons. The design is straight-forward and time-proved,
and lends itself to long, trouble-free service.
Oil hydraulic pumps and valves for 2000 pounds per square inch are commercially
available. Their familiar and extensive use in quite complex circuitry
is reassuring in evaluating this relatively simple application. By using
filtering practices that have been long and successfully established, there
is no problem in keeping the driving oil free of foreign and abrasive matter.
The variety of functional values for that pressure range, together with
standard instrumentation for process industries, opens many new opportunities
for automatic operation and control.
A metallurgical consideration in the design of the high pressure end
of the pump determines the limit of the piston size. If wall thickness
of the cylinder is excessive, it is impossible in heat treatment, to quench
to a martenistic structure. This structure is essential for maximum strength
and fatigue resistance.
Properties of Fluids at High Pressure
Fluids at high pressure have some interesting properties that must be understood
in designing and operating pumping equipment. For many purposes, unconfined
liquids may be treated as incompressible; yet at 100,000 pounds per square
inch, density of 20% over that at atmospheric pressure is common, and the
heat acquired from the work of compression may be important. On the other
hand, under normal conditions, gases obey Boyle's law approximately, and
change in pressure results in a proportional change in density. At high
pressures, compressibility becomes less, and a correspondingly smaller
portion of the applied energy goes into heating the gas. Some materials
usually thought of as solids appear quite plastic at high pressure. In
fact, some solids are useful in hydrostatic transmission of extremely high
pressure.
The effect of pressure on viscosity is familiar to most who have used
high pressure apparatus. No other property shows so dramatically the effect
of pressure. In some substances, experimenters have observed an increase
of viscosity of a millionfold. Extensive data, however, has been published(1,
4) only recently.
Applications
Powder Compaction
Powder compactions also present some interesting problems. Sometimes, for
example, if a large part of the volume change occurs at low pressure, the
work can be done more efficiently if a moderately high pressure intake
pump is used initially, leaving a smaller task for the high pressure pump.
Figure 5 shows a small double-acting, hydraulically
driven, intake pump used in conjunction with a single-acting, high pressure
pump designed for this purpose. The intake pump operates continuously when
the manually operated directional valve is in the pump-down position, charging
the whole system to the maximum pressure it can generate and setting the
single-acting piston at the bottom of its stroke. When the directional
valve is shifted to the other position, the single-acting intensifier operates.
Thereafter, the intake pump serves only to pump down as required until
the desired pressure is reached. It is obvious that a similar system works
equally well with double-acting intensifiers for pumps.
Gases at High Pressure
Gas compressors are readily available for pressures to 2500 pounds per
square inch, but the choice of apparatus for higher pressures is limited.
Using low pressure compressors in conjunction with high pressure intensifier
pumps makes it possible to compress gases to pressures as high as 200,000
pounds. As pressure increases, gases behave more like liquids, since compressibility
decreases. Two or three stages of intensifier-type pumping, depending upon
the constants of the particular problem, are all that is needed. However,
due attention must be paid to interstage cooling. The physical arrangement
of the intensifier, with its cylindrical high pressure ends, makes the
task of providing interstage cooling quite direct.
Frequently, it is convenient to use the liquid piston principle illustrated
in Figure 6. The liquid rises and falls in the accumulators
with the stroking of the intensifier. The gas is pumped on the liquid interface.
A small amount of liquid is lost by evaporation and carried into the process.
Floating pistons carried on the liquid surface reduce this loss because
a much smaller liquid surface is presented for evaporation. This also minimizes
the solution of the gas by the liquid. The floating pistons are called
upon to seal very small pressure differences, only those required to overcome
their own packing friction. Volumetric efficiency with this arrangement
is high, but the higher the intake pressure at a given stage, the less
important volumetric efficiency becomes.
Uniform Flow
Continuous, uniform flow and pressure are usually desirable in flow processes.
When gases or other relatively compressible fluids are pumped, a considerable
percentage of the stroke is used in compressing the fluid to the discharge
pressure. Under these circumstances, pips, or discontinuities in flow and
pressure at the valve crossover, are noticeable and frequently undesirable.
An arrangement of double-acting intensifiers in a system called pipless
pumping (Figure 7,
A) can eliminate pips in delivery and pressure encountered in pumping
gases or other compressible fluids, or at least can make the pips so small
that, in the smallest practical test-tube sized setup, it is difficult
to record them with standard industrial instrumentation(11).
The two intensifiers used in this system are automatically sequenced
so that while piston A (Figure 7,
A) is delivering fluid, piston C is also moving and compressing the
fluid to a pressure slightly less than that of piston A. At the same time,
piston D completes its suction stroke; and it will next be in the predelivery
compression phase. Thus, each piston in turn, goes through the sequence
of suction, predelivery compression, and delivery. Figure
7, B shows the hydraulic
circuit used to accomplish the above sequences. A similar arrangement is
possible using single-acting intensifiers.
Controllability
Another important advantage of the hydraulic drive is its ready adaptability
to automatic process-control techniques. Crank pumps do not lend themselves
so readily to these techniques because inertia of the large rotating and
reciprocating parts makes fast change in speed difficult.
In the oil-hydraulics industry, a wide variety of functional valves
built to Joint Industrial Conference standards are available. With very
little ingenuity, they can be made to operate in the hydraulic-drive circuitry
in conjunction with process instrumentation to make the pump respond quickly
and smoothly to commands issued by the instrumentation.
The more responsive a pump is to signals from control points in a continually
flowing process, the more precise automatic control is possible. Signals
of flow rate, pressure, or temperature may be used at the control point.
Frequently, all of these signals are fed to a blind controller which commands
the pump. Typical of this sort of control system is the controlled pumping
of the catalyst into a polyethylene reaction. A flow-control valve in the
low pressure drive line is operated by instrument air. The pistons can
be made to take hours for one stroke or to make the stroke in two seconds.
The valve is, of course, arranged to fail safe.
Other functional oil valves such as sequence, directional, relief, pressure-regulating,
or piloted valves which can be made to operate with instrument air, provide
a substantial variety of new methods for process control.
Safety
Controllability of the hydraulically driven pump also permits the designer
to plan several lines of defense against a breakdown. Although energy levels
are seldom high enough to make failures dangerous, safety considerations
are important, especially if the fluid is gas.
The relief valve, the first line of defense, is adjustable and at a
preset pressure will bypass the entire flow in the drive circuit, thereby
simultaneously protecting the components of the drive circuit and limiting
the maximum pressure that the intensifier can generate on its high pressure
side. The relief valve is a protective device only and not a pressure regulator.
It also serves as a guard against mishaps on the high pressure side. Should
the piston be prevented from making a full stroke, the drive pressure would
increase to the preset value, the relief valve would open, and serious
damage would then be avoided. Such an incident could occur, for instance,
if spontaneous polymerization should take place, or if the high pressure
fluid should become, with increasing pressure, so viscous that it could
not flow.
Relief valves and blowout disk assemblies may be placed strategically
in the high pressure circuit for interstage protection. Frequently, it
is more expedient to use spring or pneumatically operated relief valves
for this purpose.
In physical arrangement, high pressure portions of the pump can be barricaded
in one room with the oil hydraulic equipment in another. In other instances,
it is more desirable to barricade all the apparatus and to operate it entirely
by remote control.
Performance
Delivery
The flow from a single-acting pump is shown graphically as a function of
time in Figure 8, A which indicates the time required
to pump down — e.g., to return the piston and to recharge the pump. The
first portion of a pressure stroke is used to take up the compressibility.
while liquids at more familiar pressure levels have so little compressibility
that it can be neglected, in the higher ranges it becomes a consideration.
Essentially continuous, uniform flow may be obtained with double-acting
pumps (Figure 8, B). Ideally, there is delivery
by one end or the other all the time, and none is lost during pumping down.
Actually pips do not exist — i.e., discontinuities in pressure or flow
— their duration depending upon the speed of the spool in the crossover
valve, the capacity of the process into which the pump is delivering, and
on the compressibility of the fluid being pumped. If it is necessary to
make the flow even more uniform, the system of pipless pumping described
earlier may be used.
Capacity
The hydraulic-drive system provides high capacity in relatively small compact
units. If the piston speed is arbitrarily limited to 180 inches per minute,
Figure 9 illustrates approximately the relative
capacities of a series of Harwood double-acting pumps with various sizes
of low-pressure pistons. In this figure, DA means double-acting, and the
number following in inches of the low-pressure driving piston. The horsepower
required to pump at a given pressure and flow rate is shown. The lines
of fixed pressure are intersected by lines of fixed horsepower. They show
the maximum horsepower for a given low pressure piston size limited by
180 inches per minute piston speed.
Table I. Comparison of Pumps (rated at 50,000 pounds/sq.
inch)
|
Model
|
Piston Diameter (inch)
|
Stroke (inch)
|
Displacement per stroke (cubic inches)
|
Displacement Ratio
|
Piston Thrust (tons)
|
Crankb Torque (inch-tons)
|
| DA-14 |
2.50
|
18
|
88.36
|
1.0a
|
123
|
1107
|
| DA-10 |
1.75
|
10
|
24.05
|
3.67
|
60
|
300
|
| DA-7 |
1.25
|
8
|
9.82
|
9.0
|
31
|
124
|
| DA-4 |
0.75
|
6
|
2.65
|
33.3
|
11
|
33
|
Since all these pumps have models rated for 50,000 pounds per square
inch, that pressure range is interesting for comparison (Table
I). The piston proportions of the DA-4 model are not far from those
that might be used in a crank-driven pump. The desirability of the large
diameter piston with long stroke is shown by the ratio of the number of
strokes, required of the smaller unit, to displace the same volume as one
stroke of the DA-14 model.
Design difficulties in making crank-driven pumps with large pistons
and strokes is clearly shown in the column of piston thrust. These problems
are brought into even sharper focus when the torque (Table
I) per piston that would exist in crank-driven pumps if pistons were
of the same size and stroke used in the hydraulically driven pumps is considered.
Efficiency
Efficiency of a hydraulically driven pump is the product of efficiencies
for the low pressure oil-hydraulic pump, the intensifier, and pressure
transmission by oil in the piping. Efficiency of the oil-hydraulic pump,
depending upon its design and operating pressure, is high, about 85%, and
that of the intensifier itself is greater than 95%. With properly proportioned
piping, friction losses are low. An overall efficiency, greater than 80%
for the hydraulically driven pump is generally the rule.
References
(1) Am. Soc. Mech. Engrs., New York, "Viscosity and Density
of Over 40 Lubricating Fluids of Known Composition at Pressures to 150,000
Pounds per Square Inch and Temperatures to 425°F," vol. 2, p. 114,
1953.
(2) Bridgman, P. W., "Physics of High Pressure," p. 30,
G. Bell & Sons, London, 1952.
(3) Ibid., p.233.
(4) Bridgman, P. W., Proc. Am.Acad. Arts Sci., vol.
77, No. 4, 115 (February, 1949).
(5) Comings, E. W., "High Pressure Technology," McGraw-Hill,
New York, 1956.
(6) Dodge, B. F., Trans. Am. Soc. Mech. Engrs. 75,
331, (April 1953).
(7) Hayes, T. J., "Elements of Ordnance," p. 164, Wiley,
New York, 1938.
(8) Joint Industrial Conference, "Hydraulic Standards for
Industrial Equipment," General Motors Corp., January 1953.
(9) Lame, G., "Lecons sur la théorie (mathématique)
de l'élasticité (des corps solides)," Paris, 1852.
(10) Love, A. E. H., "Mathematical Theory of Elasticity,"
4th ed., p. 144, Cambridge Univ. Press, Cambridge, Eng., 1927.
(11) Newhall, D. H., U.S. Patent pending to be assigned
to Harwood Engineering Company, Inc., Walpole, Mass.
(12) Seely, F. B., Smith, J. O., "Advanced Mechanics of
Materials," 2nd ed., p. 334, Wiley, New York, 1952.
(13) Timoshenko, S., Goodier, J. N., "Theory of Elasticity,"
2nd ed., p. 59, McGraw-Hill, New York, 1951.
(14) Tongue, H., "High Pressure Chemical Plant," p. 51,
Van Nostrand, New York, 1934.
Harwood Engineering Company, Incorporated – © May, 1999
455 South Street, Walpole, Massachusetts 02081
phone: 508-668-3600
fax: 508-660-2276
http://www.ultranet.com/~harwood/
email: harwood@ma.ultranet.com
