3.1 Hydraulic Pump

3.1 Hydraulic pumps 2012

Hydraulic and pneumatic control lecture notes by Siraj K. Page 1

Chapter three: Hydraulic system devices

31. Hydraulic pump

3.1.1. Introduction

A pump, which is the heart of a hydraulic system, converts mechanical energy in to hydraulic

energy. The mechanical energy is delivered to the pump via a prime mover such as an electric

motor. Due to mechanical action, the pump creates a partial vacuum at its inlet. This permits

atmospheric pressure to force the fluid through the inlet line into the pump. The pump then pushes

the fluid into the hydraulic system. There are two broad classifications of pumps as identified by

the fluid power industry.

1. Dynamic (non-positive displacement) pumps. This type is generally used for low-pressure,

high-volume flow applications. Because they are not capable of withstanding high pressures, they

are of little use in the fluid power field. Normally their maximum pressure capacity is limited to

250-300 psi. This type of pump is primarily used for transporting fluids from one location to

another. The two most common types of dynamic pumps are the centrifugal pump and the axial

flow propeller pumps.

2. Positive displacement pumps. This type is universally used for fluid power systems. As the

name implies a positive displacement pump ejects a fixed amount of fluid into the hydraulic system

per revolution of pump shaft rotation. Such a pump is capable overcoming the pressure resulting

from the mechanical loads on the system as well as the resistance to flow due to friction. These are

two features that are desired of fluid power pumps. These pumps have the following advantages

over non-positive displacement pumps:

a. High pressure capability (up to 12,000 psi)

b. Small, compact size

c. High volumetric efficiency throughout the design pressure range

d. Great flexibility of performance (can operate over a wide range of pressure requirements and

speed ranges)

There are three main types of positive displacement pumps: gear, vane and piston. Many variations

exist in the design of each of these main types of pumps. For example, vane and piston pumps can

be either fixed or variable displacement. A fixed displacement pump is one in which the amount of

fluid ejected per revolution (displacement) cannot be varied. In a variable displacement pump, the

displacement can be varied by changing the physical relationships of various pump elements. This



change in pump in displacement produces a change in pump flow output even though pump speed

remains constant.

It should be understood that pumps do not pump pressure. Instead they produce fluid flow. The

resistance to this flow, produced by the hydraulic system, is what determines the pressure. For

example, if the positive pump has its discharge line open to the atmosphere, there will be flow, but

there will be no discharge pressure above atmospheric because there is essentially no resistance to

flow. However, if the discharge line is blocked, then we have theoretically infinite resistance to

flow. Hence, there is no place for the fluid to go. The pressure will therefore rise until some

component breaks unless pressure relief is provided. This is the reason a pressure relief valve is

needed when a positive displacement pump is used. When the pressure reaches a set value, the relief

valve will open to allow flow back to the oil tank. Thus, a pressure relief valve determines the

maximum pressure level that the system will experience regardless of the magnitude of the load

resistance.

Some pumps are made with variable displacement, pressure compensation capability. Such pumps

are designed so that as system pressure builds up, they produce less flow. Finally at some

predetermined maximum pressure level, the flow output goes to zero due to zero displacement. This

prevents any additional pressure buildup. Pressure relief valves are not needed when pressure

compensated pumps are used. The hydraulic power developed by pumps is converted back into

mechanical power by hydraulic cylinders and motors, which produce the useful output. Figure 3.1.1

shows a backhoe loader which uses a variable displacement, pressure compensated, axial-piston

pump to provide optimum performance in both backhoe and loader operations. The backhoe portion

of the machine performs operations such as digging a trench. The front loader portion can then be

used to load a dump truck with the earth removed from the trench dug by the backhoe. The pump

delivers the desired flow to the hydraulic cylinders at the required pressure to fulfill implement

demands. At an operating of 2200 rpm, the pump produces a maximum flow of 43 gpm (163 Lpm)

at systems pressures of 3300 psi (22,700 kPa).



Figure 3.1.1 Performance of backhoe loader is enhanced with variable displacement, pressure

compensated, axial-piston pump. (Courtesy of Caterpillar Inc., Peoria, Illinois.)

3.1.2. Pumping theory

Pumps operate on the principle whereby a partial vacuum is created at the pump inlet due to the

internal operation of the pump. This allows atmospheric pressure to push the fluid out of the oil tank

(reservoir) and into the pump intake. The pump then mechanically pushes the fluid out the

discharge line. This type of operation can be visualized by referring to the simple piston pump of

figure 3.1.2. Note that this pump contains two ball check valves, which are described as follows:

• Check valve 1 is connected to the pump inlet line and allows fluid to enter the pump only at

this location.

• Check valve 2 is connected to the pump discharge line and allows fluid to leave the pump

only at this location.

As the piston is pulled to the left, a partial vacuum is generated in pump cavity 3, because the close

tolerance between the piston and cylinder (or the use of piston ring seals) prevents air inside cavity

4 from travelling in to cavity 3. This flow of air, if allowed to occur, would destroy the vacuum.

This vacuum holds the ball of check valve 2 against its seat (lower position) and allows atmospheric

pressure to push fluid from the reservoir into pump via check valve 1. This inlet flow occurs

because the force of the fluid pushes the ball of check valve 1 off its seat. When the piston is pushed

to the right, the fluid movement closes inlet valve 1 and opens outlet valve 2. The quantity of fluid,

displaced by the piston, is forcibly ejected out the discharge line leading to the hydraulic system.

The volume of oil displaced by the piston during the discharge stroke is called the displacement

volume of the pump.



Figure 3.1.2 Pumping action of a simple piston pump.

3.1.3. Pump classification

Dynamic pumps

The two most common types of dynamic pumps are the centrifugal (impeller) and axial (propeller)

pumps shown in figures 3.1.3. Although these pumps provide smooth continuous flow, their flow

output is reduced as circuit resistance is increased and thus are rarely used in fluid power systems.

In dynamic pumps there is a great deal of clearance between the rotating impeller or propeller and

the stationary housing. Thus as the resistance of the external system starts to increase, some of the

fluid slips back in to the clearance spaces, causing a reduction in the discharge flow rate. This

slippage is due to the fact that the fluid follows the path of least resistance. When the resistance of

the external system becomes infinitely large (for example, a valve is closed in the outlet line), the

pump will produce no flow. These pumps are typically used for low-pressure, high volume flow

applications. Also since there is a great deal of clearance between the rotating and stationary

elements, dynamic pumps are not self-priming unlike positive displacement pumps. This is because

the large clearance space does not permit a suction pressure to occur at the inlet port when the pump

is first turned on. Thus if the fluid is being pumped from a reservoir located below the pump,

priming is required. Priming is the prefilling of the pump housing and inlet pipe with fluid so that

the pump can initially draw in the fluid and pump it efficiently.



Figure 3.1.3. Dynamic (non-positive displacement) pumps. (Courtesy of Sperry Vickers, Sperry

Rand Corp., Troy, Michigan.)

Figure 3.1.4 (a) shows the construction features of centrifugal pump, the most commonly used type

of dynamic pump. The operation of a centrifugal pump is as follows. The fluid enters at the centre

of the impeller and picked up by the rotating impeller. As the fluid rotates with the impeller, the

centrifugal forces causes the fluid to move radially outward. This causes the fluid to flow through

the outlet discharge port of the housing. One of the interesting characteristics of a centrifugal pump

is its behavior when there is no demand for fluid. In such a case, there is no need for a pressure

relief valve to prevent pump damage. The tips of the impeller blades merely slosh through the fluid,

and the rotational speed maintains a fluid pressure corresponding to the centrifugal force

established. The fact that there is no positive internal seal against leakage is the reason that

centrifugal pump is not forced to produce flow against no demand. When the demand for the fluid

occurs (for example, the opening of the valve), the pressure delivers the fluid to the sources of the

demand. This is why centrifugal pumps are so desirable for pumping stations used for delivering

water to homes and factories. The demand for water may go to near zero during the evening and

reach a peak sometimes during the day time. The centrifugal pump can readily handle these large

changes in fluid demand.



Although dynamic pumps provide smooth continuous flow (when a demand exists), their output

flow rate is reduced as resistance to flow is increased. This is shown for centrifugal pumps in figure

3.4(b), where pumps pressure is plotted versus pump flow. The maximum pressure is called the

shutoff head because an external circuit valve is closed when shutoff the flow. As the external

resistance decreases due to the valve being opened, the flow increases at the expense of reduced

pressure.

Figure 3.1.4. Centrifugal pump. (a) Construction features. (b) Pressure versus flow curve

Positive displacement pumps

This type of pumps ejects a fixed quantity of fluid per revolution of the pump shaft. As a result,

pump output flow, neglecting changes in the small internal leakage, is constant and not dependent

on the system pressure. This makes them particularly well suited for fluid power systems.

However, positive displacement pumps must be protected against overpressure if the resistance to

flow becomes very large. This can happen if a value is completely closed and there is no physical

place for the fluid to go. The reason for this is that a positive displacement pumps continues to

ejects fluid (even though it has no place to go), causing an extremely rapid buildup in pressure as

the fluid is compressed. A pressure relief valve is used to protect pump against overpressure by

diverting pump flow back to the hydraulic tank, where the fluid is stored for system used.



Positive displacement pumps can be classified by the type of motion of internal elements. The

motion may be either rotary or reciprocating. Although these pumps come in a wide variety of

different designs, there are essentially three basic types:

1. Gear (fixed displacement only by geometrical necessity).

a. External gear pumps

b. Internal gear pumps

c. Lobe pumps

d. Screw pumps

2. Vane pumps

a. Unbalanced vane pumps (fixed or variable displacement)

b. Balanced vane pumps (fixed displacement only)

3. Piston pumps (fixed or variable displacement).

a. Axial design

b. Radial design

In addition, vane pumps can be of the balanced or unbalanced design. The unbalanced design can

have pressure compensation capability, which automatically protects the pump against

overpressure.

1. Gear pumps

External gear pump

Figure 3.1.5 illustrates the operation of an external gear pump, which develops flow by carrying

fluid between the teeth of two meshing gears. One of the gears is connected to a drive shaft

connected to the prime mover. The second gear is driven as it meshes with the driving gear. Oil

chambers are formed between the gear teeth, the pump housing and the sides wear plates. The

suction side is where teeth come out of mesh, and it is here that the volume expands, bringing about

a reduction in pressure below atmospheric pressure. Fluid is pushed into this void by the

atmospheric pressure because the oil supply tank is vented to the atmosphere. The discharge side is

where teeth go into mesh, and it is here that the volume decreases between mating teeth. Since the

pump has a positive internal seal against leakage the oil is positively ejected into the outlet port.



Figure 3.1.5. External gear pump operation. (Courtesy of Sperry Vickers, Sperry Rand Corp., Troy

Michigen.)

Volumetric displacement and theoretical flow rate

The following analysis permits us to evaluate the theoretical flow rate of a gear pump using

specified nomenclature:

=0D Out side diameter of gear teeth (in, m)

=iD Inside diameter of gear teeth (in, m)

=L Width of gear teeth (in, m)

=DV Displacement volume of pump (in3/rev, m3/rev)

=N Rpm of pump

=TQ Theoretical pump flow-rate

The volumetric displacement of a gear pump can be found by calculating the volume of a hollow

cylinder of outside diameter 0D and inside diameteriD , where the length of the cylinder is L .

There are actually two such cylinder volumes (because there are two gears) where oil could fill the

inside of the pump if there were no gear teeth. However one half of these two volumes is taken up

by the gear teeth of both gears. Thus the volumetric displacement can be represented by equation

(3.1).



( )LDDV iD22

04−= π

3.1.1

The theoretical flow-rate (in English units) is determined next:

( ) ( ) ( )min//min/ 33 revNrevinVinQ DT ∗=

Since ,2311 3ingal = we have

( ) ( ) ( )231

min//3 revNrevinVgpmQ D

T

∗= 3.1.2

Using metric units, we have

( ) ( ) ( )min//min/ 33 revNrevmVmQ DT ∗= 3.1.2M

Equations (3.1.2 and (3.1.2M) show the pump flow varies directly with speed as graphically

illustrated in figure 3.1.6 (a). Hence, the theoretical flow is constant at a given speed, as shown by

the solid line in figure 3.1.6 (b).

Figure 3.1.6. Positive displacement pumps Q versus N and P versus Q curves.

Volumetric efficiency

There must be a small clearance (about 0.001 in) between the teeth tip and pump housing. As a

result, some of the oil at discharge port can leak directly back toward the suction port. This means

that the actual flow-rate AQ is less than the theoretical flow rateTQ , which is based on volumetric

displacement and pump speed. This internal leakage, called pump slippage, is identified by the term



volumetric efficiency υη which equals about 90% for positive displacement pumps, operating at

design pressure:

T

A

Q

Q=υη 3.1.3

The higher the discharge pressure, the lower the volumetric efficiency because internal leakage

increases with pressure. This is shown by the dashed line in figure 3.1.6 (b). Pump manufacturers

usually specify volumetric efficiency at the pump rated pressure. The rated pressure of positive

displacement pump is that pressure below which no mechanical damage due to overpressure will

occur to the pump and the result will be a long reliable service life. Too high a pressure not only

produces excessive leakage but also can damage a pump by distorting the casing and overloading

the shaft bearings. This brings to mind once again the need for overpressure protection. High

pressures occur when a high resistance to flow is encountered such as actuator load or a closed

valve in the pump outlet line.

Figure 3.1.7 shows detailed features of an external gear pump. Also shown is the hydraulic symbol

used to represent fixed displacement pumps in hydraulic circuits. This external gear pump uses a

spur gear (teeth are parallel to the axis of the gear), which are noisy at relatively high speeds. To

reduce noise and provide smoother operation, helical gears (teeth inclined at a small angle to the

axis of the gear) are sometimes used. However, these helical gear pumps are limited to low-pressure

applications (below 200 psi) because they develop excessive end thrust due to the action of the

helical gears. Herringbone gears pumps eliminate this thrust action and thus can be used to develop

much higher pressure (up to 3000 psi). herringbone gears consist basically of two rows of helical

teeth cut into one gear. One of the rows of each gear is right-handed and the other is left-handed to

cancel out the axial thrust force. Herringbone gear pumps operate as smoothly as helical gear pumps

and provide greater flow rates with much less pulsating action.



Figure 3.1.7. Detailed features of an external gear pump. (Courtesy of Webster Electric Company,

Inc., subsidiary of STA-RITE Industries, Inc., Racine, Wisconsin)

Internal gear pump

Figure 3.1.8 illustrates the configuration and operation of the internal gear pump. The design

consists of an internal gear, a regular spur gear, a crescent-shaped seal and an external housing. As

power is applied to either gear, the motion of the gear draws fluid from the reservoir and forces it

around both sides of the crescent seal which acts a seal between the suction and discharges ports.

When the teeth mesh on the opposite to the crescent seal, the fluid is forced to enter the discharge

port of the pump. Figure 3.1.9 provides a cutaway view of an internal gear pump that contains its

own built-in safety relief value.



Figure 3.1.8. Operating of an internal gear pump. (Courtesy of Sperry Vickers, Sperry Rand Corp.,

Troy, Michigan.)

Figure 3.1.9. Cutaway view of an internal gear pump with built-in safety relief valve. (Courtesy of

Vicking Pump Division of Houdaile Industries Inc., Cedar Falls, Iowa.)

Lobe pump

Also in the general family of gear pumps is the lobe pump, which is illustrated in figure 3.1.10. this

pump operates in a fashion similar to the external gear pump. But unlike the external gear pump,

both lobes are driven externally so that they donot actually contact each other. Thus they are quieter

than other types of gear punps.



Figure 3.1.10. Operation of lobe pump. (Courtesy of Sperry Vickers, Sperry Rand Corp., Troy,

Michigan.)

Vane pumps

Basic design and operation

Figure illustrates the operation of a vane pump. The rotor, which contains radial slots, is splined to

the drive shaft and rotates inside a cam ring. Each slot contains a vane designed to mate with the

surface of the cam ring as the rotor turns. Centrifugal force keeps the vanes out against the surface

of the cam ring. During one-half revolution of rotor rotation, the volume increases between the

rotor and cam ring. The resulting volume expansion causes a reduction of pressure. This is the

suction process, which causes fluid to flow through the inlet port and fill the void. As the rotor

rotates through the second half revolution, the surface of the cam ring pushes the vanes back into

their slots and the trapped volume is reduced. This positively ejects the trapped fluid through the

discharge port.

Analysis of volumetric displacement

Careful observation of figure 3.1.11 will reveaal that there is an eccentricity between the centrline

of the rotor and the centerline of the cam ring. If the eccentricity is zero, there will be no flow. The

following analysis and nomenclature is applicable the vane pump:

=CD Diametere of cam ring (in, m)

=RD Diametere of Rotor (in, m)

=L Width of Rotor (in, m)



=DV Pump volumetric displacement (in3, m3)

=e Eccentricity (in, m)

=maxe Maximum possible eccentricity (in, m)

=maxDV Maximum possible volumetric displacement (in3, m3)

From geometry, we can find the maximum possible eccentricity:

2maxRC DD

e−

=

This maximum value of eccentricity produces a maximum volumetric displacement:

( )LDDV RCD22

max 4−= π

Noting that we have the difference between two squared terms yields

( )( )LDDDDV RCRCD −+=4max

π

Substituting the expression for maxe yields

( )( )LeDDV RCD maxmax 24

+= π

The actual volumetric displacement occurs when :max ee =

( )eLDDV RCD +=2max

π 3.1.4

Figure 3.1.11. Vane pump operation. (Courtesy of Sperry Vickers, Sperry Rand Corp., Troy,

Michigan.)



Piston pump

Fixed displacement piston pumps can be classified from the arrangement of the piston and drive

mechanism and include axial, radial, and in-line crankshaft driven pumps. Axial piston pumps are

further classified as in-line or bent axis drive piston pumps. Two in-line axial piston pump designs

are common: the swash plate design and the check valve design.

In-line axial piston pumps operate at pressures to 3,000 lbf/in2 with high volumetric and overall

efficiencies (above 96% and 85% respectively) because of the close fit between the reciprocating

pistons and the cylinder bores. While speeds in the 3,000-4,000 rpm range are nominal for industry,

some airborne applications are capable of speeds over 15,000 rpm. Component parts of the swash

plate design include the inlet and outlet ports, pump housing, valve plate, rotating group, shaft seal ,

bearings, and drive shaft (Fig. 3.1.12). The rotating group consists of the splined rotating cylinder

block and spherical washer, cylinder block spring and force transmitting pins, pistons supported by

shoes, swash plate, and retainer plate. Pumping action occurs as the rotating group causes the

pistons to reciprocate as they follow the angle of the stationary swash plate. As the pistons retract in

the cylinder bores, the rotating group passes over the kidney-shaped inlet port in the valve plate and

fluid is inducted into the cylinder bore. Further rotation causes the piston to extend into the cylinder

bore as the rotating group passes over the kidney-shaped outlet port in the valve plate and fluid is

expelled.

Figure 3.1.12 Fixed displacement swash plate piston pump (co urtesy of Sperry Vickers ).



Delivery from the pump (Q) is determined by the number of cylinders (n), the speed of the pump

(Nt), the bore (db) and the stroke (5). Stroke, in turn, is dependent on the angle of the swash plate.

For fixed displacement axial piston pumps, 30° is a common swash plate angle.

Fixed displacement bent axis pumps reciprocate the pistons in the rotating cylinder block through a

bevel gear mechanism. The axis is tilted near the junction of the drive shaft and cylinder block.

Displacement angles of 15°, 20°, 25°, 30°, and 40° are standard. These pumps operate at high

efficiencies and have received widespread application in the aerospace industry. Units range in sizes

from 2 to 45 lbf and flow rates of .25 to 150 gal /min at pressures to 5,000 lbf/in2. The speed range

is from 100 rpm to 12,000 rpm, depending on the size and application. Components include the

rotating group, valve plate, and antifriction bearing s. and aluminum case (Fig. 3.1.13).

Figure 3.1.13 Fixed displacement bent axis pump (courtesy of Volvo of America Corporation ).

Pump performance

Introduction

The performance of pump is primarily a function of the precision of its manufacture. Components

must be made to close tolerances, which must be maintained while the pump is operating under

design conditions. The maintenance of close tolerances is accomplished by designs that have

mechanical integrity and balanced pressures. Theoretically the ideal pump be one having zero



clearance between all matting parts. Although this is not feasible, working clearances should be as

small as possible while maintaining proper oil films for lubrication between rubbing parts.

Pump efficiencies

Pump manufacturers run tests to determine performance data for their various types of pumps. The

overall efficiency of a pump can be computed by comparing the hydraulic power out put of the

pump to the mechanical input power supplied by the prime mover. Overall efficiency can be broken

into two distinct components called volumetric efficiency and mechanical efficiency. These three

efficiencies are discussed below.

1. Volumetric efficiency (ηv). Volumetric efficiency indicates the amount of leakage that takes

place within the pump. This involves considerations such as manufacturing tolerances and

flexing of the pump casing under design pressure operating conditions:

T

Av Q

Q

produceshouldpumprateflowltheoretica

pumpbyproducedrateflowactual=

−−

=η 3.1.5

Volumetric efficiencies typically run from 80% to 90% for gear pumps, 82% to 92% for vane

pumps, and 90% to 98% for piston pumps. Note that when substituting efficiency values into

equations, decimal fraction values should be used instead of percentage values. For example, an

efficiency value of 90% would be represented by the value of 0.90.

2. Mechanical efficiency (ηm). Mechanical efficiency indicates the amount of energy losses that

occur for reason other than leakage. This indicates friction in bearings and between other

matting parts. It also includes energy losses due to fluid turbulence. Mechanical efficiencies

typically run from 90% to 95%.

pumptodeliveredpoweractual

leakagenogassupoweroutputpumpm

min=η

Using English units and horsepower for power yields

63000/

1714/

NT

pQ

A

Tm =η 3.1.6

In metric units, using watts for power,

NT

pQ

A

Tm =η 3.1.6M



The parameters of equations (5.8) and (5.8M) are defined as follows in conjunction with figures

3.31.

=p Pump discharge pressure (psi, Pa)

=TQ Pump theoretical flow-rate (gpm, m3/min)

=AT Actual torque delivered to the pump (in. lb, N.m)

=N Pump speed (rpm, rad/s)

A

Tm T

T

pumptodeliveredtorqueactual

pumpoperatetorequiredtorquelTheoretica==η 3.1.7

Note that the theoretical torque required to operate a pump (TT) is the torque that would be required

if there were no leakage. Equations for calculating the theoretical torque and the actual torque are

as follows:

Theoretical torque

( ) ( ) ( )π2

3 psipinVlbinT D

T

∗=∗ 3.1.8

Or

( ) ( ) ( )π2

3 PapmVmNT D

T

∗=∗ 3.1.8M

Actual torque

( )rpmN

pumptodeliveredhorsepowerActualTA

000,63∗=

3.1.9

( )sradN

pumptodeliveredActualTA /

= 3.1.9M

Where

( ) ( )rpmNsradN60

2/

π=

3. Overall efficiency (η0). The overall efficiency considers all energy losses and hence is

defined as follows:


pumpbydeliveredpowerlTheoreticaefficiencyoverall =)( 0η 3.1.10

The overall efficiency can also be represented mathematically as follows:



mv ηηη ∗=0 3.1.11

Substituting equations (3.7) and (3.8) into equation (3.13), we have (for English units):

000,63/

1714/0 NT

pQ

Q

Q

A

T

T

Amv ∗=∗= ηηη

Canceling like terms yields the desired result showing the equivalency of equation (3.1.10) and

equation (3.1.11).

pumptodeliveredhorsepoweractual

pumpbydeliveredhorsepoweractual

NT

pQ

A

A ==000,63/

1714/0η 3.1.12

Repeating this substitution for metric units using equations (3.7), (3.8M) and (3.13) yields:


pumpbydeliveredpoweractual

NT

pQ

A

A ==0η 3.1.12M

Note that the actual power delivered to a pump from a prime mover via a rotating shaft is called

brake power and actual power delivered by a pump to the fluid is called hydraulic power.

Pump selection

Pumps are selected using two basic criteria: (1) those relating to the actuator; and (2) those relating

to the operation and efficiency of the pump itself. The basic parameter relating to the actuator is

volume flow rate. Those relating to the operation and efficiency of the pump include system

pressure, drive speed, cost, reliability, maintenance, and noise. A system pressure should be selected

that will provide existing and future actuators with the force necessary to move the load resistance.

Higher system pressures allow for smaller system actuators and higher efficiencies, but are less

tolerant of contaminants in the system and other lower quality components. Working pressure

should also be compatible with the rest of the system should exchange of components be necessary.

Volume flow rates are selected to provide the necessary fluid and power to move specified load

resistances required distances within available time limits.

1714

QpPhp

∗= 3.1.13

That is, must be equal to or greater than the anticipated output of the system.

The next parameter to be defined is drive speed. Standard speeds of 1,200 and 1,800 rpm are

common. Higher speeds are associated with higher pump efficiencies, but noise, vibration, and



cavitation are usually associated with high suction lifts and air entering the system. Accepted

practice places the pump as close to the reservoir as possible and provides for cooling of the fluid

(adequate reservoir size and design) as well as filtration. If HWCF fluids are used, the reservoir is

placed above the pump inlet.

Another consideration is whether the pump will be of the non positive or positive displacement

type, and, if positive, whether the displacement will be fixed or variable. System pressure is used to

determine whether a non positive displacement pump should be used, whereas cost and circuitry are

used to determine if the pump should have a fixed or variable displacement. Fixed displacement

pumps require pressure relief and flow control valves to set the pressure and flow limits and prevent

component rupture. Variable displacement pumps, on the other hand, may be pressure-, flow-, or

load-compensated to limit system pressure and volume flow rate simultaneously.

Cost factors, system reliability, and required maintenance level are used to determine if the pump

should be of the gear, vane, or piston type, although gear and vane types are usually associated with

lower pressures. Some vane pumps have replaceable cartridges that allow rebuilding within minutes

without removing the pump from the circuit. Gear pumps are usually less expensive and more

tolerant of adverse conditions such as contaminants in the fluid, but have lower volumetric and

overall efficiencies than vane pump counterparts. Piston pumps offer maximum volumetric and

overall efficiency, as well as reliability, but cost is more than proportionally higher.

Selection of an appropriate pumping unit typically follows the sequence:

1. Determine the flow rat e.

2. Select an appropriate drive speed, direction and unit.

3. Circuit conditions determine whether the pump will be of the positive or non positive type, and if

positive, if the displacement will be fixed or variable. This will depend to a large extent upon the

circuit selected.

4. Determine system pressure.

5. Select an appropriate reservoir and plumbing, including the use of a heat exchanger, if necessary.



6. Compute cos t and other factors such as noise levels, vibration, wear characteristic s, fluid

filtration and periodic maintenance schedule necessary to assure the longest possible life of the

pump.

Pump Specification

The following list shows the basic specifications that should be available to specify the pump

precisely:

• Size (displacement)

• Speed (maximum and minimum speeds)

• Maximum operating pressure (continuous/intermittent)

• For open/closed circuit

• Direction of rotation (viewed to shaft end; clockwise [R], counterclockwise [L])

• Controller (for variable displacement pumps)

• Seals (oil)

• Drive shafts

• Port connections

• Mounting type

• External dimensions

• Installation position

• Operating temperature range

• Further details in clear text

Pump noise

Noise from pumps, measured on the decibel db(A) scale or sones scale, may be used over time to

evaluate performance and service conditions. The relationship between sound level s on the sones

and db (A) scale is seen in Table 3.1.1. Increases in noise usually indicate increased wear and

imminent failure, unless a pipe or tube can be found vibrating against a machine or mounting.

Sometimes the coupling is misaligned where the pump connects to the drive motor. While noise has

been traditionally associated with pump wear and failure, recent redefinition of the term as

unwanted or excessive sound recognizes the phenomenon as a form of environmental degradation.



Department of Labor standards for noise levels in work areas also must be considered when

specifying pumps, with or without enclosures, to meet occupational safety and health standards.

Pump vibration is influenced by both mechanical out of balance forces associated with the pump de

sign and by the surges of fluid which exit the outlet port in the case of positive displacement pumps.

The frequency of pump vibration for piston pumps is defined from the cycle rat e (N ) of the pump

and the number of pistons (n) such that

( ) ( ) ( )npistonofNoNrateCycleffrequencyPump .∗=

And

nNf ∗= 3.1.14

Pressure impulses or ripples in the system resulting from fluid surges from the pump have been

shown to be equal to the pump frequency for pumps with an even number of pistons and to twice

the pump frequency for pumps with an odd number of pistons.

Table 3.1.1 comparison of sones vs. DB (A) sound scales (Courtesy of Sperry Vickers).

Documents

3.1 Hydraulic Pump