Mechanical Seals

GBH Enterprises, Ltd.

Engineering Design Guide:
GBHE-EDG-MAC-5702

Mechanical Seals

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Engineering Design Guide:
CONTENTS

Mechanical Seals
SECTION

0

SCOPE

0

1

INTRODUCTION

1

2

ADVANTAGES

2

3

DISADVANTAGES

3

4

BASIC PRINCIPLES

4

5

BALANCE

5

6

LUBRICATION

6

7

MATERIALS

7

7.1

Seal Faces

7.2

Secondary Seal

8

HEAT REMOVAL

8

9

DOUBLE MECHANICAL SEALS

9

9.1

THE 'TANDEM' DOUBLE SEAL ARRANGEMENT

9.2

THE 'BACK TO BACK' DOUBLE SEAL ARRANGEMENT


10

EXTERNAL QUENCH FACILITIES

10

11

APPLICATIONS NOT SUITABLE FOR MECHANICAL
SEAL

11

11.1
11.2

12

13

Emergency Duties - Fire Pumps Etc.
Boiler Feed Pumps Incorporating a Balance Valve

RULES FOR THE APPLICATION OF MECHANICAL
SEALS

12

DEALING WITH SEAL FAILURES

13

SELECTED BIBLIOGRAPHY

APPENDIX A -

MECHANICAL SEAL BALANCE

APPENDIX B -

USE OF TANDEM SEALS ON LPG OR OTHER
HAZARDOUS FLUIDS


0

SCOPE

This Engineering Design Guide presents the basic principles associated with the
use of mechanical seals for the benefit of those involved in the design, selection
and operation of pumps and agitators.
It does not attempt to advise on the particular make or type of seal for any
particular application. For this latter purpose reference shall be made to the
manufacturer's catalogues and difficult applications should be discussed both
with the pump vendor and the seal maker.

1

INTRODUCTION

A mechanical seal is a device with which a seal is affected on a rotating shaft by
close contact between two rigid faces normal to the axis of rotation.
Mechanical seals are precision engineering components and care has to be
taken in selection, in the design of the mounting arrangements - including
provision for cooling, and fitting if the design aims are to be realized. In general
pump manufacturers have only limited operating experience with mechanical
seals and their seal selection is usually based on seal catalogues, frequently
without discussion of the application with the seal manufacturer. If the application
is discussed with seal manufacturer, the latter's advice is normally accepted
without criticism.
Seals are easily damaged by dry-running or by the ingress of abrasive solid
particles. The latter is a particular hazard during the commissioning of new plant
and special precautions may have to be taken at this time to prevent premature
failures.
This note is based on operating experience with mechanical seals in chemical
plants. It is intended to give guidance on the integration of seals into the system
to ensure maximum reliability.


2

ADVANTAGES

The advantages of mechanical seals over packed glands are:
(a)

Reduced maintenance costs - but only if a life of some months operation
is achieved.

(b)

Reduced leakage when in good condition - hence plant areas containing
pumps handling flammable liquids can generally be classed as Zone 2
hazardous areas and not Zone 1 hazardous areas.

(c)

They require less space - hence the pump design can be more compact
with less shaft overhang etc. (It is unfortunately true that few pump
manufacturers take advantage of this since many pumps are still designed
around packed glands).

3

DISADVANTAGES

The disadvantages of mechanical seals compared with packed glands are:
(a)

More expensive - both in first cost and in replacement parts.

(b)

Require a higher standard of fitting.

(c)

Will tolerate less "run-out" of the shaft – especially axially.

(d)

Provide no support or "steadying" to the shaft.

(e)

CAN fail rapidly necessitating the immediate shutting down of the pump
(although this is the exception rather than the rule)


4

BASIC PRINCIPLES

In the soft-packed gland, sealing is obtained by pressing a flexible 'packing'
against the periphery of the rotating shaft. In the mechanical seal, sealing is
transferred to a contact between a stationary face (the seat) and a rotating face
attached to the shaft (the seal face). One of the faces is allowed limited axial
movement to accommodate wear and is pressed against the seat by light spring
loading (normally 1.4-2 bar) and the pressure of the sealed fluid. Lubrication is
necessary between the seat and seal face to prevent rapid wear during rotation.
Seal design is a compromise between providing adequate lubrication and
avoiding the formation of an excessive lubricant film that would allow a high
leakage path.
Mechanical seals thus bear a close resemblance to thrust bearings, though in
most practical seal applications lubrication tends to be in the mixed lubrication
regime rather than the full fluid regime that is the aim in thrust bearing design.
This means that, even with correct operation, wear of the faces will occur and the
seal will have a limited life though this can be many years under favorable
conditions. Moreover, the sealed fluid that acts as the lubricant will often have
poor lubricating properties so that the seal face materials have to be selected so
that they will operate under these conditions of lubrication without excessive
wear.
The lubrication regime is determined by the speed, viscosity of the sealed fluid,
and the interface pressure. The pressure on the seal faces has thus to be
controlled within fairly narrow limits for satisfactory operation. With sealing liquids
at a pressure of above about 10 bar it is necessary to 'balance' the seal to reduce
the hydrostatic pressure component on the seal faces. There are two limits to
operation; a low speed limit below which a proper lubricating film is not
developed and a high speed limit when the heat generated by viscous shear is
sufficient to destroy the lubricating film. These speed limits cannot be accurately
defined in terms of the operating parameters. In general, the lower limit is not of
concern in centrifugal pump applications; however it can apply to autoclave
agitators and similar applications.


5

BALANCE

The principle of balancing is explained in Appendix A for the normal type of seal
(e.g. a Crane 109). For other types, such as external seals, bellows seals or
seals where the floating element of the seal is stationary the principle is similar.
In all cases consider the various forces acting on the floating element of the seal.
The balance factor defines the extent to which the pressure difference across the
seal affects the mean pressure between the seal faces. In the simple unbalanced
seal because of the need to allow clearance between the shaft and the seal
faces, the balance factor is normally about 1.1 - hence a 1 bar increase in
pressure difference results in an 1.1 bar increase in mean seal face pressure.
If the seal is over-balanced there is a tendency for the high pressure fluid to get
between the faces and 'blow' them apart. For this reason in practice, balance
factors less than about 0.6 are not used. Normal value 0.75.

6

LUBRICATION

For satisfactory operation it is necessary for a lubricating film to be established
between the sealing faces. The mechanism by which this film is generated has
not been completely resolved, though most recent evidence, see Bibliography
Ref (6), suggests that with aqueous fluids it results from microwaviness of one of
the seal faces (the soft face when one is used). This can result from the lapping
process in manufacture but normally is formed during initial running. The
microwaviness is generally a 2- period sinusoidal, but there is also some
evidence for 3- period waviness, possibly resulting from relief of stress in a hard
face that has been machined using a three jaw chuck. No attempt is made by the
seal manufacturers to produce a surface profile of this type, their aim being to
manufacture faces with a high degree of flatness and surface finish. This could
be an explanation for some of the failures that occur when starting a new seal,
breakdown occurring before the required profile is developed by running in.
In practice film thicknesses lie in the range 0.25 to 2~m and the waviness height
is about half this. It is clear from these dimensions that the mechanical seal is a
precision engineering component and has to be treated as such for satisfactory
Performance. Normal practice is to arrange for the higher pressure on the outer
periphery of the seal; in this way centrifugal force tends to act against the
leakage flow.


For the majority of applications it is necessary for the fluid film to be substantially
in the liquid phase. Therefore it is essential that the seal should be surrounded by
liquid and this liquid must be sufficiently removed from its boiling point that it will
not vaporize immediately on entering between the seal faces.
Whilst in most cases the fluid entering between the faces at the outer edge must
be liquid, the fluid leaving the seal at the inner edge is often vapor. The actual
position on the faces at which the phase transition takes place or the manner in
which it takes place is not known. However, attempts to run NORMAL types of
seal deliberately on vapor have generally failed.

7

MATERIALS

7.1

Seal Faces

It is normal to have one of the sealing faces of a low friction material that can
tolerate a certain amount of dry rubbing (e.g. carbon-graphite, filled ptfe) to
minimize wear and heating in the event of intermittent face contact.
It is important for film formation that the faces should be readily wetted by the
sealed liquid. This is a function of the wettability of the face materials and the
surface tension of the liquid.
Wettability characteristics:GOOD
Cast Iron
Ni-resist cast iron
Fused Alumina (ceramic)
Carbon
Silicon Carbide

POOR
ptfe
Stellite
18/8 stainless steel


Typical surface tension values:Water
Mineral oil
Benzene
Hexane

(25°C)
(25°C)
(20°C)
(20°C)

72 dyn/cm
32 dyn/cm
28 dyn/cm
18 dyn/cm

Faces with poor wettability should not be used with liquids having high surface
tensions (e.g. water).
Faces must also be resistant to corrosion by the sealed liquid.
7.2

Secondary Seal

The secondary seal prevents leakage between the floating element of the seal
and the shaft or sleeve (or the housing in the case of a seal with a stationary
floating element). It may be an elastomeric 'a' ring or lip seal, a ptfe or metal
wedge ring or a bellows (rubber, ptfe or metal). It must be chosen to be resistant
to the chemical and thermal environment. Sliding secondary seals provide the
simpler arrangement but are liable to cause fretting damage to the component
against which they slide. This frequently necessitates replacement of the sleeve
whenever the seal is changed. A sprayed alumina coating on the sleeve provides
good resistance to fretting. Alternatively this problem can be avoided by using a
bellows type seal or one in which both elements of the secondary seal are
included in the "cartridge" which is obtained from the seal supplier. Soft bellows
are very limited in application; metal bellows) being usually made by edge
welding together preformed discs, require very sophisticated techniques and
close quality control. Rolled single and multi-stream bellows are becoming
increasingly more popular.

8

HEAT REMOVAL

For many applications (ie except those at low temperatures and pressures when
the fluid is well away from its boiling point) it is necessary to provide deliberate
means of removing the frictional heat generated between the rubbing faces.


Failure to do this may result in either:
(a)

Distortion of the faces leading to breakdown of the fluid film.

(b)

Vaporization of the liquid in the vicinity of the seal.

(c)

Accelerated age hardening of the elastomers used for the secondary seal.

all of which may lead to premature failure.
The simplest method of removing the frictional heat, applicable to most pumps, is
to provide a simple flushing connection from the pump delivery over the seal and
then back to the suction of the pump. In most cases, one of the connections,
(either from the pump delivery to the seal or from the seal to pump suction) can
be arranged internally requiring only one external pipe.
When the pumped liquid is close to its boiling point (say less than 10°C below its
boiling point) a simple flushing connection will generally not suffice and other
measures must be taken. Some possible methods of removing the frictional heat
are listed below in approximate order of increasing complexity.
(d)

A simple flushing connection as described above but with a restriction
arranged in the return passage from the seal back to the pump suction.
Hence the pressure at the seal can be higher than suction pressure thus
suppressing any tendency to boil. This method is useful for applications on
liquids with low boiling points. It has the disadvantage that the pressure at
the seal is increased.

(e)

A hollow seat (e.g. Crane 'CR' type) with a suitable coolant. For hot duties
(e.g. hot oil pumps) this method can be very effective since the frictional
heat is removed at source. However, the coolant must be carefully chosen
as too great a temperature gradient within the seat will cause thermal
distortion. On some duties (e.g. refinery hot oil pumps) saturated steam
can be used as a coolant. The steam must be wet; the cooling is really
done by the water in the steam.

(f)

A simple flushing connection with a cooler in the line from pump delivery
to the seal. This method is effective but the cooler has to remove quite a
lot of heat. In some applications where the high temperature is required for
process reasons this may be uneconomic. Alternatively, over
cooling may result in freezing of high melting point fluids.


(g)

A close loop circulation. local to the seal using a 'pumping ring' available
from seal manufacturers together with a small cooler in the circuit. Since
this 'closed loop' is separate from the main steam, the cooler does not
need to be nearly so large as in method (f) and this system is therefore
attractive on large pumps. However, there is a disadvantage in that a
small leak from the seal. which could be tolerated for a time causes a net
inflow of liquid into the seal space causing an increase in temperature and
hence possibly rapid deterioration of the seal. (Cooling jackets round the
seal chamber are not recommended. To be effective it is necessary to use
a thin steel liner With '0' ring seals inside the casting; the risk of leakage at
what are virtually uninspectable joints is generally not acceptable).

9

DOUBLE MECHANICAL SEALS

Double seals can be used in very difficult applications and properly engineered
can be guaranteed to provide an effective seal. However, because of the
necessary complication in auxiliary equipment double seals should generally be
considered as a last resort. One exception being agitator seals with top entry
shaft where liquid is not present. Double seals with a barrier liquid are usually
employed.
There are two basic arrangements of double seals namely:(a)

The TANDEM arrangement in which the two seals are arranged in series
and both are orientated so as to prevent leakage outwards,

and
(b)

The BACK TO BACK arrangements in which the two seals face in
opposite directions and are orientated so as to prevent leakage from the
intermediate space into the pump on one side or out to atmosphere on the
other.

NOTE: Crane's seal manual also shows a 'face to face'
arrangement. This is NOT a true double seal application. The outer
seal is there for the sole purpose of retaining the coolant in the
hollow 'U' type seat.


9.1

THE 'TANDEM' DOUBLE SEAL ARRANGEMENT

This arrangement is of interest for high pressure applications where the
combination of pressure difference and speed, exceed the capabilities of a single
seal. The pressure difference can be broken down, each seal taking
approximately half the total.
It can also be used with LPG's or other hazardous fluids to provide an immediate
alarm in the event of seal failure and a means of containing the leakage and
conducting it away for safe disposal. In this case the space between the two
seals should be filled with an innocuous fluid such as water/glycol. A suitable
arrangement is shown in Appendix B.
In each case it is necessary to satisfy the following requirements:
(a)

Where the tandem arrangement is used to reduce the pressure on one
seal, the intermediate pressure must be controlled and adequate alarm
and/or trip devices provided to protect the installation against failure of this
pressure control system.

(b)

A reliable means should be provided to indicate if the inner seal starts to
fail and leak, for example by detecting an increase in the intermediate
pressure.

(c)

Each seal must be separately provided with suitable means of removing
the frictional heat as described above.

9.2

THE 'BACK TO BACK' DOUBLE SEAL ARRANGEMENT

This is the more usual double seal arrangement with a separate liquid (the
sealant) supplied to the space between the seals and can be recommended in
the following circumstances:(a)

Handling toxic or highly pungent liquids where it is imperative to prevent
leaks, or;

(b)

In cryogenic applications when it is difficult to prevent vaporization at the
seal, or;

(c)

When the pumped liquid is incompatible with the seal because it is either
very highly corrosive or contains abrasive solid particles.


(d)

Agitators

NOTES:
(1)

Applications where abrasive solid particles are present can often be more
simply satisfied by providing a flushing stream inside the seal of an
acceptable clean fluid.

(2)

In many so called slurry duties, the solid content is completely harmless
and non-abrasive e.g. paraxylene slurry, because it is soft like snow.

For the full advantages to be gained from the 'back to back' double seal
arrangement the following requirement must be satisfied.

(e)

The pressure of the sealant must be maintained at approximately one bar
above the internal pressure at the seal (normally suction pressure) and
generally trip and/or alarm devices should be provided to protect the
installation against failure of this pressure.

(f)

A reliable means should be provided to indicate if the inner seal starts to
fail and leak (e.g. pressure drop or loss of level in a closed system or
increased flow in an open system).

(g)

Means must be provided for the removal of the frictional heat from the two
seals - generally this means a circulation system for the sealant.

The Bingham cryogenic pump (J Type head gear) provides an elegant solution to
these problems. The pump is a vertical type and a chamber is arranged at the
top in which the pump fluid is deliberately caused to exist in the vapor phase (by
not lagging it and in some cases even fitting heaters). The sealant (usually
a light oil) between the two seals is then pressurized by a simple head tank which
is pressure balanced to the vapor space. Leakage of oil from the system is
detected by a float operated level alarm in this head tank. Cooling is achieved by
heat transfer into the pumped fluid.


10

EXTERNAL QUENCH FACILITIES

An external auxiliary gland, a 'lip-type' seal or a simple throttle bush can be fitted
thus providing an enclosed space outside the seal into which a quench fluid can
be fed.
This facility should be provided in the following cases:(a)

Very large pumps (say over 100 kW).

(b)

Hot pumps (say over 1500C).

(c)

Cold pumps (where the boiling point of the liquid is below O0C at
atmospheric pressure).

(d)

When the fluid being pumped contains solids, it solidifies or forms crystals
in contact with air.

(e)

With highly corrosive liquids (Note: water quench should NOT be used
with some very strong acids since the addition of water can increase the
corrosion rate).

(f)

All highly flammable liquids.

(g)

Any duty where the pump cannot conveniently be taken off-line rapidly in
the event of a seal failure.

On all duties handling flammable liquids above their auto-ignition temperature, a
permanent steam quench should be provided. On cold duties operating below
O°C a permanent quench facility using methanol or similar solvent should be
provided to clear ice buildup.
On other duties listed above, the facility should be provided at the pump but not
piped up as a permanent installation.
NOTE:
On LPG installations, the use of tandem seals with a water/glycol mixture
between them (see Clause 9.1) will not only contain the leakage but can also be
arranged to give warning of an inner seal failure.


11

APPLICATIONS NOT SUITABLE FOR MECHANICAL SEALS

The following are a few notable cases where mechanical seals are
not generally recommended.
(a)

Emergency Duties - Fire Pumps etc
Mechanical seals are much more likely to fail during stopping and starting
than in continuous service. Also a mechanical seal can fail rapidly and its
performance changes from practically no leakage to a massive leak in a
short time, although this is not normal. Again, since fire pumps in
particular handle water at ambient conditions, leaks, such as can occur
with a packed gland not in perfect condition, can be tolerated. Mechanical
seals should NOT therefore be fitted to fire pumps, other emergency
service pumps unless these run continuously, and any other pump which
is called upon to operate only occasionally, but when required is essential.

(b)

Boiler Feed Pumps Incorporating a Balance Valve
In many moderate to high pressure boiler feed pumps, particularly those of
the 'ring-section' type of construction (e.g. Mather & Platt 'Plurovane'), the
accumulated thrust from the impellers is resisted by a device known as a
balance valve. In this device, the pressure in a chamber connected to the
pump delivery adjusts itself so that the pressure rating on one side of the
balance valve opposes the thrust from the impellers. For this device to
function it is essential that the pump shaft is otherwise completely free
to float axially. The axial float required for satisfactory operation of the
balance valve and to accommodate wear can be as much as 6 mm (t·,).
Very few mechanical seals will successfully cope with movements of this
order, and therefore mechanical seals are not in general recommended for
this type of pump.

12

RULES FOR THE APPLICATION OF MECHANICAL SEALS

The following section is an attempt to lay down some simple rules for the
application of mechanical seals. Generally the seal manufacturer's advice and
recommendations should be sought and this should not be disregarded without
very sound reasons. However, in some cases, the seal maker may not fully
appreciate the operating conditions, properties of the fluids etc.

Therefore all cases, where we question whether he has made the optimum
choice of materials or arrangement of seal, should be discussed between the
seal maker, the pump vendor and the user.
When any of the following rules or limitations appear to be contravened; it should
be insisted, as a minimum requirement, that the seal maker supports his solution
with evidence of successful operation in closely similar applications where we
question whether he has made the optimum choice of materials or arrangement
of seal, should be discussed between the seal maker, the pump vendor and the
user.
When any of the following rules or limitations appears to be contravened; it
should be insisted, as a minimum requirement, that the seal maker supports his
solution with evidence of successful operation in closely similar applications.
(a)

Limits
A general limitation can be applied that the product of shaft speed in
rev/sec, shaft or sleeve diameter at the seal in mm and differential
pressure in bar should not exceed about 105.

(b)

At differential pressures greater than 10 bar 'balanced' seals should be
used.

(c)

External flushing lines should be at least 10 mm bore to reduce the
possibility of blocking and should be routed so as to avoid air or vapor
locking.

(d)

With volatile liquids, if a single seal is used, the temperature of the liquid in
the immediate vicinity of the seal should be AT LEAST 10°C less than the
boiling point of the liquid at the pressure which exists at the seal. This
should be achieved by one of the methods described in Clause 8.

(e)

With liquids which are solid at ambient temperatures the seal housing and
the clamping plate should be steam jacketed and not just steam traced.

(f)

Generally, double mechanical seals should not be used except as a 'last
resort' because of the complication. A combination of a restrictor and a
mechanical seal is preferred. The mechanical seal seals to atmosphere:
the restrictor, which can be a close clearance bush or a lip seal prevents
excessive leakage of the sealant or flushing fluid into the pump.


(g)

Double seals in the 'back to back' arrangement SHOULD be used
in the following circumstances.
(1)

(2)

Where the pumped liquid is highly corrosive and it is not economic
to make all the external parts of the pump in the vicinity of the seal
of resistant material.

(3)

When the fluid contains abrasive solid particles and it is not
possible to use the simpler arrangement of a flush inside the seal
with a suitable clean fluid. See Clause 12(f) above.

(4)

(h)

Where the pumped liquid is very toxic or gives off highly pungent
odors.

When handling liquefied gases at temperatures below -50°C unless
the particular pump manufacturer can demonstrate successful
experience with a single seal.

Where 'back to back' seals are employed the following requirements are
essential:(1)

The sealant must be compatible with the process, non-toxic, noncorrosive and not be itself a difficult material to seal.

(2)

The sealant must be maintained between the two seals at a
pressure greater than the pressure on either side (usually 1
atmosphere or more).

(3)

If the pressure is controlled by instruments or other devices which
can mal-function, an alarm device should be provided to give
warning if it does fail.

(4)

When necessary, deliberate provision must be made for the
removal of friction heat from both seals. This is most conveniently
done by circulating the sealant.

(5)

If the sealant is circulated by an independently driven pump, an
alarm device should be provided to warn against failure of the
pump.


(6)

(j)

A device should be provided to give warning of failure
of the inner seal (e.g. by detecting loss of sealant from a closed
system).

Double seals in 'tandem' may be used in the following circumstances:
(1)

Where the pressure to be sealed against is too high for one seal. In
this case the pressure in the intermediate space should be
RELIABLY controlled at approximately mid way between the sealed
pressure and atmospheric. This can often be most easily arranged
with a controlled leak-off of pumped fluid. Arrangements to alarm
and/or trip on failure to monitor the correct pressure should be
provided.

(2)

As a means of containing leakage of very hazardous fluid such as
liquefied flammable gases. In this case the intermediate space
should be filled with a suitable innocuous fluid such as water/glycol
mixture. A pressure alarm should be provided to indicate failure
of the inner seal and provision should be made to contain the
leakage and direct it to a flare system or other means of safe
disposal.

(k)

In any case in which tandem seals are used, provision must be made to
remove the frictional heat from BOTH seals.

(m)

Pumps handling flammable liquids above their auto-ignition temperature or
above 200oC should be provided with steam quenching external to the
seal.

(n)

Pumps operating continuously at temperatures below O0C should be
provided with facilities to quench the outside of the seal with methanol or
similar solvent to clear ice build-up.

(p)

It is good practice to provide all pumps handling flammable liquids with an
external throttle bush, so limiting the leakage in the event of a seal failure
and facilitate the provision of quench facilities if found necessary.


(q)

Mechanical seals should not be fitted to;
Fire pumps or any other emergency pumps, Pumps used only
occasionally, Boiler feed, or other multi-stage pumps fitted with balance
valves.

13

DEALING WITH SEAL FAILURES

13.1

Problems

The previous pages present an hypothesis of seal operation. What makes seals
fail? Either a lack of understanding of the conditions at the seal leading to a
wrong design or not adhering to the condition laid down in their design. Problem
areas are the following:
a)

Material Selection
If one of the materials in the seal is attacked by the fluid handled, the seal
will fail in a short period. All seal manufacturers nowadays keep a wide
selection of material with which it is possible to avoid material problems.
Neglecting those failures which are caused by chemical attack, actual
materials may lead to wear, due for example, to the non-wettability of the
rubbing pair (Stellite and water do not go together).

b)

Lack of Liquid at the Seal
This can be due to running a pump dry, forming a gas lock near the seal
so that liquid is not present at the interface, poor venting by supplying
vapor instead of liquid or adding too much heat in the seal. Many seals are
fitted with pumping rings which pump the liquid round an external cooler.
The thermal conductivity of the seal materials and the geometrical layout
are also important in removing heat from the faces.

c)

Hot Running
Hot running is most likely to occur when the fluid supply is not readily
available at the seal due to either a bad mechanical design or the cooling
circuit being inoperative.


d)

The Presence of Solids
Due to the narrow gap between the seal faces, which is essential for a
seal to operate well, the presence of solids is always detrimental.

e)

Speed
A fairly low speed is needed to establish hydrodynamic conditions, the
exact value depending on the fluid handled. The upper limit is determined
by heat removal.

f)

Seal Geometry
When the faces fail to be flat due to mechanical or thermal distortion or fail
to be parallel due to insufficient flexibility in the support, the seal will leak.

(g)

Old Age
Plastic elements of seals e.g. "0" rings in Viton have a shelf life of less
than 5 years. The installed life is even less.

(h)

Secondary Failures
Secondary failures occur due to fretting of the secondary elements against
the shaft sleeve leading eventually to leakage or due to excessive heat
causing plastic elements to deteriorate.

(j )

Surge Pressures
Surge pressures experienced mainly when a pump is switched off are
responsible for a large number of seal failures. The high pressures and
reverse pressures caused (water hammer) are seldom considered at the
design stage.


13.2

Records & Failure Analysis

Before any attempt can be made in improving seal life, knowledge must exist on
which part of a seal fails, how often they fail and why they fail. To do this plant
records are essential. It should be noted that good records are the main clue in
establishing the cause of many a failure, not only that of seals, but also that of
bearings, valves for reciprocating compressors and many other items. A good
record should contain not only the actual running hours but also a description of
the failed element. However, even in the absence of the description, a statistical
analysis can be carried out to establish the category of failure experienced.
One of the most useful qualities of analysis methods is the Weibull analysis. See
Bibliography Ref (2).
This analysis permits by a statistical treatment of achieved performance to obtain
a value ''β' which classifies the type of failure. This analysis also provides LIO
and mean life. LIO represents the life which is survived by 90% of the population.
It is well defined in the rolling contact bearing industry, where the expected life is
calculated from a bearing operating conditions (temperature) and the bearing
load. Attempts have been made to classify seals the same way. However seal
failures are usually infantile mortality or at random and in these cases LI0 loses
much of its significance. Most seal failures can be attributed to short and random
interruptions in the conditions which occur at the seal, examples for this are,
pumps running dry, seal flush failing or seal cooling devices not being
commissioned.
From a Weibull analysis the following conclusions can be drawn.


13.2.1

A high ''β' ( >2.7)

This means wear-out.
(a)

If the mean life is adequate - accept; if not, re-design the part or parts
suffering the wear.

(b)

If the failures on-line are embarrassing, change the item at more frequent
intervals.

13.2.2

A ''β' of about I means failures are due to random effects.
(a)

If the LIO life is adequate accept.

(b)

If not, try to quantify the random occurrences.

It should be noted that a replacement on a routine basis will not
improve the failure on-line rate.
13.2.3

13.3

If ''β' is less than 1 then infantile mortality is the reason for failure,
examine each cause of failure and examine seal environment.
Replacement on a routine basis will tend to increase the failure rate
rather than improve it!

Maintainability

The correct installation and adjustment of mechanical seals is essential, yet often
difficult or even impossible to achieve. One reason is that most pumps were
designed for packed glands and mechanical seals were fitted as an after-thought.
It is hoped that, in future, pump and seal designers will co-ordinate their efforts
better. Some pump designers have done this - but it costs more e.g. Chesterton
"System 1" pump.
A partial step in this direction has been taken by the introduction of the cartridge
seal where all the critical adjustments are carried out in the workshops. A
complete seal, as a unit, can be fitted to a pump or agitator. Spares policy is
often to hold component parts of seal and assemble on site. It is a procedure
which can lead to problems if seals are assembled in situ from spare parts.


Standardization of seals will also be of benefit. At present seals are made in
many sizes even though some of them are not readily available! Standardization
would also minimize spare part holdings.

13.4

Other Solutions

Sometimes a solution to difficult seal problems is to use a canned magnetic or
electric motor with wet or dry stator, depending on the fluid handled. Vertical
cantilever shaft pumps without seals are also proving popular on atmospheric
tanks

SELECTED BIBLIOGRAPHY

The published literature on mechanical seals is very extensive and rapidly
growing. This short bibliography lists a number of general reviews that cover the
whole field of radial face seals and, through their extensive bibliographies,
provide a key to the literature on more specific aspects. In addition reference is
made to two recent papers (No 6) that produce a new important evidence on the
mechanism of lubrication of seal faces, and one paper (No 7) that lists a number
of seal arrangements that have been used successfully on chemical process
plants.
(1)

BS Nau 'Hydrodynamic lubrication in face seals', British Hydrodynamic
Research Association, BHRA Report TN843, 1965. A useful review on the
controversial issue of the mechanism of lubrication in face seals.

(2)

BS Nau 'Mechanical seals guide', BHRA Report TN904, 1967 A general
treatment of the fundamental design principles.

(3)

LVH Bernd 'Survey of the Theory of Mechanical Face Seals'
Part I

Characteristics of Seals. Lubrication Engineering,
1968, 24 479-484

Part II

Friction and Wear.

Part III

Dynamic and interfacial fluid effects. Lubrication
Engineering 1968, 3!:.. 597-604

Lubrication Engineering
1968, 24 525-530


An excellent review giving an extensive list of literature references.
(4)

E Mayer 'Mechanical Seals' Iliffe Books 1969
This is an English translation of Mayer's Axiale Gleitringdichtungen' 1966,
the only book devoted to mechanical seals.

(5)

J G Pape 'Fundamental Aspects of Radial Face Seals' PhD Thesis
Technische Hogeschool, Delft 22.10.69 (in English). The last Chapter of
this thesis ('Conclusions and Design Direction') is reproduced in BHRA
Rep. No, TNI046, March 1940.

( 6)

( a)

J G Pape 'Fundamental Research on a Radial Face Seal'
ASLE Trans 1968, !l, 302-309.

(b)

BA Stanghan-Batch 'Face Lubrication in Mechanical Seals' I Mech
E Tribology Convention 1971, Paper C59/71. These papers are
important as they produce convincing evidence that hydrodynamic
lubrication of face seals is induced by waviness of one of the seal
faces.

(7)

C Jackson 'Practical Guide to Mechanical Seals' Hydrocarbon Proc 1968,
47, 100-109. An interesting review of practical sealing arrangement on
process plant pumps by a Monsanto engineer.


APPENDIX A

MECHANICAL SEAL BALANCE

Considering the forces acting on the floating element of the seal.
(a)

PI

acting to the right

π/4 (B 2 - A 2)

(b)

P2

acting to the right

π/4 (A 2 - C 2)

(c)

The spring force acting

(d)

The mean face pressure /4 (B 2 - C 2).

Therefore:
The mean face pressure

= PI (B 2 - A 2) + P2 (A 2 - C 2)
(B 2 - A 2)

+ SPRINGFORCE
π/4 (B 2 - A 2)

= P2 + (PI - P2) x (B 2 - A 2) + Contribution
(B 2 - C 2) from Spring Force
(B 2 - A 2)
(B 2 - C 2)

IS CALLED THE BALANCE FACTOR


[t defines the extent to which the seal face pressure is affected by the pressure
difference across the seal.
In a normal 'unbalanced' seal the balance factor = 1.1 approx.
In a normal 'balanced' seal the balance factor = 0.7 approx..

APPENDIX B

USE OF TANDEM SEALS ON LPG OR OTHER
HAZARDOUS FLUIUS


NOTES:
(B1) NORMAL OPERATION
Slight vapor leak of pumped fluid carried over in sealant circulation and
released to flare.
(B2)

FAILURE OF OUTER SEAL
Loss of sealant detected by low level alarm on pot.

(B3)

FAILURE OF INNER SEAL
Major leak will cause pressure rise in sealant pot detected by high
pressure alarm.


Mechanical Seals

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