Slurry Pump Gland Seal Three Body Wear

and the Influence of Particle Properties including

Hardness, Size, Fracture Toughness and Shape

by

Nigel Ian Ridgway

School of Chemical Engineering

The University of Adelaide

A thesis submitted for examination for the degree of

Doctor of Philosophy

in

Chemical Engineering

January 2010

1

1. Introduction

1.1. General introduction

This research originated from observations of slurry pump seal failures in the field and their impact on

mineral processing costs and the escape of potentially harmful fluids to the environment. In typical

processing plants, minerals or mineral ores are first crushed and then transported between different

processes as slurries (suspensions of solids in water). Each mineral in the slurry can be identified by its

unique crystal structure.

Generally centrifugal pumps are used and they are sealed to prevent the slurry escaping. Three types of

seals are used:

1. Gland with compressed packing;

2. Centrifugal expeller with lip seal or gland packing secondary seal and;

3. Face type mechanical seals.

Of these, gland seals are the most common.

Gland seals are considered to be very basic engineering equipment and they are little studied. They are

normally designed by trial and error based on experience. It is interesting to note the observation by

Buchter: “… one wonders how industry can be so efficient despite the lack of scientific

reasoning.”(Buchter 1979)

There is a need to examine gland sealing from a scientific engineering perspective. There are numerous

uncontrolled variables in the slurry seal system including the components, the external seal water

supply, the slurry properties, and human influence on the performance of the seal during maintenance

and operation. Although there is a reasonable amount of literature published on investigating the

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hydraulic performance of gland seals in water systems, gland seal failures and the relationship with

slurry particles and the complex wear mechanisms of two and three body grinding are seldom studied.

Two body wear by definition is a form of abrasive wear in which the hard particles cause wear of one

body are fixed on the surface of the opposing body, and three body wear is caused when particles are

free to move between two contacting surfaces (Figure 1.1).

Figure 1.1 Two and three body wear (Stachowiak and Batchelor 2001)

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The slurry particles entrained in the gap of a gland seal create a lubricated abrasive wear contact

mechanism and is analogous to the wear of plain bearing journals. (In a process known as lapping,

particles are deliberately introduced into a gap for the purposes of polishing.

The following sections will provide an overview of the current knowledge of gland seals in slurry pumps,

and will identify the scientific principles and methodology for the research.

1.2. Gland seals

Gland seals have been employed to seal steam engine piston connecting rods and water pump drive

shafts from the 1850s. Early packing was manufactured from organic materials such as hemp and in

some cases the interface was lubricated with tallow. At first wrought iron was normally used for the rod

or shaft materials but this was later replaced with carbon steels. In the 1950s, sacrificial shaft sleeves

were developed using harder materials such as martensitic stainless steel. Modern packing is

manufactured from a mix of synthetic fibres which are woven into a braid for improved mechanical

strength and typically lubricated either with a PTFE (polytetrafluoroethylene) boundary lubricant or

graphite.

Gland seals are particularly used in mineral processing slurry pumps where the process pressure is high

(up to 7 MPa), the temperature is high (up to 150o C), the solids concentration exceeds 5%, the fluid is

corrosive, or some combination of these.

Common discharge diameters of the slurry pumps range between 25 and 450 millimetres. The

corresponding shaft sizes and speeds are shown in Table 1.1:

4

Table 1.1 Typical seal size range

Pump discharge size (mm)

Shaft sleeve diameter (mm)

Maximum speed (rpm)

Shaft surface velocity (m/s)

25 45 4400 10 450 220 400 5

Under ideal operating and maintenance conditions the useful life of the seal may be in the order of 8000

hours necessitating several replacements of the packing. However, should a deviation in seal conditions

occur, the useful life may be reduced significantly, typically to the order of hundreds of hours.

Functional failure results in slurry leakage possibly creating secondary failures such as corrosion and

early bearing failure. If the slurry is hazardous, seal failure may create an unsafe workplace. This is a

particular risk with the increasing trend to hot acid leach mineral processing. In today’s risk-averse

society, such an event is unacceptable. A typical failure and spillage is shown in Figure 1.2

demonstrating slurry discharging from the seal at the back of the pump to atmosphere.

Figure 1.2 Typical gland seal failure

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A typical gland seal (Figure 1.3) consists of an annular housing (stuffing box) to contain the packing, a

gland follower to adjust the packing compression, a shaft sleeve as a sacrificial wear element, and a

lantern restrictor or a lantern ring to distribute water from an external supply. The water is fed to the

gland seal to lubricate the packing, to provide heat transfer, and to flush particles from the seal gap

between the packing and the sacrificial shaft sleeve.

Figure 1.3 Typical gland seal components

The gland follower applies an axial stress to the packing which results in a radial pressure which decays

exponentially from the gland follower end to the wet end (Figure 1.4). The radial pressure in the packing

must be equal to or greater than the pumped fluid pressure at the wet end to effect sealing.

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Figure 1.4 Seal pressure distribution

1.3. Impact of wear on reliability

The significance of the gland sealing problem is demonstrated by reliability data from a sample

population of six common pumps in a process thickener underflow duty at a large (9 million tonnes per

annum feed rate) mineral processing operation in South Australia. The pumps are made of standard

modern materials and have a reliable gland water system. The pumps are operated independently but

experience similar conditions including slurry percent solids by weight, pump speed, pressure and flow

rate. The packing material was standard Garlock 20/25 with a synthetic-fibre lattice, braided

construction including PTFE dispersion and other unknown proprietary lubricants. Shaft sleeves were

duplex stainless steel substrates coated with a fused tungsten carbide/nickel chrome 500 micron layer.

When considering the failure of pump components, including seals, the type of failures reported, the life,

and the costs of failure need to be taken into account so that the right maintenance strategy can be

determined.

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Functional failure is defined as “the inability of any asset to fulfil a function to a standard of performance

which is acceptable to the user” (Moubray 1999) Useful life is defined from the Weibull distribution by

the initiation of an increase in conditional probability λ(t) and locates the parameter for planned

maintenance intervention (Figure 1.5a).

Mean time between failure (MTBF) is when 50% of the population has already failed and is a measure

of the seal reliability.

For the purposes of this research useful life has been defined as the MTBF minus the standard

deviation.

Figure 1.5 Reliability frequency of failure

In mineral processing, the key objectives of reliability improvement are to increase the useful life and to

reduce the standard deviation of the failure distribution at ageing which corresponds to a shape

parameter of the Weibull distribution when 1.0 > β < 4.0.

Over a two year period, 17 shaft sleeve and packing replacements were recorded. Assuming that all

parts were at functional failure and that failures were age related, then the cumulative percent failures

were plotted on Weibull probability graph paper.

8

Results from the investigation confirm that the failures are age related because the beta ratio (shape

parameter) for the Weibull conditional probability function is greater than one (Abernethy 1996). The

Weibull plots were reported from Isograph reliability software Avsim+ 9.0, Weibull 2 parameter fit, 2003.

The beta ratio for the sample of shaft sleeves was 1.01 (Figure 1.6) and for the packing 1.0 (Figure 1.7).

Figure 1.6 Shaft sleeve replacements: Weibull plot

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Figure 1.7 Packing replacements: Weibull plot

The mean life from the probability plot did not equal the average life (mean time between failure) and

this is reflected in the coefficient of variance for the data (Table 1.2) ie, the standard deviations were

close to the means for the shaft sleeve and packing data.

Table 1.2 Reliability data ( (Ridgway, O'Neill et al. 2004))

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In the author’s experience, maintenance history for gland sealing is difficult to capture from field records

and is a source of unreliable data for a number of reasons including:

1. gland packing is issued from central stores inventory in rolls resulting in uncontrolled local

satellite stores, and packing cut to length may be used during a maintenance intervention but

materials are not issued against the maintenance notification and excess packing is returned to

the satellite store;

2. maintenance can be opportunity-based and parts may be changed out at potential failure rather

than at functional failure. Hence, the investigation was performed over a two year period to

reduce the variation in data and based on the opportunity to source shaft sleeve samples from a

large typical mineral processing plant;

3. business enterprise systems such as SAP (proprietary computer software) and work

procedures were not used to record the pump run time hours at failure or the failure modes.

As the beta (shape factor) ratios were close to unity according to reliability centred maintenance

convention the shaft sleeve failures are age related, and there is a need to review the seal wear and

improve seal reliability. The Weibull plots were discontinuous and this indicates that the failure rate is

not constant and that there was more than one failure mode for the sample. For example, if the gland

follower adjustment is excessive, the radial stress causes increased friction and the packing will burn

resulting in an infant failure. Similarly, the observation that the plots were discontinuous also supports

the case for a mix of potential and functional failures.

From actual maintenance history, the mean time to repair for packing replacement is approximately 2

hours and for a shaft sleeve replacement 10 hours. Hence the pump availability, based on seal reliability

alone, mean time to repair (MTTR) and available operating time, varies between 95 and 99.6% and is

limited by the packing MTBF.

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Other key performance indicators for sealing are the number or frequency of maintenance interventions,

the cost of the required interventions, and the proportion which is planned. The key point is that the

MTTR for shaft-sleeve replacement and other wet end maintenance such as impeller replacement are

similar. If the shaft sleeve useful life is less than the wet end useful life (which is usually determined by

the impeller and liner life), then the replacement costs for a shaft sleeve is higher, all other things being

equal. All maintenance interventions for the sample pumps were unplanned over a two (2) year period

and this is not surprising considering the variation in the Weibull probability data.

1.4. Economic significance of wear

The processing of minerals is a significant contributor to Australia’s export income and slurry pumping

costs, from the author’s experience, are a significant part of any wet mineral processing operation. The

total expenditure on pump parts in Australia is approximately $ 200 M per annum while the current total

global spend on pump parts would be in the order of $ 1.5 billion. Assuming 5% of this is for sealing

materials, based on the proportion of seal parts to total wet end parts in a pump, $ 10 M per annum in

Australia and $ 750 M globally is being spent. The life cycle cost of a seal also includes the energy cost

(because there is friction between the packing and shaft sleeve) and the labour cost for maintenance

and operations (adjustment of seals may be executed by operations staff). Standards published by the

European Sealing Association indicate the energy costs alone for some pumping systems are as high

as 25% of the total energy costs in some plants (BHR Group, Hoyes 2005). To this needs to be added

the cost to clean up any leakage resulting from seal failures plus downtime costs and the cost of

damage to other parts of the plant caused by the leakage of corrosive slurry.

To demonstrate the significance of gland seal wear and failure, the life cycle costs were investigated for

the mineral processing operation already mentioned. Within this plant is a counter current decantation

(CCD) slurry filtering circuit with underflow slurry pumps installed for each thickener.

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The percentage of seal materials allocated to maintenance as a function of total pump materials ranged

between 30 and 40 % for the CCD pumps while absolute costs for seal materials ranged between

25 k AUD and 75 k AUD (over a two year period) per month (Figure 1.8) (Ridgway, O'Neill et al. 2004).

Figure 1.8 Gland sealing costs 2001-2002

A review was undertaken for the 220 gland sealed pumps installed (total horizontal slurry pumps

physically installed including the CCD underflow pumps) at the plant over a twelve month period

installed (approximately 100 were in duty with the remainder on standby redundancy). Ten to twenty

shaft sleeves were replaced every month (Figure 1.9a) at a cost of 20 k AUD per month (Figure

1.9b).The spike in costs for month 3 is due to the issue of a large number of spare parts from the central

store which were then stored in satellite stores before actual usage.

020000400006000080000

100000120000140000160000180000200000

PU 1 PU 2 PU 3 PU 4 PU 5 PU 6

Cos

t AU

D

CCD

CCD Gland sealing costs 2001-2002

Pump material

Seal material

Pump labour

13

Figure 1.9a Shaft sleeve usage (quantity)

Figure 1.9b Shaft sleeve usage (cost)

The gland seal must not only perform the function required but its life cycle cost must be minimised and

be consistent with corporate objectives of lowest cash quartile operating cost, including maintenance.

(Mining companies benchmark their operating performance by where they fit in cash costs based on

four quadrants, with a key objective of being in the lowest quartile to be competitive.)

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All the life cycle cost elements were reviewed for the same twelve month period and normalised against

the milled tonne feed to the plant, which ignores variation in slurry properties, resulting in a cost of

approximately $ 0.17/tonne of milled ore (Table 1.3). The most significant element was the cost of the

gland water itself (ignoring any environmental opportunity costs which are not simple to quantify).

Water for this particular plant operation is sourced from the Great Artesian Basin and the marginal cost

of process gland water includes life cycle elements such as power and maintenance. The unit gland

water cost ranged between $ 0.50 per kL and $ 1.00 per kL.

Table 1.3 Life cycle cost for 12 month period

Life cycle element Cost per annum k AUD

%

Gland process water 627 40 Shaft sleeves/spacers 307 20 Stuffing boxes 137 9 Restrictors/lantern rings 11 1 Corrosion end covers 20 1 Labour 200 13 Packing 250 16 Total cost/year 1550 100 Tonnes 9 MTPA Cost/tonne $ 0.17

In conclusion, this example demonstrates that gland seals in slurry pumps are a significant operating

cost based on the shaft sleeve materials used and the opportunity cost of gland water. Worn seal parts

increase the risk of higher gland water usage hence any improvement in seal life will improve the

environment through reduced consumption of scarce water.

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1.5. Project objectives and methodology

1.5.1. Introduction

The following research objectives were defined based on the author’s experience during 20 years in the

mining industry problem solving seal failures in mineral processing operations and feeling the frustration

experienced by end users of the consequences of unplanned seal failure. There are published

guidelines with reasonable descriptions and empirical rules on seal operation and gland water systems.

However, a number of gaps in the knowledge of slurry pump gland sealing have been identified and

remain to be solved. Prior research has focused on improving the hydraulic performance of the seal,

typically in a slurry-free environment, and over the past century many attempts have been made to

improve the mechanical design of gland seals through different mechanical arrangements, and

materials for the packing and shaft sleeve

There is a lack of understanding of the tribology of gland seals operating in a slurry pump and of the

wear mechanics during the life of the seal which incorporates all of the tribological elements in the

system. The slurry gland seal is lubricated by gland water and particles entrained in the gap contribute

to a continuum of lubricated two and three body wear; ie, the wear mechanisms don’t always exist in

isolation.

A tribological approach considering wear, friction (and lubrication) is required to improve the reliability,

useful life and ownership costs of slurry pump gland seals. Identifying the key failure modes in a slurry

environment and developing wear models and wear equations will permit more robust designs and

procedures to be adopted by the mineral processing industry, and will result in improved slurry pump

seal selection criteria.

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1.5.2. Objectives

A review of seal failures at a large copper mineral processing operation in South Australia and of the

literature has identified significant gaps in knowledge. A key deficiency is the absence of work on the

wear of seals and the development of a model explaining seal wear mechanisms. Over several

decades, there has been a quest to improve the gland seal performance by uniform compression of the

gland packing with limited investigation of the effect of particles within the system; ie, most of the prior

art was concerning gland seals in water pumps.

The major objective of this project is to improve the understanding of three body wear in slurry seals and

to develop an empirical model for the wear of the shaft sleeve in a standard centrifugal slurry pump

gland seal incorporating a range of relevant particle properties. A controlled laboratory experimental

approach was chosen to provide the optimal method of achieving the objectives. It was not the intent of

the research to design a new seal, packing or mechanical arrangement.

1.5.3. Methodology

The mechanics of two and three body wear are complex and are the subject of a significant body of

research. Given the number of controlling variables and costs involved in establishing a wear test rig, an

empirical approach was selected to measure the specific wear rate of the shaft sleeve. A wear test rig

was designed to deliberately inject slurry particles into the packing/shaft sleeve gap to cause wear and

mass loss of the shaft sleeve. The research process included:

1. Reviewing patents for gland seals and published literature on gland seals, packing and wear

mechanics

2. Quantifying the significance of gland seal wear by economic and reliability examples

3. Collecting shaft sleeve and packing samples from several mine sites followed by measurement

of the shaft sleeve wear surface profile to compare these with the theoretical pressure

distribution in the seal

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4. Designing and developing an experimental wear test rig similar in size and conditions to actual

site conditions to avoid scale effects

5. Commissioning the rig and experimenting with a wide range of particles having varying

hardness values

6. Measuring the specific wear rate (response variable) for each experiment as a function of the

key controlling variables for the particles (these included relative hardness with the shaft sleeve

material, particle size, particle shape, fracture toughness and the gland follower load)

7. Developing a predictive wear model based on the experimental data and introduction of an

equation based on a regression analysis

8. Undertaking a dimensional analysis of the seal wear system

9. Building a more developed qualitative physical wear model of the slurry seal wear from the

observations, experiment results and literature review

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2. Seal wear review

2.1. Introduction

This review of published material will initially consider patents for gland seal development (applied for

and registered in Australia, the UK and the US) followed by consideration of the effects of gland seal

pressure distribution on wear, two and three body wear, and finally the effect of the properties of

particles entering the seal system on wear. The key particle properties are hardness, shape, size and

fracture toughness.

An extensive assessment of the performance of gland seals in a liquor environment revealed a key

body of work summarized in seven reports by the British Hydraulic Research Association during the

1950s. A significant body of work by Denny (Denny 1957; Williams and Xie 1996) focused on the

pressure distribution in a seal. These ideas were further developed by Ochonski (Ochonski and

Machowski 1987) who investigated alternative mechanical arrangements which attempted to apply a

uniform radial stress to the packing. Insignificant work was identified that considered gland sealing in

slurry systems and the subsequent effects of particles on wear (Ridgway, O'Neill et al. 2005).

2.1.1. Component definition

The key components in the slurry pump gland seal are shown below in a component diagram of a

typical rubber-lined slurry pump (Figure 2.1) which is commonly employed in mineral processing to

hydrotransport mineral particles:

1. Shaft sleeve which protects the shaft from wear damage (item 34);

2. Packing rings , normally between three and five rings depending on the pump size (item 32);

3. Lantern restrictor to distribute to water evenly around the packing (item 33);

4. Annular stuffing box which contains the packing (item 35);

5. Gland follower which applies an axial load to the packing (item 31);

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6. Gland water port for entry of gland water at a pressure higher than the pump impeller boss

pressure (item 38);

The “wet end” of the pump slurry seal is defined as the side closest to the pumped fluid in the casing

and the “atmosphere” side of the seal (also known as the drive end) is adjacent the gland follower from

which any leakage from the seal is to atmosphere.

2.2. Patents

A detailed search of the Australian, British and US patent systems was performed to reveal designs

relating to the key components of gland sealing (stuffing box, packing and shaft sleeve) or combinations

of these components that have been registered or applied for to improve gland seal hydraulic

performance or wear life. If the hydraulic performance is inadequate (excessive leakage to atmosphere),

the pumped fluid with particles gradually enters the seal gap resulting in wear of the shaft sleeve or

Figure 1 Slurry seal components (Weir Minerals Ltd) Figure 2.1 Slurry gland seal components (Weir Minerals Ltd)

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shaft. The patents did not provide a complete picture of the status of gland sealing knowledge.

However, some designs are in current use and interesting tribological principles have been highlighted

by their inventors.

The Australian patents from 1975 to 2002 available on-line or in the IP Australia office, Adelaide, were

reviewed. The most extensive patent listing was discovered in the US Patent database. US Patents are

more accessible than either the Australian or British systems and available in detail from 1776 through

the US Patent Trades office web site and these have been reviewed and described in their

chronological order.

If a patent was registered in more than one jurisdiction then only the US patents have been included in

this review.

This review does not include inventions which claim to improve gland seal maintainability, nor patents

for mechanical face seals or packing chemicals, as these are not relevant to the current research.

2.2.1. Explanatory note

Gland sealing design has been largely empirical for more than a century. Consequently little

fundamental study was applied until the 1950s by the British Hydraulic Research Association. The

following comments on and criticisms of the patents are based on the author’s experience of field trials

of equipment which are largely unreported and, in some cases, protected as trade secrets.

2.2.2. Australian Patents

The patents include designs for variations in the mechanical arrangement and packing modifications

which are claimed to improve seal performance and/or wear life.

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Muragan’s “Dynamic packing seal” patent application illustrates a stepped shaft sleeve so that the

packing is axially compressed against the step at the wet end side of the shaft sleeve (Muragan 1998).

The intent is to provide some packing compression at the wet end side of the gland and alleviate the

non-uniform packing pressure distribution. Muragan recognises that the lateral pressure ratio of the

packing material will affect the packing pressure distribution and that packings with high heat transfer

capability may have a relatively low lateral pressure ratio.

Figure 2.2 Stepped shaft sleeve (Muragan 1998)

This design has several problems including:

1. no mathematical model of the pressure distribution is developed;

2. the shaft sleeve is complex and expensive to manufacture;

3. the axial wear face does not provide for/allow for axial adjustment of the impeller/throatbush in a

slurry pump;

4. the packing maximum compressive stress is still at the gland follower end contributing to low

wear life.

Chesterton’s patent application (Jean and Jean 1980) claimed that a lantern ring design (Figure 2.3)

with preferential flow towards the wet end will reduce shaft wear and reduce the packing

compression. The idea is to provide maximum flushing (flow) towards the wet end, where the

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packing compression is least, and flush abrasive particles back into the pump wet end. The patent

fails to describe the distribution of the packing compression and the gland water radial or axial

pressure distributions.

Figure 2.3 Chesterton lantern ring (Jean and Jean 1980)

Most of the Australian patents and applications relate to packing designs.

Chesterton’s patent application (Kozlowski 1977) describes the packing density in a packing under

compression; this was also filed in the US as patent 4100835. When packings are compressed in the

stuffing box the fibres at the inner diameter are compressed and the outer fibres are in tension as

demonstrated by Figure 2.4 below.

Figure 2.4 Packing stress (Kozlowski 1977)

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The trapezoidal shape is claimed to promote leakage between the packings resulting in the need for

more packing compression force which reduces packing and shaft sleeve life. Chesterton also claimed

this could be achieved by having the core strands helically twisted in one direction and the braid strands

in the opposite direction. These braid strands were angled in such a direction that fluid would be

pumped along the packing/shaft sleeve interface to the wet end.

In the author’s experience, the helix pattern is rapidly lost with braid wear and/or packing compression

i.e. the packing material becomes more homogenous in a short time as the voids close and the lubricant

is burnt or washed away by the gland water. It is interesting to note that the application was patented in

the US but not Australia.

Unasco’s patent application (Bentley 1979) claims to improve the mechanical properties of

polytetrafluoroethylene (PTFE) which is soft and easily extruded under pressure by sintering with carbon

or graphite. The carbon should also improve heat transfer but no data is provided in the patent

application to substantiate the claims.

Chesterton’s patent application (Kozlowski 1979) introduced a new manufacturing method which

allowed manufacture of the braid helix in either direction. They also included a warp strand in an

opposite helical direction to the braid strand thereby increasing the packing density and they claim

reducing the need for packing compression and reducing wear (Figure 2.5). They were concerned with

lubrication efficiency and the costs of sealing. The ring would be made from an elastomeric material.

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Figure 2.5 Braided packing construction (Kozlowski 1978)

Chesterton (Kozlowski 1978) patented the basic construction of braided packing and the intent of this

patent was for a braided construction that would provide uniform cross section density when the packing

is compressed around a shaft sleeve and this was seen to be an improvement over the standard

herringbone weave by Garlock in US patent 3646846 (1972). However, detailed examination of this

patent indicates that the principal claim is the addition of graphite or other solid lubricants to the

impregnating suspensoid.

It was recognised by Chesterton's designer Kozlowski that, as the packing is compressed to control

leakage, the lubricant is forced from the packing and washed away by the gland water which increases

the friction and starts the wear process.

Robins, an employee of Finreco, lodged a patent application (Robins 1981) for a self lubricating

packing. They claimed that conventional packings with dispersed lubricants were a waste as the

lubricant was needed at the shaft sleeve/packing interface. His idea was to incorporate a hollow ring

filled with lubricant to be discharged from radial holes when under compression (Figure 2.6).

25

Figure 2.6 Self lubricated packing (Robins 1981)

Manegro’s patent application (Adorno 1995) is for a composite braid as distinct from composite fibre

(Figure 2.7). Many braided packings use PTFE yarns which are easily extruded from the stuffing box at

high pressure. This problem has been solved by incorporating aramide fibres of high strength and

coated with PTFE at the corners which improves the packing life but increases the shaft sleeve wear

rate. The patent claimed new art by positioning the aramide fibres in the centre of the packing braid and

the softer yarns adjacent the shaft. The application claims that the packing protected the shaft from

“uneven wear”.

Figure 2.7 Composite yarn (Adorno 1995)

Carrara’s patent application ((Carrara 1996)) claimed similar ideas to those of Chesterton’s application

77/511574 and US patent 4100835 noting the uneven compression. Carrara’s unique idea involves the

creation of trapezoidal-shaped packing in the shape of an “isosceles trapezium” with the larger base on

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the “extradotal” outer side. When the packing is compressed in service the packing shape will be more

uniform. A reduced force would be required for compression and hence there will be less wear on the

shaft.

Tompac’s application for a "gland sealing compound injector" (TomPac 1989) specifies a compound

mixture of lubricants and binders such as aramide fibres which is injected into the gland annulus with a

manually-operated positive displacement pump. The intent is to improve gland seal maintainability and

to describe a new gland adjustment process that could be automated. Similar products are available

today from Chesterton, Utex and Rainsflo. These products have not been successful in practice as the

mixture sets hard, forming a solid bush.

Garlock ( (Harrelson 2000) describes a knitting method where filler yarns are pre-stretched during the

construction thereby reducing keystoning and uneven pressure distribution. A further claim is that using

interwarp knitting methods localizes the distribution of yarns and hence low cost filler yarns can be

positioned where they have little effect on the packing performance while low friction yarns are

positioned adjacent the shaft surface where needed most to reduce wear and friction (Figure 2.8).

Figure 2.8 Interwarp knit construction (Harrelson 2000)

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2.2.3. British Patents

The National Research Development Corporation (Richards 1966) patented a gland seal using hydraulic

uniform radial loading either by means of external grease under pressure or using the sealed fluid itself.

The inventor claimed success with a wear life of “hundreds of hours” and conditions were reported to be

600 PSI (4 MPa) and 1000 fpm (5.1 m/s) shaft speed. The idea is very similar to the “Radpack”

marketed by Garlock in the 1960s.

Roth Company (Roth 1964) recognised that the packing compression is non-uniform and, when PTFE

type packings were used, there was considerable shaft damage at the gland follower end. Roth also

claimed that PTFE packing had a low coefficient of friction and his design was relevant for packings with

a low coefficient of friction. Roth’s patent described a stuffing box with radial grooves in the external

diameter allegedly to “preventing the pressure in the pump cavity from unduly pushing the packing back

out of the stuffing box and against the gland member”. As demonstrated by Ochonksi, the radial stress

distribution is a function of the packing mechanical properties and applied stress. The radial grooves will

have little if any impact on the radial pressure distribution as the friction coefficient between the packing

and stuffing box bore is not changed.

The Chemical Construction Company (Construction 1965) patented a packing combination using Teflon

O rings and metallic guide rings. Their claim was that there would be a uniform (linear) pressure drop

across each packing ring. The net effect of the design is a series of packings in the form of large O rings

and there is nothing novel in the design (Figure 2.9). There is no explanation of how the uniform

compression may improve life of the packing /shaft sleeve and oil is referred to as the hydrostatic

packing fibre boundary lubricant (Ochonski 1988).

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Du Pont (Rhodes and Ruddy 1973) described a new packing yarn containing a composite of PTFE and

polyamide filaments (ie, a mixed filament yarn) which was claimed to prevent shaft leakage in pumps.

Crossley (Taylor 1970) patented a similar idea noting that PTFE has a low coefficient of friction and

highly chemical resistant but had the disadvantage of poor thermal conductivity and excessive

expansion which results in swelling of the packing, excessive heat, and premature failure of packing if

the gland follower load is excessive. They revealed a composite yarn including graphite filaments to

improve heat transfer.

Flexibox (Flexibox 1974) patented a radially loaded hydraulic gland seal mechanical arrangement noting

that shaft sleeve wear is uneven with a conventional axially loaded gland seal. The packing is contained

within a rubber carrier and an external pressure is applied to the carrier (Figure 2.10).

Figure 2.9 Packing spline arrangement (Construction 1965)

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Figure 2.10 Hydraulic radial gland (Flexibox 1974)

Esso (Shaw 1969) also revealed that the radial packing pressure distribution is uneven and any

“excessive local pressure causes undue wear a the region of contact with the shaft”. The patent

revealed an internally spring loaded gland with a reduced number of packings to even out the radial

pressure distribution (Figure 2.11).

Figure 2.11 Combined tandem seal (Shaw 1969)

2.2.4. US Paten

Kerr (Kerr 1907) proposed

In the author’s experience

rods are exposed to the flu

The spring loaded gland

despite wide recognition in

seal life, this method of app

Fig

nts Mechanical Arrangements

a reverse packed gland as a means of evening o

the key problem with this gland type is that the in

id and would seize in a short time (Figure 2.12).

Figure 2.12 Reverse packed gland (Kerr 1907)

follower was patented by Everett (Everett 1947

n the industry that gland follower adjustment is a

plying the packing stress is not in use today (Figur

gure 2.13 Spring loaded gland follower (Everett 19

30

out the radial packing stress.

nternal gland follower and tie

7) and it is interesting that,

an important issue in gland

re 2.13).

947)

31

Johansson (Johansson 1980) patented a radial hydraulic seal in 1980 based on experience in the pulp

and paper industry (Figure 2.14). Water or a fluid is applied to the external surface of the packing rings

and item 48 has lips at either end to seal this fluid from the pumped medium. As the external surface

area is greater than the inner diameter surface area, the pumped medium may be used as the hydraulic

adjustment medium. This concept is very similar to the Tompac sealing injector.

Figure 2.14 Radial hydraulic seal (Johansson 1980)

Chesterton’s patent (Rockwood and Antowiak 1983) covered the vertical packing gland or radial gland

seal, which is still on the market today. One of the key features is that the shaft sleeve is eliminated but

replaced with a radial sacrificial wear plate (item 209, Figure 2.15).

Figure 2.15 Chesterton vertical packing chamber (Rockwood and Antowiak 1983)

Chesterton had claimed tha

packing is compressed ev

sensitive and either the g

relationship between the p

For a conventionally appl

beyond the residual limit w

particular packing (K incr

exacerbated).

There are very few vertic

rhetoric. Another problem i

axial gland seal has indepe

Ahlstrom (Vaisanen 1990)

and each seal is clamped a

follower bolt (Figure 2.16).

at this seal would solve all of Alcoa’s alumina refin

venly. In practice it was found that the adjustm

gland would leak excessively or the packing wo

packing lateral pressure ratio and adjustment met

ied load by gland follower bolting the packing w

which is defined by its load versus lateral press

reases as the load increases and if there is o

cal packing chambers in service today despite

s that the seal adjustment and impeller positions a

endent components.

patented a gland system incorporating two glan

at a different rate (deformation) based on differen

Figure 2.16 Series gland seal (Vaisanen 1990)

32

nery sealing problems as the

ment of the follower is very

uld burn indicating that the

thod was never considered.

was too easily compressed

sure ratio (K) curve for that

nly one row this would be

the patent and marketing

are both related whereas an

nd seals effectively in series

t thread pitches of the gland

Warman (Burgess 1998) p

Alcoa. No published repor

indicate the design was s

(Figure 2.17).

Fig

Nippon Pillar Packing (Has

self-centring capability (Fig

increases substantially and

life. The author Hashiguch

packed glands and this pa

by applying the load at both

is that the higher radial p

preferred compared to the

atented a uniform compression gland as part of its

rts exist on the trials as the design was not succe

sensitive to adjustment similar to the Chesterto

gure 2.17 Uniform compression gland (Burgess 19

shiguchi and Ueda 2002) patented a back to ba

gure 2.18). When gland seals are offset to the sh

d there is a larger gap for particle entry and hence

hi also recognised the problem of exponential rad

tent was an attempt to provide a means to even

h ends of the stuffing box. One of the key points w

ressure acts against the shaft sleeve nearest th

conventional pressure distribution.

33

s development program with

essful but anecdotal reports

on vertical packed chamber

998)

ck spring-loaded gland with

haft sleeve the leakage rate

e wear, shortening the useful

dial pressure in conventional

out the pressure distribution

with a reverse oriented gland

he pumped fluid and this is

Figure 2.18

The patent allowed for sha

have shafts with low slende

Alcan (Delaney 2002) high

and labour and noted that

leakage begins. They prop

box bore, and claim that th

2.2.5. US Paten

Considerable developmen

limited amount of commerc

of the technology is secret.

Nippon Pillar spring loaded gland (Hashiguchi and

afts that “vibrate or go eccentric” but this is not rele

erness ratios (L3/D4 where L is the shaft length and

hlighted in some detail the problems of frequent g

t as the packing/shaft sleeve radial gap increases

osed a combined gland seal and mechanical face

e face seal would prevent entry of particles into th

nt packing design

t has taken place in the last 200 years with gla

cially-available data on the mechanical properties

..

34

d Ueda 2002)

evant to slurry pumps which

d D the diameter).

gland adjustments with cost

s, slurry enters the gap and

e seal installed in the stuffing

he packing radial gap.

and packings but, given the

s and lubricants used, some

35

Packings have been employed at least since the 1750s for steam engine piston rod glands. In the late

1880s, as the pressure of engines increased with the introduction of superheat and the design of higher

pressure boilers, packings were developed to suit.

Winans, 1881 described square formed packing with a rubber core and the spaces filled with tallow

(grease). The rubber was designed to give the packing resilience and the braiding was relatively simple

in design, built up in sequential layers to provide mechanical strength (Figure2.19).

Figure 2.19 Winans braided packing (Winans 1881)

Interwoven braided packing was introduced in the 1930’s and a patent by Blaisdell (Blaisdell 1930)

describes the complicated weaving pattern to manufacture the packing (Figure2.20). Until the 1930s,

packings were very easily damaged as the yarns were layered and would disintegrate as the packing

wore. Braided packing features interlocking of the threads and series within the packing centre. As the

packing and shaft (or shaft sleeve) wears, the shape is retained.

36

Figure 2.20 Blaisdell woven braid (Blaisdell 1930)

Garlock introduced impregnated packing in 1964 to improve wear life. The lubricant cited is Teflon in the

form of a paste which is applied to each strand. The Teflon is a fine powder mixed with a thickening

agent of sufficient viscosity to keep the Teflon in suspension. Methylcellulose is mentioned as a

thickening agent and “Triton X-100” as a wetting agent. The distribution by weight is shown in Table 2.1:

Table 2.1 Components used in Teflon lubricating paste

cellulose 1.2% Triton 0.2%

Teflon 40 mesh size 49.0% water 49.6%

As the packing is formed the shape is calendered and the packing is heated in an oven to dry off any

residual moisture.

The patent was more concerned with cost saving as scrap Teflon is cited rather than new Teflon

material. A typical braiding machine (Webster, Houghton et al. 1964) which is used today by all

manufacturers of packing is shown in Figure 2.21. The yarn carriers are guided by tracks in the bed and

driven by gears hidden from view.

37

Figure 2.21 Braiding machine (Webster, Houghton et al. 1964)

Garlock’s US Patent (Houghton and Dixit 1972) introduced the lubricated herringbone weave packing

using graphite fibres which became the accepted industry standard and is still in common use today

(Figure 2.22). This packing differed from prior art by the addition of solid lubricants such as graphite

mixed with a dispersion liquid (also described as fluorocarbon resin such as the polymer PTFE) .The

ideal mixture described is equal weights of graphite, water and PTFE suspensoid. The fibres are coated

before and after braiding. A “crystalline lattice layer structure” is briefly mentioned.

Figure 2.22 Herringbone weave (Houghton and Dixit 1972)

Garlock's claim suggests re

cost of these graphite pac

PTFE boundary lubricants

Gore and Associates pat

filaments which may be us

microstructure of the amor

impregnation of packing fib

compromised. The Teflon

cooled (Figure 2.23).

Gore and Associates furt

disadvantage that the ther

by PTFE, which has low th

paste whereby the graphit

fibrils when the mixture is

fibres (Gore 1976a).

educed friction on pump start up reducing early lif

cking materials is excessive compared to stand

.

ent (Gore 1976a) revealed the manufacture of

sed in the construction of packing. The key poin

rphous Teflon consists of nodes interconnected b

bres with lubricants i.e. there is space between th

is expanded at a temperature exceeding 327o

Figure 2.23 Teflon structure (Gore 1976a)

ther developed packing recognising that PTFE

rmal energy developed during the start up/break i

hermal conductivity. The patent revealed a combin

te acts as a lubricant and conductor of heat and

sheared, hence making the lubricant less suscep

38

fe failures. However, the unit

ard aramid fibre types with

f porous Teflon continuous

nt of their design is that the

by fibrils which facilitates the

he nodes and strength is not

oC (melting point) and then

fibre or lubricant has the

n process is not transferred

nation of PTFE and graphite

d is entrapped by the PTFE

ptible to separating from the

Nippon Pillar (Ueda 1997)

of the packing width and st

As explained by the pate

construction method must

that graphite tape wound

exaggerates leakage betw

wear of slurry pump gland

They also claimed that the

other methods and therefo

Ueda (Ueda, Konaka et al.

(Figure 2.25). The main cla

the shaft and hence less pa

introduced graphite tape rather than fibres. The ta

trengthened by longitudinal fibres (Figure 2.24).

Figure 2.24 Folded graphite tape (Ueda 1997)

nt, expanded graphite has a low tensile strengt

consider how the packing is to be cut for mainten

d around a fibre or packing punched from s

ween layers. This material is in common use toda

seal shaft sleeves.

folded layer eliminates slip between the laminated

re improves shape retention.

. 1993) then patented similar laminated packing b

aim was that, due to the packing shape, there is no

acking relaxation as creep is reduced.

39

ape is folded in the direction

th and is fragile hence the

nance purposes. They claim

sheet lacks versatility and

ay to reduce the friction and

d packings manufactured by

ut with a zigzag folded layer

o slip in the axial direction of

40

Figure 2.25 Laminated packing (Ueda, Konaka et al. 1993)

Further work by Ueda revealed moulded graphite packing. Laminated graphite tape is spirally wound

and then moulded to shape, possibly including some metal reinforcing to maintain shape. The key

benefit claimed is that this type of packing requires a low gland follower force and causes low shaft

sleeve wear.

Tsukamoto (Tsukamoto 2001) designed laminated packing using expanded graphite tape and PTFE

combined by twisting them into a yarn (Figure 2.26).

Figure 2.26 Expanded graphite/PTFE composite (Tsukamoto 2001)

The idea was to combine the strength of PTFE with the lubricating and heat conducting properties of

graphite and compensate for its low tensile strength. The patent revealed that there is an optimum

graphite density of 0.8 to 2.2 g/cm3: above this the surface roughness is low and results in low bond

strength to the PTFE while below a higher surface roughness compromises the sealing ability. The

41

number of twists per meter is approximately 55 to 70 and is related to optimum bond strength. To

successfully bond with the graphite, the PTFE must be porous with a pore size of 0.05 to 15 μm.

2.3. The Effect of pressure distribution on wear

2.3.1. Introduction

The scientific work on gland seals is primarily concerned with liquor process pumps which do not have

particles in the pumped fluid. The following review is of major works from the BHRA (published as

internal reports for corporations with commercial interests), and scientific papers in the public domain.

2.3.2. British Hydraulic Research Association reports

The BHRA published an extensive set of internal reports on the performance of gland seals in a liquor.

They performed a range of experiments on different packing materials and stuffing box designs.

Denny published internally a seminal work on the force analysis of stuffing box seals and derived a

mathematical relationship between the radial and axial pressures (Denny 1957). Denny assumed that

the packing transmits pressure like a fluid and that the lateral pressure ratio equalled unity.

tl

R ePP /)( 21. μμ +−= Equation 2.1 Denny radial pressure

Where:

PR = radial pressure (kPa)

P = axial pressure (kPa)

µ1 = friction for packing on the shaft

µ2 = friction for packing on the stuffing box

l = stuffing box depth/packing (mm)

t = packing thickness (mm)

42

The packing radial pressure decays exponentially due to packing friction in the stuffing box housing.

This was proved by developing a linear test rig where packing was stacked in series and the applied

and reaction loads were measured. The exponential distribution was proven by loading successive

numbers of linear packings.

Turnbull, 1958 revealed that at low pressures the distribution is more even than for high pressures

where the bulk of the pressure drop, (up to 80%) occurs across the first row of packing (atmosphere

end) adjacent the gland follower.

Denny, 1959 evaluated the inverted stuffing box design and found that leakage rates are reduced with

this design for high speed duties. This is an expensive and complicated arrangement seldom used

today.

Austin and Nau, 1971 conducted a number of experiments with a purpose-built test rig using different

packing materials and monitored leakage and gland force. The seal was lubricated with water and no

particles were in the system. They reported wear of the shaft but it was not quantified. Wear was

observed to be greatest at the wet (fluid pressure) end and the grooves were up to 4 x 10-3 inches

(101.6 µm) deep.

Nau, 1968 conducted experiments with radially packed glands and reported uniform leakage rates over

10,000 hour tests. Shaft wear was observed and again there was no quantification of the wear or

explanation of the mechanism.

Nau and Austin, 1969 completed a life cycle cost review of a standard gland seal test rig incorporating

two glands mounted back-to-back (Figure 2.27) with a range of operating liquids including fresh water,

lubricating oil, diesel and saltwater.

43

Figure 2.27 BHRA gland seal test rig (Nau and Austin 1969)

Wear was highest at the gland follower end (position 20) with the exception of saltwater. Saltwater was

the most abrasive fluid at the wet end (fluid pressure) side of the shaft. Premature failure was caused by

crystallization in the gap and occurred within 100 hours. The groove depth was typically 11 x 10-4 inches

(27.9 µm) compared to 1.55 x 10-4 inches (3.9 µm) for fresh water. Profile measurements were based

on “spot measurements at 0.1 inches (2.54 mm) intervals” (Figure 2.28). They indicated that wear was

not uniform across the shaft in all experiments and did not follow an exponential function. No

experiments were performed with slurries.

Figure 2.28 Shaft sleeve wear (Nau and Austin 1969)

44

2.3.3. Scientific papers

Following the early works of the BHRA, analyses of the pressure distribution were published through the

International Fluid Sealing Conferences and in other technical papers. (Thomson 1961) published an

extensive pressure analysis of a gland seal which is consistent with the prior BHRA work. It was noted

that the wear is proportional to the contact stress and highest at the atmosphere end.

Coopey, 1965 revealed that, when seals were exposed to a slurry, the highest wear occurred at the

pump wet end (fluid pressure) due to entrainment of particles into the gap.

Allen and Rieder, 1968 did new work on the design and experimental evaluation of an inverted and

skewed gland seal where the packing rings were not normal to the axis of rotation but were rotated with

the stuffing box. They found that the coefficient of friction was reduced and the seal more stable to

operate. The design concept attempted to improve packing heat transfer by alternately covering and

uncovering the seal interface. Reduced coefficients of friction were reported but no wear results were

detailed.

Seal shaft sleeve wear became a more important issue from the 1970s. Austin highlighted the following

factors in seal wear:

1. Type of fluid and abrasives contained

2. Shaft sleeve material

3. Packing material and lubricants

4. Packing stress/pressure distribution

5. Shaft speed and temperature

Bohner, Blemke et al, 1975 published work on the lateral stress ratio, deformation and relaxation of

packings and made a significant contribution to the knowledge of packing properties including definition

of their physical, mechanical and chemical properties. Early work by Denny et al failed to consider the

45

viscoelastic behaviour of the packing and the lateral pressure ratio which is the ratio between applied

axial stress and radial stress (K = 0 when incompressible, = 1.0 when hydrostatic). A complex

relationship exists between the coefficient of friction, applied gland follower load, time and lateral

pressure ratio. For small K values and a given leakage rate the gland force required is small; however,

the pressure distribution is less even. As the K value increases, the gland load required is smaller but

the rate of decrease in radial pressure is higher than smaller K value packings. The work also identified

the property of packing stress relaxation as a contributing factor to wear (Figure 2.29). As the number of

gland adjustments (deformations) increases the residual stress increases and the packing approaches a

solid bush over infinity (K =0).

The stress relaxation reduces over time after each successive adjustment of the gland follower until

there is negligible relaxation. Before this point is reached the initial relaxation may be up to 75%, hence

the gland follower load and adjustment strategy are key variables in gland seal performance and wear.

(Wallace and Collins 1975) revealed that the radially packed gland was commercially manufactured and

tested by Flexibox with the aim of providing uniform compression and reducing wear of the shaft at the

atmosphere end. Their work showed that the fluid pressure distribution was also a function of the

hydrostatic gland water pressure, run time and gland follower load.

Figure 2.29 Stress relaxation (Bohner, Blenke et al. 1975)

46

Ochonski and Machowski, 1987 compared a number of new seal arrangements to provide a more

uniform radial pressure distribution . Several designs were tested including a reverse “inverted” gland,

loaded by individual disk springs; individually loading rings; and stepped packing sizes. Typically friction

was 20 to 25% lower for the “reverse inverted” gland compared to the standard gland seal (Figure 2.30).

Figure 2.30 New seal arrangements (Ochonski and Machowski 1987)

The radial pressure distribution (Figure 2.31) was based on the Denny equation with the insertion of the

lateral pressure ratio:

)()( xKxq σ= Equation 2.2 Radial stress and

)/2(.)( sKlD ex μσσ −= Equation 2.3 Axial stress

Where:

q(x) = radial packing stress (kPa)

K = lateral pressure ratio

µ = coefficient of friction between the packing and stuffing box

σD = gland follower applied stress (kPa)

σ(x) = packing axial stress (kPa)

s = packing thickness (mm)

l = length of packing (mm)

47

Figure 2.31 Radial pressure distribution in a gland seal (Ochonski and Machowski 1987)

Ochonski later conducted experiments to verify the actual radial pressure distribution compared to the

theoretical distribution. He developed a new relationship to describe the packing lateral pressure ratio

as a function of a number of material constants.

(Flitney 1986) verified that the radially-packed gland seal was not commercially viable due to the

complexity of manufacture, despite the benefits of a more uniform pressure distribution. The mechanical

properties of PTFE, graphite and aramide packing material were discussed. Flitney concluded that

aramide fibre displayed better creep and recovery than PTFE but was abrasive to shafts. Wear was not

quantified for the different packing materials.

48

Shevchenko, 1989 reported another novel method of uniform pressure distribution by deliberately

introducing deflection into the shape of the stuffing box bore using a plastic material so that the

deflection under the gland load would compensate for the packing radial pressure distribution (Figure

2.32). The stuffing box was termed a “compliant casing” and was manufactured and installed in a

commonly sized 150/100 pump. No data is reported in the work to validate the design.

Figure 2.32 Compliant casing (Shevchenko 1989)

Warman International considered a new uniform compression gland, the “cone gland” (Figure 2.33), as

a trial for slurry pumps installed in Alcoa’s refineries (Bushell 1992). This design competed with the

Chesterton vertical packing chamber which was claimed to be easier to adjust. However in practice the

vertical packing chamber was not favoured by the market as it was very sensitive to adjust and was

found to cause packing burn out when the gland follower was adjusted too tightly.

49

The seal type is 10 to 15 times more sensitive to adjust than the conventional packed gland (Ridgway,

O'Neill et al. 2004). The cone gland would also have been difficult to manufacture and, because the

vertical packing chamber was not successful, the cone gland was not manufactured for experiment.

Recently several suppliers have marketed bushes to replace the standard lantern restrictor which, they

claim, will centrifuge slurry particles away from the packing and back into the pump, thereby reducing

shaft sleeve wear (Figure 2.34 and Figure 2.35). However, there is little scientific data available to verify

their performance .

Figure 2.33 Cone gland (Bushell 1992)

50

Figure 2.34 Centrifuge type bushes (Technologies 2001)

Garlock (Figure 2.35) are claiming, in recent publications, a reduction in seal gland water of 50 to 80%

and that the bush “reduces significantly shaft/sleeve wear” without any verification of the claim

(Technologies 2008).

Figure 2.35 Garlock seal bush

2.3.4. Summary

The Patent literature catalogues a large number of gland seal developments over time but the actual

design principles are often hidden by legal language that is designed not to be interpreted irrespective

51

of the patent’s protection. Improved packing construction and materials have been developed in the last

century. Most authors acknowledge that the wear life of the seal (shaft sleeve and packing) needs to be

improved. In recent years, the material properties of the packing have been examined closely at the

micro level but little of this information is publicly available. Recent patents provide some direction for

tribological review of PTFE/graphite packing lubricant composites with shaft sleeves used in slurry

pumps.

The patented gland seal designs and published literature demonstrate that many inventors and

engineers have been chasing the holy grail of uniform compression without any useful result to the end

user. This supports the need for a tribological model of the actual wear process and a better

understanding of the influence particles have on seal wear during non ideal conditions.

Neither Ochonski nor Schevchenko considered modeling the vertical packing chamber or some

combination of radial and packed glands.

2.3.5. Seal wear experiments

Relatively little work has been undertaken on seal wear experiments. Hart et al recognised this

deficiency in their analysis of lip seal wear (Hart 1999), clearly demonstrated by their words “practically

no work has been done on the failure modes of rotating shaft seals”. They employed laser-induced

fluorescence techniques to analyse lip seal wear and found that dirt builds up to a critical thickness

which then breaks the oil surface tension and forms pits in the dirt layer allowing a mass flow of oil.

Ayala used an identical technique to quantify the oil film thickness and particle bed thickness under a

rubber lip seal. It was found that a dimpled lip seal surface with raised protrusions broke the dirt clusters

and allowed the lower regions to act as reservoirs of lubricant for the high contact pressure regions

(Ayala, Hart et al. 1998).

52

A shaft sleeve tester was developed by Warman International Ltd and consisted of a shaft sleeve

specimen rotating in a slurry bath with packing subjected to a constant load applied by a lever. It is

important to note that the packing load is constant and applied evenly over the shaft sleeve length which

does not represent the actual tribolological conditions (Mitchell, Huggett et al. 1990). This test was used

effectively to rank the wear resistance of different shaft sleeve materials (Figure 2.36).

Figure 2.36 Shaft sleeve tester (Mitchell, Huggett et al. 1990)

As expected, fine particles have a greater ability to enter the seal gap and increase wear rate. Sharp

beach sand produced a higher wear rate than rounder river sand particles. The wear rate was reported

as a “wear rate factor” equal to the sample wear rate (mass loss) divided by the wear rate for a

martensitic stainless steel shaft sleeve over a 20 hour test (Figure 2.37). Increasing the shaft sleeve

hardness was found to reduce the wear rate factor (Huggett and Varjvandi 1989).

53

Figure 2.37 Warman shaft sleeve tests (Huggett and Varjvandi 1989)

2.4. Two body abrasive wear

Two-body wear occurs between two surfaces. When they are in contact, material is removed from one

of them by the relative movement. Archard’s equation (originally devised for adhesive contact) is

normally used to describe two body wear. The wear is directly proportional to the load and sliding

distance, and inversely proportional to the surface hardness of the wearing material (Williams 1999).

HPsKW /×= and HKSWR /= Equation 2.4 Archard’s equation

Where:

K is a dimensionless constant or non dimensional wear coefficient.

H = hardness of wearing material (Vickers Hardness Number or GPa)

P = contact load (N)

s = travel distance (m)

W = wear volume removed (mm3)

The ratio K/H is the specific wear rate SWR or dimensional wear coefficient. High stress abrasion

occurs when the particles are fractured during the wear process.

54

During the life cycle of a seal, entrained particles are expected to be initially wedged in the packing,

forming a two body abrasive wear mechanism; then, as the gap increases relative to the particle size, a

three body wear mechanism develops. Particles can be represented by a three dimensional pyramid

with 2ψ the dihedral angle and θ the attack angle to the wearing surface (Figure 2.38).

Figure 2.38 Asperity or particle geometry (Williams 1999)

The particle shape influences the dihedral angle, and range of potential attack angles shown in Figure

2.39 below. The rates of wear are also different between the contact processes of cleaving, cutting, and

ploughing. Micromachining (cutting) rates are relatively higher when the volume of material lost

increases the machining efficiency; hence reducing friction actually causes a relatively higher machining

rate.

Figure 2.39 Abrasive wear modes (Kato, Hokkirigawa et al. 1986), (Williams and Xie 1996)

55

When the particle shape is sufficiently acute (sharp) the material deformation mode moves from elastic

to plastic. Given a sharp particle, and provided the attack angle is greater than some critical value,

cutting is possible (Williams and Xie 1996).

Williams expanded this work with a wear map showing the ratio Hs/Hb which is a measure of the softer

materials work hardening; the solid lines show the separation between cutting and ploughing processes.

It is important to note that spherical particles have an angle approaching 180 degrees and low attack

angle and vice versa (Figure 2.40) , hence spherical particles cause less wear.

Figure 2.40 Wear mode map (Williams and Xie 1996)

This wear map emphasizes the importance of particle shape and hardness in two and three body

lubricated abrasive wear (note the attack angle is usually represented by θ). Particles can be

considered as machine tools and the mechanics of orthogonal cutting apply (Figure 2.41). Smaller and

angular particles have higher attack angles and a higher shear rate of the deformed surface. When the

56

attack angle is low, ploughing is the dominant mechanism; as the attack angle is increased, a critical

value is reached and the dominant wear mode is cutting or micro-machining.

Figure 2.41 Orthogonal cutting mechanics (Kalpakjian 1985)

2.5. Three body abrasive wear

Three body abrasion occurs when hard particles are free to roll or slide between two surfaces. There is

relatively little work published in this area. There are three tribo elements in the wear system: the two

wear surfaces and the particle. Direct observation during wear is difficult and therefore few deterministic

models have been developed. The wear rate is generally an order of magnitude less than the two body

wear under comparable conditions; however, this may not be the case when hard particles become

embedded in a softer surface (Williams 1997) .

The wear rate for three body wear has been found by experimental testing and empirical models as well

as numerical simulation. There is a range of laboratory test methods including rubber wheel, steel

wheel, pin on disc, and ball cratering tribometers (Stachowiak and Stachowiak 2002). Their work with

three body testing reported that the wear mechanism is a function of the critical ratio of load to slurry

concentration. Two body cutting wear was related to high load/low concentration and rolling wear to low

load/high concentration.

57

Misra et al (Misra and Finnie 1979) reviewed the literature and defined closed three body wear when

loose particles are trapped between two sliding or rolling surfaces close to one another. The

classification can then be further subdivided into high or low stress. They designed and conducted a

range of three body wear tests with a “low stress open three body wear tester” and found that the wear

rate for three body wear is not linear at low loads but then approaches linearity (Figure 2.42).

Figure 2.42 Wear of AISI 1020 steel with SiC particles (Misra and Finnie 1979)

Fang et al (2004) revealed that, when all the particles slide together, three body abrasion can be

considered as two body abrasion. Their work modeled abrasive wear by a probability distribution

combining the cutting and ploughing contributions. (Fang, Liu et al. 2004) predicted the wear rate of

materials by a Monte Carlo simulation of spherical particles in a three body system and compared this

with a ring on plate test. Their work revealed the shear stress distribution below the surface is different

for rolling and sliding particles: sliding particles have stress closer to the wear surface for wear by plastic

deformation (ploughing).

2.6. Effect of particle hardness on wear

Minerals are quantified in hardness by a variety of methods. Moh developed a relative measure of

hardness based on the ability of a mineral to be scratched by another. This scale ranges from talc at

58

one and diamond at ten; however, this range is not linear when compared with the actual

microhardness of the minerals. The earth’s crust contains mainly silica (quartz) which has a hardness of

Moh = 6.0 to 7.0 (Nesse 2000). Today, hardness is quantifed by mechanical test methods such as the

Vickers hardness tester through measurement of the applied load and indentation.

The ratio of metal hardness (H) to the abrasive hardness (Ha) has been demonstrated by Kruschov,

Mutton and Hurricks et al to be important in determining the abrasion resistance of materials (Mutton

1980), (Williams 1997). A critical ratio of H/Ha is shown to exist at a value of 0.6. Above this ratio the

wear resistance of a material significantly increases (Figure 2.43). It is important to note that most of this

work relates to two body abrasive wear.

Figure 2.43 Effect of metal to abrasive hardness ratio (Mutton 1980)

Three body abrasive wear of mild steel and white iron were investigated by Stachowiak and it was found

that wear increased with the hardness of the abrasive particles (Stachowiak and Stachowiak 2002).

However, it was also found that quartz has a higher wear rate than silica; although they have the same

hardness, the quartz is more angular in shape.

Williams summarised the work on relative hardness, neatly noting that an asperity in one material can

only abrade another should the difference in hardness be at least 20% greater -- which is the basis of

Mohs scale (Williams 1997). This work reveals that there is a relationship between the particle size,

59

shape and hardness. For high attack angles (sharp, lower roundness factor) the mechanical strength of

a particle reduces. The ratio H= Ha/Hs (note this ratio is the inverse of Mutton’s work) must be

exceeded for abrasion to be possible. For example, if Ha/Hs is equal to or less than 1.4 then asperities

with attack angles less than 45 degrees would be fractured rather than cause wear on the surface.

Figure 2.44 Relative hardness and shape (Williams 1997)

Consequently the hardness of the triboelements in a slurry seal are all relevant to the wear model.

Archard’s equation for two body wear, originally proposed for adhesive wear, is a useful starting point

and includes the hardness of the wearing material. However, it does not consider the hardness of all the

triboelements in a three body system and published work on relative wear resistance is for two body

abrasive wear experiments.

The seminal work of Williams et al on lubricated contacts highlights the “overlooked hardness effect”

also known as the hardness paradox, whereby particles become entrapped in a softer surface in a three

body lubricated system (Williams and Hyncica 2001). The work of Axen et al shows that the wear of a

surface actually increases when the counterbody is softer as particles become embedded in the softer

surface and conditions effectively move from three body to two body wear (Figure 2.45). This means

that making a surface softer compared to another may actually increase the specific wear rate and is a

paradox of relative hardness.

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Figure 2.45 Axen dimple grinder test (Williams and Hyncica 2001)

Miller and Miller also found that the wear rate of materials is a function of the particle hardness. Their

experiment, now standardised by ATSM G75-89, is a three body test. Four wear materials are

reciprocated under load simultaneously between slurry particles and a neoprene lap (Miller and Miller

1993). The rate of mass loss is measured and converted to a standard Miller number and Slurry

Abrasion Response of Materials (SAR) number. This work revealed a direct relationship between Miller

number and particle hardness (Figure 2.46). The test was originally designed to evaluate the

degradation of particles rather than wear rate.

Figure 2.46 Miller number (Miller and Miller 1993)

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Published values of the mineral hardness in terms of the Vickers Hardness number have been included

in the chart for comparison and the correlation with Miller number is strong which accentuates the

importance of particle hardness in a seal wear system.

2.7. The effect of particle size on wear

There is little published work on the impact of particle size on seal wear. Slurry particles in mineral

process circuits range between a top size of 30 mm in a milling circuit down to several microns.

Typically in a flotation circuit the particle size is less than 100 microns. Roco, 1990 characterized a

material with a non-uniform size distribution by a representative particle size equal to the 85% passing

size (d85) and this was employed in an erosive wear model for pump impeller and liner wear.

Misra and Finnie, 1981 found that there is a limiting critical particle size in both two and three body

abrasion of ductile materials and the wear process is less efficient below a particle size of 100 microns

(Figure 2.47). They concluded that there is a shallow surface layer in the abraded material which has a

higher flow stress than the bulk material for fine particle wear.

Figure 2.47 Particle size effect for copper sample (Misra and Finnie 1981)

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Generally the erosion wear rate is proportional to the particle size by a power constant between 0.3 and

1.2 (Ghandi and Borse 2003) . The mean diameter was found not to be representative of the equivalent

size; it is thought that the finer particles provide a fine boundary layer to lessen the impact of larger

particles resulting in a reduction in collision efficiency.

2.7.1. The effect of seal gap

The published work in lubricated abrasive contacts has been limited until the work of Williams et al

(Williams and Hyncica 2001). This work found that there is a critical gap in journal bearings in relation to

the particle size entrained in the lubricating fluid.

The Archard equation is used for two body wear. In three body conditions, where there are slurry

particles in the seal gap (h) between the packing and shaft sleeve, the wear mechanics are more

complicated.

Williams highlighted that, when the particle size is small, particles tumble through the gap and cause

random erosive wear. As the size approaches a critical ratio the mechanism changes to micro

machining and ploughing. Particles enter the gap and become embedded in the softer surface (bearing

journal) and rotate until a force equilibrium is reached and machining of the harder and softer surface

results with the line of force acting offset to the particle centre (Figure 2.48). The resulting angle of

particle inclination is also a function of the surface relative hardness. If the surfaces A and B have

similar hardness the reaction forces are aligned with the particle centre. The orientation θ2 of the particle

on the right is related to the hardness of the two surfaces. If the relative hardness between the surfaces

is greater than unity, then θ2 > θ1 the particles are encouraged to embed in the softer surface and the

wear rate of the harder surface is actually increased. Embedment of particles in seal packings is

intuitively expected as they have a braided lattice structure and a low overall hardness. Note that this

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analysis by Williams does not include the physical properties of hardness and fracture toughness of the

particle.

Figure 2.48 Particle equilibrium (Williams and Hyncica 2001)

2.8. The effect of particle shape on wear

The effect of particle shape on wear is revealed in early works on the analysis of wear debris found in

condition monitoring of engine lubricants. This work relates to the shape description of the particles

resultant from the worn surface rather than the foreign particles in the system.

The science of shape description is complicated and today there is no one universal method of accurate

quantification. Particles have a three dimensional shape and are anisotropic as the shape and

properties of a particle are not consistent depending on the particle orientation and direction of analysis

ie orientation of the particle in relation to the view by digital imaging or optical measurement.

For many years the mining industry has described particles as round, semi-angular or angular, based on

simple observation. Particles may also be described in geometric terms as oblate (rounded) or prolate

(elongated) (Pellegrin and Stachowiak 2002). When particles fracture, they may break along cleavage

64

planes of the greatest atomic density or there may be no clear demarcation at all, resulting in a

conchoidal, irregular, hackly or splintery surface (Nesse 2000).

A particle surface is a complicated mechanical surface which does not have a direct analogy to an

orthogonal machine cutting tool. A particle may, in fact, have multiple vertices where the cleavage

planes meet, capable of wearing a surface by machining, ploughing or fatigue.

Kato et al examined the three dimensional shape effect of particles on two body abrasive wear and

found that the dihedral angle of a particle has an effect on wear rate and wear mode. Generally the

wear rate increases linearly with attack angle in both dry and lubricated conditions (Kato, Hokkirigawa et

al. 1986).

There is a plethora of quantitative descriptors available based on the interpretation of two dimensional

SEM images for particle sizes generally less than 100 micron in size. There are many potential errors in

using these images including the resolution and magnification of the particle (for example, many

surfaces exhibit a wide range of topographical features over a wide range of scales). This was revealed

by Majumdar et al for the analysis of topography, noting that different resolutions and scan lengths give

different parameter results commonly known as “parameter rush”. In addition the particle boundary

should be measured with the surface topography (Majumdar and Bhushan 1990). There are a range of

reasonable descriptors today for the particle boundary; however, the characterization of topography is

still a relatively new area of science.

Roylance Raadnui, 1993 and classified the morphology of particles by three qualitative features:

1. Outline shape (regular, irregular, circular or elongated)

2. Edge detail ie, the boundary (smooth, rough, straight, serrated or curved)

3. Surface texture (smooth, rough, striated, etc)

65

They applied the standard geometric parameters of aspect ratio and roundness factor, and analysed the

variance with standard deviation, skewness and kurtosis. Their work was intended to enable the

identification of particles and the wear processes from which they originated.

Stachowiak expanded the use of curvature analysis of two dimensional particle shapes by describing

the degree of angularity with the spike parameter quadratic (SPQ) fit which is based on Richardson’s

technique whereby a yardstick is walked around a particle perimeter. A circle of equivalent surface area

is superimposed on the shape with the same centroid; any vertices (spikes) protruding from the circle

are deemed of interest and the intersections with the particle perimeter are found. The corresponding

“apex” angle is found and averaged. This work has highlighted a practical descriptor; however, the

weakness is that the method requires a high degree of mathematical processing and does not consider

the particle size range of a slurry.

Stachowiak revealed that the wear rate in a two body experiment was linear with SPQ based on 20

particles per SPQ result (Stachowiak 1998); see Figure 2.49 (gb = glass beads, sic = silicon carbide, q

= quartz, ca = alumina, d = diamond, g = garnet, ss = silica sand).

Figure 2.49 Two body abrasive wear (Stachowiak 1998)

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There are a number of weaknesses with this method including:

1. Mechanical wear is caused by a sharp edge or vertex in the orthogonal cutting model;

however, a spike may not in fact be sharp or acute at the tip of the vertex.

2. There is no weighting for the range of particle sizes in a typical slurry distribution. Pellegrin

et al highlight the importance of shape and size and their relationship in other work; to

quote, “… size and shape are both logically and mathematically inseparable. This duality is

reflected in the fact that an elongated particle has certain projections that, from the

perspective of the workpiece , appear small and sharp. Other orientations present large and

therefore blunt projections , and equiaxed particles present the same projection irrespective

of orientation” (Pellegrin and Stachowiak 2005).

3. Protrusions less than 20% of the equivalent circle diameter are not filtered out; ie, there

would be a diminishing probability these protrusions contribute to the wear rate.

4. The method is scale dependent and there has been relatively little published work

highlighting the impact of scale on the standard deviation, and in comparison with other

descriptors.

This work also showed a similar linear result for three body wear (Figure 2.50).

67

Figure 2.50 Three body abrasive wear (Stachowiak 1998)

Stachowiak et al noted that, in three body wear, there are other effects such as particle embedding,

particle fracture and rolling behaviour which influence the linear response (Stachowiak and Stachowiak

2001). Quartz was found to have a relatively lower wear rate in this work compared to sand which has a

similar hardness, as the particles were elongated and tended to orientate so that the minor axis is

perpendicular to the direction of motion.

Mandelbrot revealed the difference between Euclidian geometry and fractal geometry in 1967 and the

use of fractal mathematics has been applied to two and three dimensional surfaces (Majumdar and

Bhushan 1990). Mandelbrot discovered that some rough surfaces in nature have a length that is scale

dependent and that length cannot be an integer (by definition there is no fractal number for a Euclidian

shape). He proved that, for decreasing units of measure, length increases. A log log plot of the length

versus unit of measurement gives the fractal dimension between 1 and 2 and a higher value equals a

more irregular surface (Figure 2.51).

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Figure 2.51 Fractal and topographical dimension

A fractal boundary may be thought of as a set of Euclidian shapes connected in series. The fractal

dimension for a surface ranges between 2 and 3.

Stachowiak reviewed the Richardson technique with a range of step techniques and calculated the

boundary fractal dimension with a maximum error of 2.7%. The important point is that there was no

comparison with the standard numerical statistical methods (roundness factor, etc).

It was found from the review of the information there was little scientific comparison of the statistical

methods based on SEM digitization and the newer methods of SPQ or fractal geometry to make a

comparison between the methods based on analysis of variance techniques. This was a key finding.

Kaye warned users of fractals “… any short cut should not lull anyone into thinking that, in the real

world, a fractal dimension used to describe a rugged system is independent of the scale of scrutiny” ie

not all particles will have a unique fractal dimension for different positions of measurement; for example,

different orientations of a particle for digitization (Kaye 1989).

Later works by Stachowiak reveal more complicated methods of quantifiying the surface topography by

development of multi scale methods such as wavelet transformation and fractal methods. The latter

methods include blanket, box counting, 2D Hurst analysis, and patchwork models. The key point is that

69

fractal methods work well with isotropic surfaces; however, mineral particles are generally anisotropic

and do not exhibit self similarity in all directions (Stachowiak and Podsiadlo 2001).

In summary, statistical methods describe the disorder of a particle shape, and fractals describe the

order behind the disorder present in the surface topography or perimeter. Other limitations of the fractal

method are:

1. Particles are assumed to have unlimited roughness but this may not actually be the case for

a machined surface, for example

2. Some fractal dimensions may be the same for different surfaces

3. The calculation is difficult if the nominal fractal dimension is unknown i.e there may be an

excessive number of iterations

The important point is that an irregular fractal boundary or surface does not imply necessarily that a

particle may exhibit sharp acute vertices likely to cause machining wear.

In orthogonal cutting and micromachining it is the attack angle and particle dihedral angles that cause

wear (material removal). Table 2.2 summarises the methods reviewed for quantifying a particle

boundary.

Hamblin et al reported some comparisons of the SP, SPQ and reciprocal of the shape factor. In all

cases (Figure 2.52) it was found there was a linear relationship between two body abrasive wear, SPQ

and reciprocal of the shape factor (Hamblin and Stachowiak 1996).

70

Table 2.2: Particle shape descriptors

Particle 2 D profile descriptor Comments Value Roundness factor Easy to measure Medium Shape factor Easy to measure Medium Ra, arithmetic mean roughness Same Ra may have different angularity is well recognised Low Rq, standard deviation =RMS Measure of variation in profile similar to fractal

ruggedness. Rugged surface may have range of dihedral angles

Low

Rsq, skewness Measures offset from the mean Low Rku, kurtosis Measure of peakedness is a very sensitive unreliable

parameter Low

Aspect ratio Easy to measure Low Convexity Easy to measure Low Spike parameter quadratic(Stachowiak 1998)

Vertex angle is averaged. Size fractions, scale effects are ignored

Medium

Groove function (Pellegrin and Stachowiak 2005)

Analysis is tedious as modelling is required to calculate/simulate the areas at different depths

Medium

Fractal (yardstick boundary) Richardson technique (Podsiadlo and Stachowiak 1998)

Fractal dimensions are sensitive to scale and particles are anisotropic. Some programming ,effort to measure medium

Low

Fractal (power spectrum) (Majumdar and Bhushan 1990)

Fractal dimensions are sensitive to scale and wear particles are actually anisotropic. Is a measure of ruggedness.

Low

Fractal (structure function) (Yuan, Li et al. 2003)

Fractal dimensions are sensitive to scale and particles are anisotropic. Methodology is complex, requires programming

Low

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Figure 2.52 Normalised wear versus shape parameter (Hamblin and Stachowiak 1996)

Mikli also found that the SPQ can be employed as a reasonable measure of shape and compared a

number of the statistical parameters (Figure 2.53) including:

1. RN = roundness = 1/shape factor or 1/form factor

2. RNF = roundness factor

3. IP = irregularity parameter

4. SPQ = spike parameter quadratic (Mikli, Kaerdi et al. 2001).

Figure 2.53 Shape factor ranges (Mikli, Kaerdi et al. 2001)

This work indicates that the roundness factor has sufficient sensitivity between a circle and triangle for a

reasonable quantification of shape boundary.

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Given the complexity of the fractal methods and the absence of published scientific comparisons of

fractal and statistical methods analysis of variance with other descriptors, roundness factor was chosen

as one of the controlling variables in this research.

2.9. The effect of particle fracture toughness on wear

Slurry gland seals are exposed to mineral particles during the wear process. When particles enter the

seal gap, they are exposed to significant compressive stress until the shaft sleeve is worn and the seal

gap and pressure differential allow particles to flow through the gap. Mineral particles, when deformed

under stress, generally exhibit a brittle tenacity and have a relatively higher tenacity than other

materials.

The published work on particle fracture is mainly concerned with comminution in mineral processing and

the size reduction of particles. Size reduction is required to increase the liberation of mineral bearing

ores and for the separation of particles based on size.

Particles are fractured under compressive stress and cracks are propagated because there are tensile

stresses acting normal to the crack plane.

There are two measures of fracture toughness: GIC which is the critical energy release rate per unit area

of crack plane and the stress intensity factor KIC for crack propagation. By definition cracks are predicted

to grow when the stress intensity factor is equal to or greater than the fracture toughness of a particle.

5.0.aYK CIC σ= Equation 2.5 Stress intensity factor

Where:

KIC = fracture toughness stress intensity factor (MPa.m0.5)

Y = crack shape geometry factor

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σc = critical crack tensile stress (Pa)

a = crack flaw size (µm)

and is related to GIC ;

( ) 2/1ICIC EGK ≈ Equation 2.6 Stress intensity factor and critical crack energy release rate

Where:

E = isotropic tensile elastic modulus (Pa)

GIC = critical crack energy release rate (J/m2)

Tromans et al (Tromans and Meech 2002) examined a range of minerals including sulphides, halides,

silicates and oxides, and treated the minerals as ionic solids. They estimated the fracture toughness for

each mineral using the Born model of ionic bonding. The toughness values were calculated based on

published elastic constants for a single crystal and represent the lowest possible values for ideal brittle

fracture in pure single phase polycrystalline minerals.

The work found that there are four fracture types: intragranular, grain boundary, interfacial and

interphase.

1. Intragranular fracture is an ideal brittle fracture without any plastic deformation.

2. Grain boundary fracture is cracking along the grain where atoms are arranged irregularly.

3. Interfacial fracture is the crack propagation along the interface between poorly bonded

subparticles within a mineral particle and relates to the geological history of the mineral.

4. Interphase fracture is the cracking between two different crystalline phases for example in rock.

Tromans revealed that the relative toughness values can be related by:

IFICIPICGbICIC GGGG )()()( >>>> Equation 2.7 Relative toughness

Where:

GIC = critical crack energy release rate (J/m2)

74

(GIC)Gb = GIC for grain boundary fracture (J/m2)

(GIC)IP = GIC for interphase fracture (J/m2)

(GIC)IF = GIC for interfacial fracture (J/m2)

The intragranular fracture can be taken as the upper limit. The toughness values are not exact because

mineral particles contain inherent internal and surface flaws of random orientation, position and size.

Particles are anisotropic and contain flaws, voids and cracks which relate to the prior geological history

of formation. Tromans et al also evaluated the fracture toughness of covalently bonded minerals based

on a Morse-type bonding model (Tromans and Meech 2004).

It has only been in recent times that the toughness values of individual particles have been

experimentally determined by microindentation of samples and measurement of the crack length (Broz,

Cook et al. 2006). Broz et al revealed the stress intensity factor values for the minerals in the Mohs

hardness range. This work showed that toughness generally increases with Mohs hardness (Figure

2.54). The low hardness minerals such as gypsum showed inconsistent cracking and talc no cracking at

all. The orientation of the crystal plane was noted for each experiment as the results are sensitive to the

direction of cleavage. Crystals break along planes of weakness which are cleavage planes; if the

mineral contains no weakness, the fracture will not show perfect cleavage but range between hackly,

fibrous and conchoidal fracture.

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Figure 2.54 Toughness versus microhardness (Broz, Cook et al. 2006)

Whitney later expanded the experimental study to include metamorphic minerals such as silicates

(Figure 2.55) because not all materials in the Mohs range are common rock-forming minerals and the

silicates are relevant to geological research. This work found minerals with a wide range of hardness

between 7 and 17 GPa have a similar fracture toughness in the order of 1.5 MPa.m0.5.

Figure 2.55 Toughness versus microhardness (Whitney, Broz et al. 2007)

(Cz = cubic zirconia, And = andalusite, Grs = garnet, Ky = kyanite, Or = orthoclase, Per = periclase, Sil = silmanite ,

Alm-Prp = almandine pyrope garnet)

76

If they are very tough, abrasive particles may not fracture in a wear process; any sharp edges or

vertices become rounded over time and reduce micromachining. Brittle particles will fracture and

present new surfaces for wear and, at the same time, reduce in size, resulting in a complicated wear

relationship between particle toughness and size (Stachowiak and Batchelor 2001).

2.10. Summary and discussion

In general it was found the patent literature was duplicated, many patents were multiple listed, and

applications for patents had expired. The detail contained in the patents of the inventions was vague

and explanations of key design features were, in many cases, brief or absent, probably to provide

commercial protection. (Denny 1959) stated that “… the traditional design is deeply rooted in drawing

office procedure” which supports the research objective for a scientific approach.

There is an absence of work that reviews in any structured way all of the triboelements in a slurry pump

gland seal system. This current work is therefore considered to be novel through its systematic

consideration and unification of the relevant properties of the particles that enter the slurry seal gap

during failure. Most of the published design criteria and experimental work on seals has been in a liquor

environment and does not consider the entry of particles into the seal gap between the shaft sleeve and

packing. The review of the literature on particles confirmed that the properties relevant to the wear

model were size, hardness, fracture toughness and shape and that these properties have some

interdependency. No definitive relationship was found between the sacrificial wear component (shaft

sleeve) and slurry particle properties.

The historical approach has been to design a seal that has the optimum hydraulic performance and

uniform pressure distribution through novel mechanical arrangements, new packing materials and

packing braid structures. This approach has had little consideration for the contact mechanics and

particle motion when particles do enter the gap during seal failure.

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Most authors agree that the science of quantifying particle shape and surfaces is still in development.

Defining the fracture toughness of crystals is more recent work and combining these properties in one

model is a novel approach to the wear model.

There is little work on three body wear (compared to two body wear) and the properties and motion of

particles. The gland seal wear problem in tribolological terms may be considered as lubricated three

body wear analogous to plain journal bearing wear because both systems have a rotating shaft inside a

housing and are lubricated by an externally supplied fluid. In the case of the bearing, the external

housing supports the load; in the case of the seal, the external housing locates the gland packing.

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3. Hypothesis

3.1. Reliability

In practice, radially-packed uniform compression gland seals have not proven successful as the gland

follower is very sensitive to adjustments. To improve the reliability, useful life and ownership costs of

slurry gland seals, a tribological approach is required. Identifying the failure process (wear model and

wear equations) in a slurry environment will permit more robust designs and procedures to be

developed for the mineral processing industry and will result in improved slurry seal selection criteria.

The intent of any new potential design is to increase the reliability (useful life) and reduce the standard

deviation for a population of failure events (Moubray 1999) and this is relevant for the shaft sleeve and

packing components in a gland seal.

Figure 3.1 Reliability and life of seal components

The sealing problem is very complex and to simplify it a cause and effect (Ishikawa) model was

developed with focus on the effect of shaft sleeve wear.

The causes can be classified as design, materials and operations. Design causes arise from the

mechanical and chemical properties of the triboelements in a seal, and may include the gland water

system and the influence of impeller back vanes on the seal performance. Operations influence the

useful life of the components: for example, some users operate pumps until failure while others may

replace seal components based on condition (ie, the maintenance strategy influences useful life). Other

operating variables include the gland follower adjustment (frequency and load), fitting tolerances (which

79

can have an effect on shaft sleeve/stuffing box concentricity), and variations in gland water quality which

influence shaft sleeve useful life; for example, depending on the mine site, gland water may include

dissolved salts and organic materials. Process causes relate to the slurry properties such as the particle

size, shape distribution, percent solids, acidity, and pump duty.

The hypothesis relates to the slurry properties and their impact on shaft sleeve wear which are shown in

the Ishikawa diagram.

Figure 3.2 Shaft sleeve Ishikawa diagram

3.2. Chronology of seal failure

Gland water acts as a hydrostatic lubricant and is normally supplied at a pressure (35 to 200 kPa)

greater than the pump discharge pressure to lubricate the packing/shaft sleeve interface and at a flow

rate sufficient to flush any solids from the seal into the pump and reject thermal energy developed by

friction between the packing fibres and shaft sleeve.

Shaft sleevefailure

Material(mechanical) Wear process

Design Operations

Material(chemical)

Hardness

Porosity

Type

Topography

Wear distributionStribeck lubrication

Lubrication type

Gland water flow

Lubrication regime

Wear contact/particle properties

Wear debris

Gland water pressureFriction

Oxide products

Redox potential

Substrate leaching

Force balance Pressure distribution

Energy balance

TemperatureGland function

Useful life

Variance

Eccentricity

Adjustmentfrequency

Maintenancestrategy

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Common conditions leading to the functional failure of a seal are:

1. Packing hardens with age as the boundary lubricant is squeezed or melted from the packing

and the voids between packing fibres close under compression so the lateral pressure

coefficient approaches zero (ie, it becomes a solid);

2. Gland water pressure is lower than the pump impeller boss pressure (boss pressure is less than

the discharge pressure);

3. Gland water flow is too low (gland water may have been isolated during pump start for

example).

In normal operation, the gland follower is designed to be adjusted as the packing relaxes to compensate

for the viscoelastic behaviour of the packing (Bohner, Blenke et al. 1975). Packing fibres touching the

shaft sleeve cause localised heating resulting in further lubricant loss of the packing, when the lubricant

melting point is reached, and further relaxation. This process is shown in Figure 3.3.

Figure 3.3 Seal failure model (Ridgway, O'Neill et al. 2006)

Eventually, the packing becomes so hard that particles are able to enter the shaft sleeve packing gap

thereby initiating two and three body abrasion. When slurry irreversibly departs (leaks) to the

surroundings and no gland follower adjustment is available, the seal is at functional failure (ie, has

reached the end of its useful life). This generally occurs when the packing is hard and/or the shaft

sleeve has worn to a depth in the order of 1 to 5 mm either in specific/discrete places or over the entire

contact area.

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Figure 3.4 is a new model which illustrates schematically the gland water duty and pressure

relationships for a gland seal water supply system and slurry pump gland.

Figure 3.4 Seal gland water system curves (Ridgway, O'Neill et al. 2006)

New packings have an interference fit with the shaft sleeve resulting in a steep system resistance curve,

as shown by the new gland curve, and gland water flow is less than the limit of the flow control valve

(Maric brand). The flow is limited by an O ring constriction depending on the valve pressure differential

and the performance curve of the valve. As the packing relaxes the gland water flow increases and

reaches the flow control valve limit as shown by the tight gland system curve. As well, incorrect

selection of the flow control valve may contribute to failure if the flow rate is not great enough to

compensate for seal wear.

A typical performance curve for a Maric-brand valve is shown in Figure 3.5 which indicates the upper

range and flow rate of gland water for a typical large slurry pump. (The valve is rated at 34 L/min and

was selected for use with a 20/18 (450 metric) size slurry pump.) These valves are useful in the

operating range of 200 to 1000 kPa pressure differential across the valve. At high pressure differentials

the O ring constricts in size to reduce the flow.

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Figure 3.5 Typical performance of a flow control valve

If the gland seal fails, the maximum gland water flow rate is limited by the valve capacity and this

prevents excess flow and consequential failure of other gland seals supplied by the same (common)

gland water supply system. In either case, fine particles (typically microns in size) enter the

packing/shaft sleeve gap resulting in a complex tribological environment (Figure 3.6) where the number

of tribo elements in the seal system increases from four at zero time to eight at failure (excluding steam

which is an extraordinary consequence of failure) (Ridgway, O'Neill et al. 2005).

Figure 3.6 Seal tribo element model

Once the limit to gland follower adjustment is reached and no extra gland water flow is available,

increasing wear reduces the flushing effect thereby accelerating the failure rate. The packing radial

pressure is exponential and decays in relation to the friction between the packing and stuffing box,

packing lateral pressure ratio and dimensions in a complicated relationship; hence it is expected that the

0

20

40

60

80

100

120

0 200 400 600 800 1000 1200

% R

ated

flow

Pressure differential kPa

Typical flow control valve performance

83

slurry entry and highest wear would be at the wet end side of the seal, ignoring effects of packing wear

on the shaft sleeve at the wet end; ie, the wear profile would be expected to be an inverse function of

the radial pressure distribution.

An equation for shaft sleeve wear in a given slurry seal system with defined tribo elements is proposed

below (Equation 1). For simplicity, the shaft sleeve is considered to be constructed from homogenous

material. The volume removed from the shaft sleeve depends on the relative material hardness (Ha/Hs)

between the particles and shaft sleeve, particle fracture toughness (stress intensity factor, KIC), particle

shape (φ) and equivalent particle size (d85). It is also related to the length of contact (L), diameter of the

shaft sleeve (D), velocity of the shaft sleeve surface (V), fluid pressure (P), time of wear (t) and gap

between the packing and shaft sleeve (h).

),,,/,,,,,,(),( 85dKHsHahtVPDLtxSWR IC ϕΦ≈ Equation 3.1 Seal specific wear rate

The objective of the work was to find the empirical correlations between the wear rate and key variables

by laboratory experiment.

As finer particles are expected to initiate the wear process the specific wear rate should be inversely

proportional to the equivalent particle size. It is recognized that particle properties are not all intrinsic

and the variables selected in the above model will possess some level of dependency. For example,

smaller particles possess lower surface area and are less likely to have surface defects such as twins or

impurities which may be crack initiation sites. Hence, there is a tradeoff between the particle size effect

for less efficient two and three body abrasion and higher fracture toughness. For instance, if small

particles have conchoidal shape or smooth cleavage resulting from attrition of particles with low

toughness and small dihedral angles, the specific wear rate is increased. Mineral crystal formation

histories are unique and in practice the particle shape is some continuum between fracture and

cleavage of crystallographic planes (Nesse 2000). For convenience the d85 size of the relevant size

range was employed as a simple measure to compare samples as developed by (Roco 1990).

84

Particle fracture toughness is also related to hardness and crystal bonding. Crystals with closer packed

atoms along cleavage planes, smaller atoms or ions and higher bond density are less likely to fracture

and generally exhibit a higher hardness. Clearly, the properties are related. The probability of a particle

fracturing in the seal is related to the fracture toughness of the particle as particles under compressive

stress will possess some residual tensile component. As particles fracture the particle size distribution is

modified and new particle vertices become available for orthogonal cutting in three body wear (Figure ).

These may increase cutting efficiency. Quantification is complicated by the different and multiple mineral

phases and defects present in each particle. Recent work by Broz et al found that the property of

fracture toughness has a correlation with hardness indicating that fracture toughness is relevant (Broz,

Cook et al. 2006).

Figure 3.7 Particle fracture and 3 body wear

Very tough particles are expected to be round whereas brittle particles fracture which reduces their size

and maintains their sharp dihedral angles. Particles of medium toughness self-sharpen and maintain a

reasonable size. These considerations are relevant to the model (Stachowiak and Podsiadlo 2001).

Based on normal tribological considerations for two and three body wear, the specific wear rate should

be inversely proportional to the velocity and travel distance of the shaft sleeve.

Force

Packing

Slurryparticles

Fatigue

Shaft sleeve

Attrition

85

The dominant mechanism of wear is expected to be micromachining and ploughing for a homogenous

shaft sleeve material. From orthogonal cutting mechanics, there is a higher probability of wear when the

dihedral angle is minimised; ie, for larger attack angles (Figure 3.8) (Kalpakjian 1985). Kato and

Williams have shown that the Archard equation can be adapted based on the asperity attack angle and

higher rates of wear are related to micromachining and higher attack angles (Williams 1999). The point

is that high attack angles are most likely when the particle dihedral angle is small (between 60° and

120°). For lower attack angles, it is more likely that particles are rounded with dihedral angles

approaching 180°.

Figure 3.8 Orthogonal cutting mechanics

The potential interactions between the shaft sleeve and particle for a single asperity encounter are

summarised in Table 3.1.

The wear of the shaft sleeve is caused by multiple asperity encounters and the wear track in the surface

is an input into the wear of the shaft sleeve; ie, as successive particles contact the shaft sleeve surface

their motion and contact is influenced by the wear groove from the ploughing or machining of the

previous particle contact.

86

Table 3.1 Particle shape influence on cutting mechanics

Wear model Dihedral angle Rake angle Attack angle Shear angle High specific wear Low

60 to 120o cutting 0 to 60o cleaving

+ ve High

High shear strain, ploughing

Low - ve Low Low

High shear strain, ploughing

High - ve Low Low

Historically, efforts to improve gland seal reliability have focused on several main areas:

1. Adopting harder shaft sleeve materials and ceramic coatings

2. Selecting the right seal type (for example, avoiding the use of mechanical seals in large particle

applications)

3. Educating end-users to design and maintain effective gland water systems

4. Trialing lip seal and packing combinations

5. Trialing new packing materials and lantern restrictor designs

Most new developments are tested by field trials and these are problematic because trials are often not

supervised; they are interfered with by operations staff and the results are lost so there is no data as a

basis for scientific improvement.

Thus a novel experimental laboratory test rig was developed to test the wear model and develop a wear

equation.

87

3.3. Research Hypothesis

The project hypothesis is that uniform packing compression is not the main contributor to gland seal

useful life in a slurry duty and that a quantitative model relating the shaft sleeve wear rate with the

particle properties of hardness, fracture toughness, shape and size can be found from laboratory

experiment with these controlling variables being unified in one empirical equation. It is expected that

the conditions of a slurry pump in mineral processing service can be replicated in a laboratory test rig

with similar geometric scale and that wear of the shaft sleeve can be deliberately induced by injecting

slurry particles into the seal gap.

88

4. Method

4.1. Methodology and process

The overall objective of the work was to develop a new quantitative wear model for slurry seals and find

a relationship between the relevant particle properties in the seal system. This had four main

components:

1. Measuring the wear profile of shaft sleeves from industrial (mining) service

2. Designing an experimental test rig and standardising the experimental process

3. Measuring the specific wear rate (output variable) using the laboratory experiment

4. Combining the results into a regression equation and checking that the results were statistically

significant

4.2. Measuring shaft sleeve wear profile

Initially an empirical study was undertaken to measure the seal wear in practice and compare it with the

theoretical radial pressure distribution. Random samples of worn shaft sleeves collected from local

copper and lead/zinc mineral processing plants were analysed over a twelve month period with the

objective of determining an empirical wear model and the seal specific wear rate (or dimensional wear

coefficient from the Archard equation (Archard 1953).

Trial experiments were conducted using an M5 laser distance sensor and traversing the beam across

the shaft sleeve surface (Figure 4.1). The inherent linearity accuracy of the laser was in the order of 1

micron; however, this accuracy could not be achieved in practice because of the rough surface of some

shaft sleeves which reflected the light at different angles. This method was changed to measurement

using a linear variable displacement transducer (LVDT).

89

Figure 4.1 Trial laser surface profile measurement

The transducer was mounted in a lathe cross slide and traversed the worn sample which was mounted

in a chuck. The transducer was fitted with a 2 mm ball tip and repeatability was within ± 3 microns.

The surface profile was measured for twenty sleeve samples and the data was downloaded into a

computer via a filter to eliminate noise, and converted to digital data.

4.3. Laboratory experiment design

The research was extended by controlled experimental work in the laboratory to test the validity of the

proposed wear model. Two methods were considered for experimental testing: a test rig with a slurry

pump gland and a Miller test which measures three body reciprocating wear in a standard Miller test rig.

Table 1 shows a comparison of the operating parameters between the test methods.

Table 4.1 Comparison of test methods

Miller Gland Velocity (m/s) 0.16 7 Tribo elements Neoprene lap, slurry, 25.4 x 12.7 mm sample Shaft sleeve, packing, slurry Wear mechanism 3 body abrasion 2 and 3 body abrasion Load (N) Radial 22.24 Axial (200 to 500) and radial Load measurable Yes Yes (axial only) Test Four samples simultaneously Sequential Sleeve profile N/A Yes Time to test (hours) 6 1

90

A comparison between the Miller number reported in the literature and mineral hardness values

revealed an almost linear relationship (Figure 2); consequently the Miller number was not expected to

provide any more value than a measure of relative hardness between the shaft sleeve, packing fibres

and slurry particles (Miller and Miller 1993).

Figure 4.2 Miller number versus hardness

The gland test rig was selected as the preferred method to generate data and determine the

effectiveness and value of the test. Compared to other standard wear tests, using a test rig gland

represents, as closely as possible, the actual two and three body wear conditions; it also avoids the

problems of scale errors by using the same-sized components as a small slurry pump gland seal.

An experimental test rig was designed, constructed and commissioned to test the proposed model. The

purpose of the experiment was to deliberately inject particles into the seal gap and induce wear. This rig

included a standard gland seal arrangement with braided glass fibre packing lubricated with PTFE and a

standard 316 stainless steel shaft sleeve directly coupled to a 2.2 kW motor and rotating with a shaft

surface velocity of 7 m/s. All shaft sleeves were finished by surface grinding with a design surface

roughness of Ra equal to 0.8 microns. The shaft sleeve material (316 stainless steel) was selected to

increase the relative hardness gap with the sample particles.

Mineral property

0500

10001500200025003000

limes

tone

coal

baux

ite

phos

phate rut

ile

haem

atite as

h

magne

tite sand

copp

er ore

alund

um

silico

n carb

ide

Mineral

Ave

rage

M

iller

No.

0200400600800100012001400

Hv Hv

Miller

91

The packing was supplied by Weir Minerals Ltd and is a standard slurry pump packing material code

Q05 which is a cross braided packing, 8 mm square in cross section and with continuous filament

“Fortaglas” yarns, impregnated with PTFE and a mineral oil-based break-in lubricant. This packing is

manufactured by Flexitallic Ltd and was their material code 713L. Because both packing fibres and shaft

materials were constant for the experiments, the relative hardness between them was fixed.

In a standard gland seal, three to four packing rings are fitted, depending on the pump size. For the

experiment the lantern restrictor was not required (because there was no external gland water supply)

and so an additional row of packing was fitted (making a total of five rings) to maximize the shaft sleeve

wear contact length. The motor selection was based on the estimated friction power assuming 600 kPa

axial pressure applied by the gland follower and a friction coefficient of 0.17 when the seal is in new

condition (Karassik 1990). The power for grinding the shaft sleeve was calculated based on a depth of

cut of 10 microns and published constants (Table 4.2) (Kalpakjian 1985).

Table 4.2 Power estimates

Friction power kW 1.5 Grinding power kW 0.6

The motor current draw was checked to be below full load current during commissioning to confirm the

power calculations. It was not within the scope of the research to measure friction power.

Slurry was mixed in a separate tank to a controlled density of 10% solids by weight and pumped into the

seal opposite the gland follower end at up to 1 MPa at 3.3 litres/minute (5 x 10-6 cubic metres/second);

this was similar to the gland water design flow rate for a new seal. Particles were screened below 500

micron by manually sieving before mixing. Also the tank volume (25 litres) was large in relation to the

flow rate so that the slurry was only replaced 8.2 times for the duration of a test which minimised the

effects of attrition (change in particle size). The diameter of the pipe feeding the seal was initially chosen

92

to produce a flow rate close to the settling velocity to prevent any particles settling in the pipes and then

increased to 19 mm to account for the very short pumping distance and residence time. During the

screening experiments a flow rate of 1 litres/minute was trialled but this resulted in bogging of the seal

gap and pipeline; the flow rate was increased to 3.3 litres/minute for the actual tests.

The key components in the test rig were (Figure 4.3):

1. Slurry Tank

2. Screw Pump

3. Pressure Relief Valve

4. Pressure Transducer

5. Pressure Gauge

6. Seal

7. Mixer

Figure 4.3 Experiment process diagram

The development of the test rig took several years because of the novelty of the work and the need to

build robust components that would operate for the duration of the experiments.

Strain gauges were mounted on the gland follower bolts, and a pressure transducer on the seal inlet

with both outputs to an amplifier and digital Picolo recorder.

93

The mixer speed and pump speeds were constant for the experiments and the slurry density was

measured with a Marcy scale. Generally the solids concentration was stable after 10 minutes of mixing.

At a trial slurry density of 40% the number of particles was found to be excessive and caused blockage

of the seal (Figure 4.4).

Figure 4.4 Seal blockage at slurry Cw = 40%

Slurry discharged from the seal was captured and fed back to the tank by gravity (Figure 4.5).

Figure 4.5 Test rig layout

4.4. Lean experiment

A lean thinking approach was applied to the experiments through reducing any waste (materials, time,

cost and variance in results) in the experimental process. A flow chart was developed for the experiment

and refined during the initial screening experiments to reduce the time between experiments, variation

in results, and errors (Figure 4.6).

94

Figure 4.6 Experiment lean process

Specifically, the following improvements were made:

1. Standard tooling was purchased to reduce the time to set up the experiment and then remove

the shaft sleeve from the test rig.

2. A standard wear test record was used to record the raw data from all experiments.

3. The mixing tank was designed with minimal bolting to enable the vessel to be easily removed

thereby reducing the cleaning and changeover time to a different slurry sample.

4. A standard operating procedure was produced for the experiments to minimise any variation in

testing, including the processing of the optical imaging data. This procedure was constantly

refined during the project.

95

5. Quick release hose clamps were used to reduce the time needed to clean hoses between slurry

samples.

6. The seal parts selected were all from a standard 2/1.5 B AH Warman slurry pump with gland

seal so parts were available when required.

7. All hoses were transparent so the slurry was visible during experiments.

8. The mixing tank was designed to provide a positive intake head for the Mono screw feed pump

and a positive head from the test rig drain back to the tank, reducing pumping requirements.

9. The strain gauges were designed for a range of zero to 1 kN (5000 mV) and the preload for

each experiment was in the order of 300 N. This margin reduced the risk of overloading the

gauges during the experiments and the time needed to calibrate them. The gauges were rated

for 2 kN to ensure they were robust. The gauge zeros were checked before each experiment for

drift which could be caused by mechanical damage or water ingress and contributed to error

proofing the process. The error from the change in temperature between the start and end of an

experiment was in the order of 2%.

4.5. Mineral sample range

The problem of gland seal wear is highly complex with a large number of variables and complex

interdependencies. The initial approach was based on a 2k factorial experiment design: two values (low

and high) of each property (k = number of properties) were selected to enable a full statistical

experiment design (Montgomery 1976).

The k factors proposed were particle size, relative hardness and angularity (WDA) and are shown in

Figure 4.7. Since all the factors are quantitative, and there are more samples than factors, a design with

centre points was relevant. Assuming no centre points the experiment design considered was 23 which,

when each experiment was repeated, gave 16 runs or observations. The response variable is specific

wear rate.

96

Figure 4.7 Geometric view of factorial design

If there are centre points at each possible location the number of points is 26 which, when repeated,

gives 52 runs or observations (3 factors at 3 levels repeated is 54 runs).

However, this experimental strategy was found to be not relevant for the following reasons:

1. The process requires factors at a predefined level. For example, particles can be sieved or

separated by other means such as cycloning or milled to achieve a defined size; however, this

was found to be expensive for a small sample and not within the project budget. A “one factor at

a time” approach was rejected because this approach did not consider the dependency

between the controlling variables.

2. It is not possible to modify the fracture toughness property of a sample because it is intrinsic to

the mineral; hence the range was limited by the range of minerals commercially available.

3. The laboratory experiments were labour-intensive, requiring a high level of supervision and set

up time and so the number of runs had to be limited.

4. Many commercial minerals are milled and screened to a finely-graded particle size less than

150 micron. Larger samples were not always available from mining sources because of the

complexity of transporting hazardous goods and occupational health and safety concerns.

97

5. Fertilisers were a mineral source provided the content was homogenous; however, many were

found to be heterogeneous and this limited the samples available.

6. Samples obtained from local mine sites were generally a wet concentrate with a very broad

range of particles sizes. It was not possible to dry the concentrate effectively and screen out

particles greater than 500 microns and excluded from testing. Haematite and copper bearing

ores from local mineral processing operations were difficult to obtain in the target size range

less than 500 microns.

7. Although there are many processing plants in Australia, many of the mineral products are

hazardous (eg, lead sulphide or galena) and could not be transported without state government

approval and were subject to hazardous goods legislation. This added additional complexity to

the choice of the mineral sample range.

Consequently the experimental strategy developed towards testing commercially available minerals

which covered the maximum extent of the Mohs hardness range (excluding diamonds) and then finding

a regression model and testing for significance of the model and interaction between the controlling

variables; ie, the experimental strategy was to vary all the factors together.

A wide range of minerals was tested within the constraints of cost and the commercial availability of

particle samples. The experiments targeted the upper and lower limits of the properties that influence

the specific wear rate as summarised in Table 4.3.

98

Table 4.3 Range of particle properties

Factor Lower level Upper level Range Roundness factor 1.0 2.0 1.0 Fracture toughness (MPa.m0.5) 0.1 4.1 4.0 Hardness (Moh) 1 9.5 8.5 Size (micron) 5 500 495

The particle angularity (dihedral angle) was quantified by optical scanning of a range of the particles

across the sample size, digitizing the perimeter then deducing an average of the roundness factor. All of

the samples were checked for crystal homogeneity by x ray diffraction.

The fracture toughness values were found from work reported by Broz et al. Values for gypsum were

not available and so the values for anhydrite were used as an approximation; similarly, calcite values

were used as an approximation for dolomite (Sorrell 1973). No information was found on talc

(magnesium silicate), other than that experimental values are difficult to determine because no cracks

are formed during depth sensing indentation (Broz, Cook et al. 2006), and a low value of 0.1 MPa.m0.5

was assumed for the purpose of the regression analysis.

The hardness values were estimated by averaging the Mohs hardness published for each mineral and

converting to a Vickers number and Pascals (Nesse 2000).

4.6. Measuring wear

The shaft sleeve was weighed before and after each test and the mass loss converted to a volumetric

loss for calculation of the specific wear rate (SWR). SWR was defined as:

)...../().10()./( 93 loadtimeRPMdmasslossmNmmSWR πρ= Equation 4.1

Where:

mass loss = shaft sleeve mass loss [kg]

mm3 = m3 x 109 conversion factor

99

ρ = shaft sleeve density [kg/m3]

d = shaft sleeve diameter [m]

RPM = speed of the motor shaft

time = experiment duration [min]

load = average applied axial load [N]

Specific wear rate was defined for the purpose of the experiments based on normalising the shaft

sleeve wear rate with the average gland follower load, recognising that gland follower load affects the

applied packing stress and particle motion in the seal gap. It is not possible to determine the specific

load on each particle per se.

Experiments were conducted over a 62.5 minute duration with 125 data points at 30 second intervals.

The average load was found using the trapezoidal rule for area under the force versus time curve.

The wear profile across the shaft sleeve surface was not measured as the mass loss was only in the

order of several grams for each test and excessive test times in the order of days would be required for

a significant profile result to compare with the field data.

100

5. Wear test results

5.1. Field wear data

Figure 5.1 illustrates the wear profile versus normalised length for four typical samples (a small number

has been chosen for clarity). It was apparent that the maximum localised wear does not occur at the wet

end side of the shaft sleeve and consequently there are other variables to be considered in the wear

model.

The test 1 data is neither uniform nor exponential (see Equation 2.1). The test 2 data does demonstrate

a higher wear at the wet end but the profile is not exponential. Likewise, the test 3 data indicates wear is

not uniform and fails to follow any exponential relationship. Periodicity in the test 4 shaft sleeve profile

indicates that the packing bulges and wears more at the packing centre. However, such behaviour was

not frequently observed in the samples. This behaviour may be related to the packing fibre lattice

construction: extra fibres are located at the packing edges to reduce the risk of mechanical damage and

maintain shape which means the packing properties are not uniform across the section. It appears that

the rows of packing are not a homogenous triboelement and are effectively discontinuous. The same

data also indicates increased wear in the centre of the shaft sleeve rather than the wet end – packing

Figure 5.1 Shaft sleeve profile

Shaft sleeve profile

-4

-3.5

-3

-2.5

-2

-1.5

-1

-0.5

0

0.5

0 20 40 60 80 100 120

% from drive end

Pro

file

(mm

)

Test 1Test 2Test 3Test 4

0% Drive end

100% Wet end

101

contact which does not fit the Denny exponential pressure distribution model suggesting there are other

variables in the wear of the shaft sleeve.

Four field samples were collected in failure sequence from a thickener underflow pump in a lead/zinc

processing plant. Again the data does not display any pattern (Figure 5.2).

There are several limitations with this approach based on collecting samples (site data) from mineral

processing operations including:

1. Both the gland follower load and gland adjustment are unknown as the routine maintenance

schedule was not logged and the load was not measured.

2. Packing and shaft sleeve useful life are rarely recorded even for world best practice operations.

Hence the distance travelled (product of speed of rotation and diameter) cannot be determined

with any reliability.

Figure 5.2 Lead/zinc thickener underflow shaft sleeve wear profile

S haft s leeve w ear p ro file

-5

-4.5

-4

-3.5

-3

-2.5

-2

-1.5

-1

-0.5

0

0.5

0 20 40 60 80 100 120

% from driv e e nd

Wea

r (m

m) Tes t 1

Tes t 2Tes t 3Tes t 4

Higher w ear at w et end in general

D isc ontinuous w ear profile

1234 hours

1275hours

Life unknow n

0% D rive end 100% W et end

102

3. Gland seal repair maintenance is generally unplanned or within a few days of functional failure

which renders collection of data problematic.

4. Information about the tribological system is not captured at the time of failure eg, the slurry

sample which could yield valuable data about the mineral and gangue particle size, particle

fracture toughness and shape.

5. Packing is treated as a commodity by mineral processing operators; consequently stock tends to

be uncontrolled in satellite stores and failed packing is scrapped unless a routine collection

system is in place.

6. Normally the shaft sleeves and packing are replaced at seal failure or during pump wet end

maintenance (opportunity-based). Hence, the shaft sleeve life is some random (and unknown)

percentage of the useful life; ie, they may be at potential failure. This means that the opportunity

to establish the wear profile over the total life of the shaft sleeve and packing is not available for

comparison with the Czichos model and phases of wear over time (Czichos 1978).

The problem of meaningful measurement of failures in mineral processing operations and collection of

accurate maintenance history is not uncommon. Holmberg summarised the issue stating: “The most

promising approach to collecting accurate history data that can be used for endurance life estimation is

using automatic and operating condition monitoring modules directly connected to performing

components. Then there is no human influence…” (Holmberg 2001) There are developments by mineral

process operators in progress to link maintenance management systems (SAP®) with control systems

such as Cytec ® which will aid in identifying operating hours for components but these advances are not

widespread. Consequently the field empirical approach proved that additional variables other than the

exponential packing radial stress are important, and a new and controlled experimental methodology

was required to validate the wear model in the hypothesis.

103

5.2. Repeatability and variance

The initial laboratory experiments were designed to screen the factors and seven replicates for sand

(quartz) were used to prove repeatability (Figure 3). The coefficient of variation was found to be 23.7%

for the specific wear rate which is typical for wear experiments with a wide range of controlling variables.

In comparison the coefficient of variation may be as high as 35% in erosive wear based on actual site

pump wear data, hence the repeatability is reasonable. (Walker 2001).

Figure 5.3 Repeatability of the experiment during screening

A measurable wear result was achieved within minutes of testing and continuous slurry flow proved

practical at the small scale. The experiments were conducted over a 62.5 minute period and mass

losses varied between 1.5 and 6 grams. A typical wear sample is shown in Figure 5.4

The effects of attrition are well known in laboratory testing and, due to the relatively small flow rate in

relation to the tank volume, the particle properties were not compared before and after testing; ie,

attrition effects were ignored. The shaft sleeve was the same material (316 stainless steel) for all

experiments and the passive layer was not measured.

SWR (quartz)

0.00E+00

5.00E-05

1.00E-04

1.50E-04

2.00E-04

2.50E-04

3.00E-04

1 2 6 7 8 9 10

Test

SW

R m

m3/

N.m

104

Figure 5.4 Wear sample shaft sleeve

The specific wear rate was found to range between 9.23 x 10-6 mm3/Nm (test 15) and 3.59 x 10-4

mm3/Nm (test 3).

5.3. Effect of gland follower load

The effect of the gland follower load applied by the bolts either side of the gland follower was observed

during the screening experiments with coarse sand. The peak pressure was significantly lower when

zero load was applied to the gland follower. Figure 5.5 demonstrates an exponential decay in pressure

which is consistent with the work by Bohner and this simple experiment confirmed the need to measure

the gland load so results between the experiments could be compared and that the initial pressure could

be set close to actual pressures of typical slurry pumping in the order of 600 kPa (Bohner, Blenke et al.

1975).

105

Figure 5.5 Seal pressure versus time

5.4. Effect of relative hardness

Published values of average Mohs mineral hardness, converted to Vickers hardness, were used to

determine the relative hardness (Ha/Hs) of the shaft sleeve wear specimen. The results indicated

specific wear rate increases with the relative hardness (particle hardness/shaft sleeve hardness)

according to a power law relationship with coefficient of regression equal to 0.4762 (Figure 5.6).

Gland pressure at 6.3 l/min

-20

0

20

40

60

80

100

120

0 10 20 30 40

Time (minutes)

Pres

sure

(kP

a)Test 1Test 2

106

Figure 5.6 SWR versus Relative hardness

Test 17 (silicon carbide) specific wear rate was slightly lower than test 3 (corundum) which was not

consistent with their hardness difference and similar size. Tests were conducted for a soluble mineral

(malachite) which, in effect, was a slurry without particles and the specific wear rate was found to be

very low at 3.88 x 10-6 mm3/Nm. The density was constant at 10% by weight for the experiments and a

single random experiment at 5% indicated a reduction in specific wear rate by an order of magnitude for

test 12 (magnetite).

It was observed that the wear profile along the shaft sleeve was not sufficient in wear depth for all

experiments for the wear to be quantified as a function of the shaft sleeve length. This would have

required experimental times in the order of several days for comparison with field data.

SWR vs Ha/Hs y = 5E-05x0.6358R2 = 0.4762

1.00E-05

1.00E-04

1.00E-03

0.10 1.00 10.00 100.00

Relative Hardness (Ha/Hs)

SWR

(mm

^3/N

m)

107

The non dimensional wear coefficient (from the Archard equation 2.4) ranged between 1.81 x 10-5

mm3/Nm (test 15) and 1.23 x 10-02 mm3/Nm (test 3) which is consistent with the two body and three

body coefficients reported by Williams and strongly indicates a range of wear processes exists in a

continuum between two body and three body wear proposed in the wear model (Williams 1997).

5.5. Effect of Critical gap

During the screening experiments it was found that the gap between the gland follower and shaft sleeve

was critical to eliminate wear in this region, and so the gap was set at 2 mm on radius and the thickness

of the compression ring reduced to 500 micron to enable particles to be discharged from the seal.

The wear rates during the commissioning process were observed to be an order of magnitude higher

between the gland follower and shaft sleeve for small gaps. It is thought that the particles are trapped

between the shaft sleeve and the gland follower resulting in a two body micro machining mechanism.

Williams et al found that there is a relative hardness effect (hardness paradox), typically in journal

bearings: particles can become embedded in a softer surface and increase the wear rate of a harder

surface (Williams and Hyncica 2001). Intuitively it was thought that particles would embed in the packing

fibre structure and contribute to two and three body wear mechanisms; however, during removal of the

packing after experiments, particles were not observed to be embedded. It is thought that the elastic

deformation of the packing surface allows particles to tumble through the gap which is typical of three

body wear (Figure 5.7). Particles are acted upon by a shear force caused by the rotating shaft sleeve

and normal loads from the packing compression and may rotate and/or slide. If any force equilibrium

exists between the particles, it is only momentary as the deflection of the elastic packing surface shifts

the position of force action on each particle and establishes torque (moments) for each.

108

Figure 5.7 Three body wear model for hard and soft surfaces

When both surfaces are inelastic and there is a small gap between shaft sleeve and gland follower, the

three body mechanism is different and it is thought that the particles become trapped and wedged

between the surfaces and establish a conglomerate or quasi particle which results in micromachining of

both surfaces. The particles are effectively locked together in a force equilibrium until the geometry of

the conglomerate is altered. Note that the particles have axial and radial force components. Particles

with small dihedral angles (and high attack angles) in contact with the hard surface experience high

Hertzian stress and fracture, resulting in smaller particles and new vertices being exposed for

machining.

Particles then move towards the low pressure (atmosphere) side of the seal to irreversible departure,

discharged from the seal and are replaced by new particles entering the gap. These new particles

establish another force equilibrium and the process is repeated. This three body process is shown in a

new model in Figure 5.8.

109

Figure 5.8 Three body wear model for hard surfaces

Hence the relative elasticity (a dimensionless property) of the surfaces is thought to be an important

controlling variable in three body wear and requires more research.

It is expected that the smaller particles are more likely to be free to rotate in the gap and contribute to

three body wear. This is similar to a lapping, wet grinding process where the water is used to flush out

the debris from wear.

As found in the site samples, the packing mechanical property is not uniform across the length of the

seal hence, when quasi particles do form and present new vertices for orthogonal cutting, deeper cuts

are machined in proximity to the packing lattice where it is most elastic thought to be halfway along the

packing cross section. An example of these deeper cuts is shown in Figure 5.9. In this example there is

highest wear adjacent the lantern gland water distribution ring.

110

Figure 5.9 Shaft sleeve wear (DE = drive end)

5.6. Effect of Hydraulic performance

It was observed that the packing gland load relaxed after a relatively short time (in the order of 5

minutes) in an exponential manner (Figure 5.10). This is consistent with/in accordance with the Bohner

packing deformation model. The maximum gland pressure ranged between 50 and 1000 kPa without

any clear relationship between particle size and pressure.

Figure 5.10 Typical Pressure and load versus time

050

100150200250300350

0 50 100

Time (min)

Pack

ing Lo

ad (N

)

050100150200250300350

Inlet

Pres

sure

(kPa)

Total Packing Load (N)Inlet Pressure (kPa)

111

5.7. Effect of fracture toughness

It was found there was a correlation of specific wear rate with published values of fracture toughness,

based on the stress intensity factor. The method of quantifying fracture toughness for particles by

microindentation is a laborious process and is subject to variance through selection of the cleavage

plane, mechanical properties also vary with direction (Figure 5.11). Consequently the property of particle

fracture toughness was included in the proposed model.

Figure 5.11 SWR versus fracture toughness

5.8. Effect of Particle size

The experimental particle size range included particles typically from 5 to 500 microns and it was found

that generally the specific wear rate increases with the equivalent particle size d85 and this is consistent

with the size effect reported by Misra et al (Misra and Finnie 1981).. Individual particle sizes were

measured by AnalySis software which deduces the arithmetic mean of all diameters of a particle for

angles in the range from 0 to 179 degrees with step width of one degree. The particle sizes were sorted

by frequency to determine the 85% passing size for each sample. This relationship is best approximated

with a log function (Figure 5.12).

SWR vs Fracture Toughness y = 0.0002x0.7616

R2 = 0.3837

1.00E-06

1.00E-05

1.00E-04

1.00E-03

0.1 1 10

Stress intensity factor

SWR

(mm

^3/N

m)

112

Figure 5.12 SWR versus particle size

Peak gland pressures reached 1 MPa during some of the experiments with fine particle slurries;

however, no clear correlation could be defined based on particle size alone (Figure 5.13). It was

expected that the coarser corundum particles would result in high peak pressures and form larger quasi

particles more readily. Generally it was observed that the finer slurries resulted in higher peak start up

pressures which is consistent with a higher packing density and additional yield stress effect for flow to

occur; ie, it is thought that with the finer particle slurries the rheology of the slurry is also a contributor to

the peak pressure. There was also little correlation between gland follower peak load and the particle

size. The notable exceptions were test 4 (talc) and test 15 (copper oxide) which had high initial packing

loads and high peak pressures. For these tests the start-up packing load was in the order of 300 N and

the gland pressure increased within several minutes to the order of 1 MPa before exponential decay

commenced. It is thought that the rheology of these fine particle slurries was a significant contribution to

the high pressures. It is deduced then that for fine particles a gland seal is, in some cases, a more

effective barrier for preventing leakage and particles are less likely to flow across the gap. There were

tests of different particles with similar size which had very low packing loads and pressures as

demonstrated by the scatter in Figure 5.13. This was a key finding.

SWR vs d85

1.00E-06

1.00E-05

1.00E-04

1.00E-030 100 200 300 400 500 600

d85 (micron)

mm

3/N

m

113

Figure 5.13 Pressure versus particle size and load

5.9. Effect of particle shape

5.9.1. Particle shape characterisation

The literature review provided some direction on particle shape characterization; however, it is a science

still in development. Because there was not a clear comparison between all the methods available and

analysis of variance, it was thought that the development of a new parameter and its comparison with

contemporary methods was needed to bring a higher level of definition to particle shape enabling its

relevance to the wear model to be explored.

During the screening experiments a number of parameters were explored to quantify the particle shape

perimeter in a two dimensional plane (particles are anisotropic and three dimensional in nature). They

included spike parameter linear fit (SPL) and spike parameter quadratic (SPQ) fit and two new

parameters, spike parameter quadratic linear fit (SPQL) and weighted dihedral angle (WDA). The fractal

dimension was found to be of low value and excluded from the parameter investigation because a high

fractal dimension, indicating an irregular boundary, may not represent vertices with small dihedral

Peak pressure vs size and load

0.00

100.00

200.00

300.00

400.00

500.00

600.00

700.00

800.00

900.00

1000.00

0 100 200 300 400 500 600

d85 micron

kPa

0.00

100.00

200.00

300.00

400.00

500.00

600.00

700.00

800.00

900.00

Load

(N)

PressureLoad

114

angles that are relevant to cutting wear by sharp particles in a seal system. This was confirmed by

personal discussion with Dr Pawel Podsiadlo of the University of Western Australia.

The SPL is based on Richardson’s technique and constructs a series of triangles of different step sizes

when walking around the particle perimeter and using all possible starting points (Figure 5.14). The

cosine rule is used to calculate the spike dihedral angle and internal angles and the spike height SV.

The spike value is normalised by the spike height and the maximum SV for each step length averaged.

This was repeated for all particles and averaged again for the sample.

Figure 5.14 SPL boundary walk of a particle

The SPQ estimates the equivalent circle having the same area and centroid as the particle (Figure

5.15). All spikes outside the circle are considered of interest and the SV is calculated using quadratic

equations.

115

Figure 5.15 Schematic illustrating SPQ calculation method, circle of equivalent area and intersections

A new parameter was proposed by Oram through combining the SPQ and SPL methods which have

inherent limitations. The equivalent circle diameter is plotted on the particle centroid and the SPL

method is applied to the coordinates outside the circle (Oram 2007).

The SPL method is limited by the fact that some vertices with small dihedral angles that cause cutting

wear will not always be captured within the boundary walk. Conversely some insignificant boundary

features which are not likely to cause wear can also be captured, making the computing time excessive.

The SPQ method, in comparison, measures only the relevant vertices and has a relatively shorter

computing time. There were some limitations found when there is a change in direction, for example

when there are two vertices close together (two humps) and the quadratic equations don’t fit the change

in slope (Figure 5.16).

116

Figure 5.16 SPQ poor quadratic fit

In addition when there is a rounded vertice the SPQ finds one large dihedral angle and the SPL finds

two angles as shown in Figure 5.17.

Figure 5.17 SPL and SPQ variance

The theoretical value range for SPQ and SPL shown in Figure 5.18 indicated that to make a meaningful

comparison between the methods a phase shift was required in the SPQ by 0.5 hence the equation for

ASPQ was:

cos 0.52cos 0.5,

2aspq ASPQ

nθ

θθ

⎛ ⎞⎛ ⎞ +⎜ ⎟⎜ ⎟⎛ ⎞ ⎝ ⎠⎝ ⎠= + =⎜ ⎟⎝ ⎠

∑ Equation 5.1

This correction was made to ensure a small dihedral angle equated to a high angularity value as shown

in Figure 5.18.

117

Figure 5.18 Angularity versus angle range

A new parameter Weighted Dihedral Angle (WDA) similar to SPQ was proposed in which only the

relevant vertices outside the equivalent circle diameter are considered (Figure 5.19). Each vertice angle

is quantified based on Cartesian coordinates of the particle boundary from SEM imaging and

trigonometry rather than solving a quadratic equation. It is important to note the angle is not normalized

by the “spike height” as the angle is the critical factor in orthogonal cutting mechanics as found by

Williams et al. The angle is also averaged over the size ranges 0 to 10, 11 to 100 and 101 to 500

microns, recognising that shape and size are both statistical distributions and dependent (see Section

2.8).

-0.6

-0.4

-0.2

0

0.2

0.4

0.6

0.8

1

1.2

0 0.5 1 1.5 2 2.5 3 3.5

Angle (rad)

Ang

ular

ity

SPQSPLAdj. SPQ

Figure 5.19 Proposed parameter WDA

118

Random samples of mineral (haematite, coarse sand, gypsum corundum and apatite) were selected

with particle sizes less than 500 microns and sieved manually into the three size ranges with

approximately 10 particles in each fraction. Demineralised water was added to each sample and the fine

clay particles were decanted several times before the final concentrate was dried. The particles were

viewed with an Optical Microscope MZ316 and the particle boundary quantified with iTems software

resulting in Cartesian coordinates, shape area and centroid coordinates.

It was found that the ASPQ parameter was consistently in the order of 50% lower than the SPL. Another

significant finding was the insignificant variation between the samples (Figure 5.20).

Figure 5.20 Particle angularity

The WDA parameter also found a similar result to the spike parameters with approximately four degrees

variation in the data (Figure 5.21).

0.20

0.25

0.30

0.35

0.40

0.45

0.50

0.55

0.60

Alumina Coarse Sand Hematitie

Part

icle

Ang

ular

ity

ASPQ

SPL

119

Figure 5.21 WDA results of random samples

The corundum sample was expected to be the sharpest sample as this material is commonly employed

for sand blasting processes and felt reasonably sharp to touch. A visual overview of the particle

boundary images supports the data and accentuates that there is little difference between the samples

but a wide variation of particle shapes within a sample. There is a mix of prolate and oblate perimeter

shapes within any sample and this was a key finding (Figure 5.22).

To improve accuracy a larger population of particles would need to be examined but this was not

possible because of the scanning time and imaging costs.

84.0

85.0

86.0

87.0

88.0

89.0

Alumina Coarse Sand HematiteA

ngle

(°)

120

Figure 5.22 Variation in particle shape (Oram 2007)

A histogram of the particle parameter for each sample was investigated to check for any bias over the

range and the alumina results indicate that all of the methods consistently result in similar shape

distributions. Generally it was found that for any given sample the median shape range is between 60

and 120 degrees and was not an important controlling variable for the wear model (Figure 5.23).

121

In summary the particle shape parameters were found to be very time consuming to quantify and the

results showed that, on a two dimensional basis, there was not a significant variation between the

samples. Consequently, based on the angularity results from the parameter data found in the size range

up to 500 micron and the literature review summary revealing the complexity of quantification, it was

Figure 5.23 Angularity frequency (Oram 2007)

122

determined that particle shape was of limited value to the wear model and a standard shape descriptor,

form factor (reciprocal of roundness factor), was evaluated for the experimental range of samples.

Particles with small form factors are less round and spherical particles have a factor of one. This

standard shape data was computed from the optical images with minimal computing time and was

analysed as a final investigation of the effects of particle shape on seal wear.

5.9.2. Particle shape data

The relationship between specific wear rate and particle shape was found to be constant across the

range of shapes in the experiments based on roundness (inverse of the form factor). Particles less than

500 microns in size were scanned using a Leica MZ16FA stereo microscope and recorded using a

Leica IC3D camera and Leica Application Suite software. Fine particles were scattered on aluminium

SEM carbon sticking tabs and coated with 6 nm of platinum to make them conducting. They were then

imaged using an Phillips XL20 scanning electron microscope. Particles greater than 5 micron in size

were examined by optical microscope with a maximum magnification of 115 times, and less than 5

micron by SEM. The particle shape and size was quantified using AnalySIS image analysis software.

Test 18 (glass beads) with a shape factor of 1.08 resulted in one of the lowest wear rates (7.37 x 10-6

mm3/Nm); however, the results clearly indicate that particle shape for the size range was not a relevant

factor to the model (Figure 5.24).

123

Figure 5.24 SWR versus Particle shape (roundness)

Comparing the data for all of the methods highlights the relative shape similarity between the samples

based on the average of each sample. This also supports the qualitative finding that there was a range

of oblate and prolate shapes within each sample and that the average shape was similar between the

samples (Figure 5.25). Note that for comparison purposes, WDA in degrees has been converted into a

shape value by using the equation for ASPQ.

It is recognized that form factor or roundness does not have the inherent relevance to orthogonal cutting

wear as the SPQ; however, the results demonstrate that the extrinsic property of shape was not relevant

to the shaft sleeve SWR and that the form factor was similar in accuracy to the other descriptors.

SWR vs Shape

1.00E-06

1.00E-04

1.00E-02

1.00E+000 0.5 1 1.5 2 2.5

Roundness

mm3/N

m

124

Figure 5.25 Shape descriptor comparison

5.10. Effect of crystal content

All of the samples were analysed by X ray diffraction. They were found to be of homogeneous crystal

content with the exception of the apatite sample which was contaminated with sand; consequently this

result was excluded from the regression equation. The SWR was reviewed as a function of the crystal

form and it was found that there was no relationship between them (Table 5.1). In practice the minerals

found in mineral processing and to which sealing systems in slurry pumps are exposed may or may not

present perfect crystal faces and shapes depending on how they are formed and grow, hence many

crystals are not well formed or euhedral (Nesse 2000).

Shape Descriptor Comparison

00.10.20.30.40.50.60.70.8

Alumina Apatite Gypsum Sand

Mineral

Shap

e valu

eWDASPQFF

125

Table 5.1 Crystal system for mineral samples

Test No Mineral Crystal System Formula 1 Coarse Sand Hexagonal(trigonal division) SiO2 2 Coarse Sand Hexagonal(trigonal division) SiO2 3 Alumina Hexagonal(trigonal division) Al203 4 Talc Monoclinic Mg3Si4O10(OH)2 5 Apatite Hexagonal(trigonal division) Ca Mg(CO3)2 6 Beach sand Hexagonal(trigonal division) SiO2 7 Unimin 30/60 Hexagonal(trigonal division) SiO2 8 Unimin 30/60 Hexagonal(trigonal division) SiO2 9 Unimin 30/60 Hexagonal(trigonal division) SiO2 10 30/60 50N mix Hexagonal(trigonal division) SiO2 11 Magnesite Hexagonal (rhombohedral) MgCO3 12 Magnetite Isometric FeFe2O4 13 Dolomite Hexagonal (rhombohedral) CaMg(CO3)2 14 Gypsum Monoclinic CaSO4.2H2O 15 Copper oxide (cupric oxide) Isometric CuO 16 Manganese oxide (manganite) Monoclinic MnO 17 Silicon Carbide (carborundum) Hexagonal SiC 18 Glass beads Amorphous - n/a SiO2 19 Malachite Monoclinic Cu2CO3(OH)2 20 Magnetite Hexagonal (rhombohedral) FeFe2O4

126

The particles were observed and their fracture type, crystal faces, and habit are summarised in Table

5.2 .

Table 5.2 Particle crystal properties

Test No Mineral Fracture Crystal face Habit1 Coarse Sand hackly subhedral equant2 Coarse Sand hackly subhedral equant3 Alumina conchoidal subhedral equant

4 Talcbasal

cleavage anhedral platy, equant5 Apatite hackly subhedral equant6 Beach sand uneven anhedral spherical/equant rounded7 Unimin 30/60 uneven anhedral equant rounded8 Unimin 30/60 uneven anhedral equant rounded9 Unimin 30/60 uneven anhedral equant rounded

10 30/60 50N mix uneven anhedral equant rounded11 Magnesite conchoidal anhedral equant12 Magnetite hackly anhedral equant13 Dolomite hackly subhedral equant14 Gypsum splintery subhedral platy15 Copper oxide (cupric oxide) no cleavage anhedral spherical tufts16 Manganese oxide (manganite) uneven anhedral equant17 Silicon Carbide (carborundum) conchoidal subhedral equant18 Glass beads n/a n/a equant spherical19 Malachite n/a anhedral fibrous, acicular20 Magnetite hackly anhedral equant

It was expected that particles with sharp edges formed by hackly or conchoidal fracture, well-formed

crystal faces and equant habit would contribute to high SWR (highlighted cells in Table 5.2). Conversely

it was expected that particles with anhedral shape and fibrous in habit would be related to lower SWR.

Generally it was found that all particles did not exhibit euhedral crystal form either due to weathering or

irregular fracture caused by impurities and defects in the crystals and mutual interference during

formation, hence this descriptor was not useful for the wear model. The quartz samples exhibited a

range of habits from equant to rounded as can be seen in the SEM images (Figure 5.26).

127

Figure 5.26 Test 10: Sand mix equant and rounded particles

Test 3 (alumina) and 17 (silicon carbide) revealed conchoidal fracture and sharp edges and vertices

which would contribute to orthogonal cutting of the shaft sleeve and relatively high SWR (Figure 5.27

and Figure 5.28).

Figure 5.27 Test 3: Alumina

128

Figure 5.28 Test 17: Silicon carbide It is important to note that while the test 20 (magnetite) and test 13 (dolomite) had all the indications of

sharp particles, with hackly and equant habit, the SWR results were not high and this is a function of

their other properties of fracture toughness and size (Figure 5.29 and Figure 5.30).

Figure 5.29 Test 20: Magnetite

129

Figure 5.30 Test 13: Dolomite

The magnetite particles were very fine, and the dolomite has a relatively low if insignificant fracture

toughness. It is thought that the particles of low fracture toughness attrite into smaller particles until they

reach a low particle size limit where the fracture energy available is less than the molecular bond

strength.

5.11. Packing observations

The packing material from each experiment was examined to identify the embedment of any particles in

the braided lattice; despite the intuitive appeal of this particle motion, the number of particles contained

in the packing was negligible. Samples of new (as received) packing were selected at random and

examined under SEM which showed the glass fibres and PTFE boundary lubricant (Figure 5.31).

130

The observation of used packing showed very fine particles (less than 5 microns) bound by the lubricant

and fibres in the test packing samples (Figure 5.32) along with broken glass fibres. The important point

is that the observation is post factum and the evidence for particles greater than 5 micron size is

destroyed when removing the packing from the stuffing box housing. The observation also supports the

finding that particles and quasi particles are trapped in the gap by a balance of forces in the gap rather

than the inherent geometry of the braided lattice.

Figure 5.31 New packing SEM view

Figure 5.32 Packing after test SEM view

131

Consequently there was little value in recording pictures of each packing sample since bound particles

and quasi particles were not evident after the experiments.

Each packing was examined and two examples are included as typical observations. Test 10 (Unimin

sand mix) showed tearing of the glass fibres and indication of the particles jamming on one side of the

seal gap (Figure 5.33).

Figure 5.33 Test 10 packing

Test 14 (gypsum) showed the effects of heat and discolouration of the lubricant caused by excessive

friction (Figure 5,34). The lubricant contains a combination of hydrocarbon and PTFE, and it is thought

that the hydrocarbon component reached its critical combustion temperature in this example.

132

Figure 5.34 Test 14 packing

5.12. Seal Dimensional Analysis

From the project hypothesis the seal functional relation is proposed as:

The dimensions for fracture toughness (KIC) are simplified to MT-2 (equal to the dimension for GIC) by

excluding the L-1/2 term . KIC is directly related to GIC (see Equation 2.5).

The variables and their dimensions are shown in Table 5.3.

Table 5.3 Seal variables

Variable Description Dimension SWR Specific wear rate (mm3/N.m) = wear volume normalised by

load and travel distance (Vt) and is equivalent to the dimensional wear coefficient in the Archard equation

[LM-1T2]

x Packing length from gland follower [L] P Gland follower applied load to packing [ML-1T-2] h Wear depth at position x [L] V Shaft sleeve surface velocity [LT-1] HShaft Hardness of shaft sleeve [MT-2L-1] HParticles Hardness of particles [MT-2L-1]

),,,,,,,,,(),( 85, dKHHHhVPLtxSWR ICPackingShaftParticles θφ= Equation 5.2

133

HPacking Hardness of packing fibres [MT-2L-1] θ Weighted dihedral angle of slurry particles [1] d85 Equivalent particle size [L] KIC Particle tenacity/fracture toughness [MT-2]

The experiment was designed at the same scale as operating practice, hence geometric similitude was

satisfied and scale effects minimised.

There are eleven (11) variables and three (3) dimensions (M,L,T). By the Π theorem there would be up

to eight (8) parameters and twenty four (24) equations with three (3) unknowns to solve for one set of

repeating units. As there are a large number of variables and repeating unit options this makes the

solution complex. The repeating variables are unknown and solving the equation is by inspection and

substitution.

By definition θ is a function of arc length and radius so:

θπ =1

There is interaction between the shaft sleeve, packing and particles. The relationship between particles

and the packing is not a significant contribution to the model and is ignored, so by inspection:

ShaftPacking HH /2 =π

ShaftParticle HH /3 =π

The particle motion and forces are a function of the particle size and gap between the packing and shaft

sleeve so:

hd /854 =π

(and d85/L by inspection is also dimensionless which is ignored as h is a function of L by an exponential

relationship for the packing according to the Denny equation). The variable h is in effect the seal gap

and the ratio indicates a critical gap:

134

d85 = [L] and d85/V = [T]

HParticle = [ML-1T-2] and by substitution:

[M] = HParticled853/V2

SWR = LM-1T2 and substituting for L, M and T, SWR = HParticle-1

ParticleHSWR.5 =π

KIC = [MT-2 ] and by substitution of M and T;

)/.(/. 285

22385 dVVdHParticle=

then ParticleIC HdK .85=

)./( 856 ParticleIC HdK=π

The hardness of the packing is resisted by the gland follower load (ie, the axial force is balanced at

equilibrium), then HPacking = P. (Note hardness may be multiplied by the gravity constant and

approximated to a pressure in SI units of MPa or GPa. The actual hardness indenter surface is not

normal to the surface and the conversion is an approximate equivalent only.)

PackingPacking HPPH //7 ==π (inverted terms)

For the experiment and model, thermal effects [T] on the packing pressure, friction, lubrication and

particle properties are ignored. From experience it is known that the gap between the shaft sleeve and

packing is directly related to the gland follower load or pressure.

Then Pch .= where c is some constant (for example the packing lateral pressure ratio)

Phc /.8 =π resulting in eight proposed Π dimensional groupings.

The dimensional equation is then:

ParticleHSWR.

)/,/,,/,/,/),./(( 8585 PchHPhdHHHHHdKf PackingShaftParticleShaftPackingParticleIC θ=

Combining the HPacking terms;

135

)/,/,,/,/),./(( 8585 PchHPhdHHHdKf ShaftShaftParticleParticleIC θ=

h is not known as experimental measurement is difficult , hence h can be excluded. The equation then

reduces to:

)/,,/),./(( 85 ShaftShaftParticleParticleIC HPHHHdKf θ=

and inverting the first term:

)/,,/,/).(( 85 ShaftShaftParticleICParticle HPHHKHdf θ=

The dimensional analysis found two additional numbers of importance. The latter term relates to the

seal itself and, as the hardness was constant for the experiments, can be taken as the pressure alone.

The SWR was examined as a function of the hydraulic pressure which is directly related to the gland

load and no correlation was found (see Section 1.6).

The relative hardness and particle shape are discussed in Section 1.4 and Section 1.9 respectively.

The remaining number found is in effect a particle dimensionless number (or particle wear number) and

the new number is proposed as:

ICParticle KHdPDN /).( 85= Equation 5.3 (inverse of π6)

Charting the PDN number (based on the particle abrasive hardness in Pascals) revealed a reasonable

correlation with SWR on a log – log basis with a correlation coefficient of 0.5666 and the line of best fit

equation (Figure ).

0510 107)(103 −− ×+×= PDNSWR Equation 5.4

136

PDN vs SWRy = 3E-10x + 7E-05R2 = 0.5666

1.00E-06

1.00E-05

1.00E-04

1.00E-03

1.00E-02

1.00E-01

1.00E+001 10 100 1000 10000 100000 1000000

PDN

SWR

Figure 5.35 SWR versus PDN

This is a significant discovery noting that the correlations of SWR with the controlling variables of

particle shape and relative hardness (both in the dimensional analysis) and fracture toughness had

lower correlation coefficients. The dimensionless analysis may be extended by combining the particle

number which is in effect a material parameter with the relative elasticity of the opposing surfaces which

would, in effect, combine to form a three body wear number.

The correlation coefficients are summarised in Table 5.4:

137

Table 5.4 Correlation summary

Variable R2

PDN 0.5666 Particle size 0.3684 Fracture toughness 0.3837 Relative hardness 0.4762

5.13. Erosive wear and vorticity

It can be shown by solving the Navier Stokes equation that the flow rate in the seal gap (between the

shaft sleeve and packing) is approximately a function of the gap, based on a power law and the

pressure drop across the seal, and the lubricating fluid (gland water) viscosity which is the viscous flow

equation in thin films (Turnbull 1976)

)/.(12/3 dxdPhQ μ−= Equation 5.5 Viscous flow

The seal may also be considered as an annulus (Figure 5.36):

Figure 5.36 Annulus schematic

138

Given small clearances between shaft and packing, flow would expected to be laminar when the seal

components are in as-new condition.

The flow rate Q (m3/s) is given by Hagen Poiseuille Law (Streeter and Wylie 1981):

4H

mean xD PQ v A

128 LΔ⎛ ⎞= = × ⎜ ⎟μ ⎝ ⎠

Equation 5.6 Hagen Poiseuille

Where:

Q = volumetric flow rate (m3/s)

vmean = average velocity in the annulus (m/s)

Ax = X-sectional area (m2)

L = length of shaft sleeve (m)

DH = hydraulic diameter (m)

ΔP = applied pressure loss (Pa)

μ = dynamic fluid viscosity of fluid ( Pa.s)

The hydraulic diameter is given by:

( )( ) ( )

2 2i o

H i oi o

4 D Dcross sec tional areaD 4 4 D D (annular clearance)wetted perimeter D D

π −= × = × = − = δ

−

The corresponding Reynolds number is:

( ) ( )( ) ( )

x H o imean H2 2

o io i

Q A D 4Q D Dv D 4QReD DD D

ρ − ρρ ρ= = = =

μ μ π + μπ − μEquation 5.7 Reynolds Number

As the shaft sleeve wears it is expected that the flow rate will increase to the power of three at least,

and is reasonably sensitive to the gap. Also the Reynolds number is proportional to the seal gap, and

turbulent flow is possible depending on the gap dimension.

When the flow is fully turbulent it is expected that there would be erosive wear of the shaft sleeve as

particles entrained in the slurry randomly impact on the shaft sleeve while there is a component of

velocity that transports the particles from the high pressure region at the pump impeller boss to

atmosphere.

139

The particle motion is complicated by the Taylor vortices induced by the relative motion of the shaft

sleeve to the packing. When the inertial forces acting on a particle exceed the fluid viscous force,

vortices are formed in Taylor Couette flow. As the Taylor number based on the shaft sleeve angular

velocity and seal gap is exceeded, toroidal vortices form and these may be stacked on each other or be

in the form of helixes. A range of complicated flows may exist in the seal gap and the seminal work by

Andereck et al (Andereck, Liu et al. 1986) found that in the flow between independently rotating

cylinders there are a large variety of flow regimes (Figure 5.37).

Figure 5.37 Taylor vortice flow regimes (Andereck, Liu et al. 1986)

The flow transition states were found by defining the Reynolds number for the inner and outer cylinder

diameters. The seal gap was not defined for the experiments as the packing is a visco-elastic material

and there were five packing rings; consequently it was not possible to predict exactly what flow regime

would be expected for a shaft sleeve with significant wear. Erosive wear was not observed in the shaft

sleeve specimens. Future work may be able to deliberately induce the vortices by pre-machining the

shaft sleeve outer diameter and commencing experiments with a measurable gap (ie, no contact

between the packing and shaft sleeve). From an end-user’s perspective, this is an extreme state at

which the seal would be at functional failure.

140

5.14. Summary

A field empirical approach indicated variables other than the packing radial stress are relevant. Twenty

experiments were completed, excluding the screening and replicate testing , on a range of particles

commercially available within 500 micron in size . The experiments were successfully quantified by

mathematical models and supported by observation and repeatability was measured by replicates of the

sand (quartz) sample.

The SWR increases with relative hardness between the sample particle and shaft sleeve according to a

power law relationship, and increases with published values of particle fracture toughness by a power

law relationship. SWR also increases with equivalent particle size d85 and is approximated by a log

function.

Particle shape descriptors were evaluated and found not relevant to the SWR because the samples

contained a range of oblate and prolate shapes.

In summary the experimentation proved that development of a SWR quantitative model was possible

and the methodology of testing.