Smith G.T. Cutting Tool Technology: Industrial Handbook

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e capacity for the workpiece, or even a chuck

component to cause damage, or personal injury is di-

ectly related to its kinetic energy (KE). us, the kin-

etic energy of a body is proportional to the square of

its speed. As a consequence, the capacity of a fast ro-

ating body (i.e. chuck assembly) such as a 1 kg work-

iece rotating at 2,000 rev min

–1

would have a KE of

≈ 7 Joules , whereas at 6,000 rev min

–1

the value would

be 62 Joules – approximately the same as a small cali-

bre (0.25) handgun bullet ballistically hitting a target

t 100 m range! Even though the KE of the workpiece

represents a serious safety hazard, the energy stored in

the clamping jaws is far greater. For example, a chuck

assembly tted with its jaws (i.e. shown in partial sec-

tion in Fig. 229b) of φ2

00 mm rotating at 6,000 rev

min

–1

would produce a KE of 1460 Joules (Fig. 229c),

which is about the same as a 0.44 Magnum handgun

bullet as it ballistically-issues from the muzzle! With

only 33% of the static clamping force remaining at

the maximum chuck rotational speed, there is con-

siderable energy being stored in the rotating chuck

assembly. It is important that operators understand

how to determine safe operational speeds when using

non-standard jaws, coupled to its residual grip, while

having an appreciation of the energy stored in these

rotating parts.

As one might surmise from the application of HSM

turning chucks, friction plays an important role in

terms of the chuck’s actual performance in-service. Of

note, is that a freshly-assembled and greased power

chuck will oen exceed its maximum clamping force

by almost 20%, despite this fact, aer a few weeks of

use, or indeed non-usage, the maximum-rated clamp-

ing force may be reduced to as little as 30%. is

considerable drop-o in clamping performance is at-

tributable to the absence of eective lubricant and the

presence of particulates on the chuck’s sliding surfaces.

Substantial losses can result even when the chuck has

been frequently lubricated, resulting from the wrong

grease, or deposits that are not periodically removed.

In essence, grease comprises of oils and solids that are

bound together with soap. us, the oils lubricate the

sliding surfaces and the purpose of the solids are two-

fold: rstly, they impede the escape of oils when the

pressure between the sliding surfaces would usually

squeeze them away; secondly, they directly lubricate by

a shearing mechanism at higher pressures. So, greases

having a higher solid content would normally produce

optimum clamping performance. However, when em-

ploying the chuck in the regions HSM, the centrifugal

eects act to separate out the solid and oil constitu-

ents of the grease, causing the oils to be thrown out

of the chuck and leaving solid matter remnants. Over

an extended time period, these solids from the grease

collect in various: gaps; recesses; and cavities; combin-

ing with small amounts of debris (e.g. nes and parti-

cles) generated during previous machining operations,

which impair the chuck’s performance. While the ap-

plication of coolant exacerbates this situation still fur-

ther, as it tends to leach-away the grease and accelerate

its break-down, possibly causing corrosive damage to

a very expensive chuck. In order to combat these un-

desirable eects, additional compounds are necessary,

such as polymers that can improve the lubricant’s: co-

hesion; adhesion; and water-resistance.

Special care must be utilised to ensure that a chuck’s

reliable performance occurs when HSM operations are

employed, as they are especially susceptible to cen-

trifugal separation and leaching of grease – by coolant

application. Under such circumstances, it is probably

advisable to monitor the chuck performance over a

period of time and apply grease, or service the chuck

once a certain wear pattern emerges allowing one to

create a specic maintenance schedule. Measurement

of chucking performance is vital and simply applying

grease at regular intervals is no guarantee of its per-

formance.

When a chuck becomes congested with solid mat-

ter – separated grease constituents, it prevents fresh

grease from reaching the critical surfaces, so in eect

provides little, or no improvement. Clamping force

measurements should be taken both before and aer

the application of grease, so ensuring that it is evident

as to whether, or not, the chuck needs to be serviced/

cleaned. Measurements of the static and dynamic

clamping forces can be simply determined using a ‘Ra-

dio Frequency Gripmeter’ (RFG). e RFG essentially

comprises of just a load cell and handset. e load cell

is clamped in the chuck’s jaws and the handset displays

the measured clamping forces (i.e. the ‘grip‘ being

accurate to 1kN per jaw), thus avoiding any lengthy

and time-consuming calculations. ese RFG’s are

available from reputable chuck manufacturers, with a

typical handset being able to store up to 120 separate

readings, having a PC-link to Windows

compatible

soware, allowing graphical trends and further ana-

lyses to be undertaken – as necessary.

In any HSM applications for turning, ‘centrifugally-

balanced chucks’ can be utilised as they incorporate

a counter-balance mass that equalises the centrifu-

Machining and Monitoring Strategies 463

Figure 229. Thermal expansion tooling its operation and high-speed turning chuck details.

464 Chapter 9

gal loss of the jaws – when rotating at typically high

turning speeds (i.e see Fig. 229b for a diagrammatic

cutaway assembly of an HSM quick-change chuck).

ese quick-change chucks incorporate a traditional

edge-style and lever mechanism, that instead of dir-

ectly acting on the jaws, the radial force acts through

the actuator and lever mechanism, prior to transfer-

ring the eort to the jaws. So, when the chuck rotates

at high-speed, the actuators are thrown outward by

centrifugal force, but are restrained from moving by

the lever, which pivots about the central connected

sphere. At the opposite end of the lever, the jaws are

also thrown outward and act to move the pivot in the

opposing direction (i.e. to that of the actuators) – ef-

fectively balancing each other. e performance of the

counter-balanced chuck depends upon the accuracy of

the balance achieved, as the actuator mass is constant

and the top jaw mass being variable depending upon

the particular top jaws in use, thus the state of the

balance will also vary. As a consequence, the clamp-

ing force may fall with rotational speed, or actually

increase with heavy and light jaws, respectively. With

standard hardtop jaws, the clamping force remains al-

most constant across the range of the operating speed,

making it unnecessary to calculate the clamping force

losses. An additional feature is that the static clamping

force can be much lower, since there is no centrifugal

loss. is lower static clamping force application, has

the benet that when turning either thin-walled, or

more delicate workpieces that may otherwise distort

with higher clamping forces, such chucks are unlikely

to aect these components, when an HSM turning

strategy is utilised.

Much more could be said concerning HSM turn-

ing operations, particularly relating to the calculations

and working practices, but it was not the intention

ere, to give a comprehensive account of such tech-

nical aspects, simply a concise account of the antici-

pated problems and possible solutions when turning

at high rotational speeds. In the following section, a

discussion concerning toolholder coupling to the ma-

chine tool’s spindle will be briey reviewed.

9.4.2 Toolholder Design

and Spindle Taper

Introduction

In the past, the taper cone and its associated driving

dogs and pull-stud, provided adequate location and

torque for the cutter assembly when mounted into

the machine tool’s spindle. e tool’s cone taper an-

gle was adequately manufactured so that it perfectly

‘wedged’ into its mating spindle taper and the prob-

lem of the single-contact mechanical interface was not

really exposed as decient, until very high rotational

speeds were being utilised, coupled to much greater

feedrates that the newly-developed tooling geometries

and tool materials could now exploit. In recent years,

both dual- and triple-contact tooling systems have

been introduced, these designs will now be briey re-

viewed.

Dual-Contact Tool/Spindle Design

One of the most signicant developments in maintain-

ing a complete mechanical interface between the tool-

holder and the machine’s spindle was the dual-contact

7/24 taper system

. e CAT Standard incorporates

this 7/24 taper, but also allows simultaneous contact

on both the toolholder’s ange and taper, when HSM

machining is the requirement. By achieving this dual-

contact, the CAT-shank toolholders minimise any

orm inherent imbalance at say, 2,000 rev min

–1

. How-

ever, if the cutter assembly is to be rotated at 10,000 rev

min

–1

, the toolholder must cope with a × 25 increase

in centrifugal force, which may compound any unbal-

ance present in the tooling assembly. Further, if the ro-

tational speed is increased still further, into the HSM

ange, then here, the centrifugal force is × 100 greater

and the onset of considerable imbalance may create

chattering conditions. At such high rotational speeds,

if coolant is utilised in the machining process, the

HSM conditions could develop a vortex around the

cutting tool, that conventional ood coolant pressures

cannot penetrate. In these circumstances, possibly the

only realistic option is to utilise a through-the-spindle

coolant delivery application at pressures of >690 kPa

(i.e. 1,000 psi), coupled to perhaps, micro-ltration of

the coolant with special pipes and couplings. e CAT

system of dual-contact oers reasonable rotational

control of the tooling assembly at moderate-to-high

rotational speeds, as the mechanical interface system

of face-and-cone provides a certain security against

24 ‘Dual-contact 7/24 taper system’ , refers to the taper being to

the 7 inches of taper per 24 inches of length. is 7/24 system

incorporates several Standards: CAT and BT 40- and 50-taper

tooling.

Machining and Monitoring Strategies 465

the onset of imbalance. Typical applications for these

HSM dual-contact systems include: aerospace part

production; precision die and mould making; automo-

tive component production; as well as medical compo-

nent manufacturing.

It is worth digressing somewhat, to explain the situ-

ation of why the single-cone mechanical interface is

simply not eective for HSM production applications.

hen rotational speeds begin to approach 20,000 rev

min

–1

, it is not an unusual occurrence for the single-

contact conventional, or standard CAT V-ange tool-

ing assembly to be eectively sucked into the spindle

(i.e. as there is no mechanical contact at the ange),

this being the result of a combination of the pull-stud

pressure and the machine’s spindle ‘taper swelling’ – due

to the very high centrifugal force acting at such high

rotational speeds. In fact, this minute amount of ‘taper

swelling’ can cause the tool holder to separate from

the spindle’s surface and as a result cause considerable

damage to both the cone’s male and female surfaces.

In order to alleviate this HSM problem and run the

tooling assemblies at even faster rotational speeds, the

HSK dual-contact toolholders were developed, which

will now be briey mentioned.

Hsk Dual-Contact Tooling

ere are a number of toolholder designs that are al-

ternatives to the conventional steep-taper spindle con-

nection. Probably the most popular version for HSM

is the HSK-designed tooling connection (i.e see perti-

nent HSK tooling details in Fig. 126c). HSK toolholder

connections oer simultaneous tment on both the

taper and face, at the front of the spindle. e reason

for their acknowledged popularity amongst the HSM

machining companies, is because the increased rigid-

ity of the joint, coupled with their inherent reduction

in dimensions, compared to the equivalent conven-

tional steep-taper connection. In Fig. 126c, the HSK

8° (included angle) short taper with its gauge face con-

tact and simultaneous taper interference can be seen,

which was designed in Germany to Standard: DIN

69893, being introduced in 1993. HSK is a German

acronym that translates into English as: ‘Hollow short

taper’. us, the HSK connection provides:

•

both high static and dynamic stiness,

•

oering great axial and radial repeatable accuracy,

•

with low mass and stroke,

•

having inner clamping.

erefore, with all these proven design advantages

over conventional spindle connections, it allows the

HSK tooling assemblies to utilise the increased rota-

tional speeds necessary for an HSM strategy.

Triple-Contact Tool/Spindle Design

e triple-contact connection is being oered by a

few toolholder manufacturers (i.e. shown in Fig. 230).

e triple-contact design relies on an inner expand-

ing sleeve which maintains uniform contact between

the machine tool spindle and the: toolholder’s top ta-

per; bottom taper; and ange; this being regardless of

the spindle speed employed. Of particular note is the

nner expanding sleeve which functions particularly

well at high spindle speeds. So, as the centrifugal forces

increase – with higher rotational speeds, it causes the

spindle to grow (i.e. ‘swell’), the toolholder’s spring

mechanism forces the split-cone sleeve to proportion-

ally-expand with the spindle. Further, the expanding

sleeve also acts as a vibration-dampening device. e

expanding sleeve extends the tool’s life on average by

between 300 to 500%, by virtually eliminating vibra-

tion. As a result of this ‘vibration-free interface’ be-

tween the tool and workpiece, it provides smoother

machining of: tool steels; aluminium alloys; plus other

metallic alloys. is triple-contact connection system,

also performs eciently with extra-long tools (i.e see

Fig. 231), notably when utilised on horizontal machin-

ing centres. e main reason for the enhanced triple-

contact tool’s cutting performance with extended

tooling assemblies, is the result of the ‘oating’ inner

sleeve (Fig. 230) which acts to minimise any potential

Z-axis deection, thus maintaining its rotational con-

centricity.

Such triple-contact tooling is not inexpensive to

purchase, but these toolholders really do amortise

their cost, by signicantly extending cutter life, while

improving part production rates. Further, it is claimed

by the tooling manufacturer that the toolholder is

‘

maintenance-free’ , while its spring-mechanism in

‘life-testing’ has achieved upward of one million tool

changes. With the advent of either the double- and

triple-contact systems, enabling contact between the

machine tool’s spindle and the toolholder’s mechanical

interface: top-taper; bottom-taper; plus ange; while

‘eliminating vibration’; this has been achieved under

the unique conditions that arise with today’s HSM and

high-accuracy and precision manufacturing needs.

466 Chapter 9

9.5 Dynamic Balance of

Toolholding Assemblies

Introduction

Balancing tools that are intended for HSM applica-

tions is vitally important and there are quite a few In-

dustrial/Manufacturing engineers and users who do

not really understand the concept of how to achieve

balanced tooling, or why it is really necessary. Either

very long extended tooling required for say, for deep-

pocketing (Fig. 231), or tooling that is out-of-balance,

will more than likely produce: chattering eects; goug-

ing of a step, or face; loss of workpiece accuracy and

precision; not to mention uneven and premature cut-

ter wear. Whenever a new tooling assembly is destined

Figure 230. Triple-contact tool connection

system is ideal for any potential HSM operations.

[Courtesy of Heartech Precision Inc. (HPI)]

Machining and Monitoring Strategies 467

Figure 231. Tool runout (≥10 µm) should be of prime importance when machining deep pockets. [Courtesy of

Sandvik Coromant]

468 Chapter 9

for HSM applications on a workpiece, a balancing

operation needs to be undertaken, this statement is

also true for many sub-HSM applications, particularly

when extended tooling is used for whatever reason

(Fig. 231). In fact, every rotating object (i.e. chuck, or

tooling assembly, etc.), will generate vibration.

As has been explained in the previous section, this

vibration results from a number of sources, but princi-

pally here, from centrifugal forces produced by the ro-

tation of an unbalanced mass. ere are several types

of unbalance that could arise, but here, we are mainly

concerned with what is termed dynamic unbalance,

hich increases by the square of the rotational vel-

ocity. For example, any vibration produced by a tool-

ng assembly at 3,000 rev min

–1

, is × 100 greater than

an identical tooling conguration that is rotating at

00 rev min

–1

. Moreover, what is oen either misun-

derstood, or indeed overlooked, is that any change to

the tooling assembly – no matter how small it might

seem, requires re-balancing! ese tooling modica-

tions include any occasion when a cutting tool is ad-

justed, or changed, or similarly if the toolholder is also

either adjusted, or changed. Such changes to the ‘sta-

tus quo’ of the tooling, will directly aect its ensuing

balance, even minutely when just a ‘few microns’! So

that, these miniscule changes to the tooling’s dynamic

condition, causes a degree of tooling oscillation, hence

an out-of-balance condition – with the likely problems

that this creates.

With the wide variety of tooling that is held in: tool

storage carousels; magazines; turrets; etc.; they must

all be ‘balanceable’ by some means. A range of balanc-

ing techniques can be employed here for either single-,

or dual-plane balancing – more will be said concerning

these eects will be made in the following section. e

techniques utilised in achieving tool balance could in-

clude:

•

‘Hard-balancing’ (i.e. see Fig. 234b) – when the

complete assembly either has to have material re-

moved, or added at a certain part of its assembly.

NB 

e major problem associated with ‘hard-bal-

ancing’ is that if the tooling setup changes, so will

the likely rotating mass change, which will mean

modifying the amount of material to be either

added, or subtracted from this newly-distributed

mass,

•

‘Adjustable balancing rings’ (i.e. see Fig. 232) – by

rotating the twin lower and higher balance rings

either clockwise, or anti-clockwise they minutely

modify the balance-condition, allowing single-

plane balance to be achieved.

NB 

ese matched pair of balance rings are in a

symmetrical state of unbalance (i.e. they are both

‘unbalanced’ to the same degree). Letting the user

adjust the pair to counter any unbalance in the cut-

ting tool/toolholder assembly and locking them

into place – usually achieved on commercially-

available balancing machines (i.e. see Fig. 234a).

e state of unbalance is not merely a subject to the

‘caprice’ of the machine tool operator, a tool assembly’s

balance is given by various quality Standards, such

as ISO 1940/1, or ANSI S2.19 – being basically exact

reections of each other. In the following related sec-

tions, they deal with how and in what manner rotat-

ing cutter assembly balance is achieved, utilising such

HSM balance calculations and associated graphical

details as necessary, from these Standards.

9.5.1 HSM – Problem of Tool Balance

Unbalance of a rotating body (i.e. here we are con-

cerned with a complete tooling assembly), can be

dened as: ‘e condition existing when the principal

mass – axis of inertia – does not coincide with its ro-

tational axis’ (i.e. shown schematically in Fig. 232).

For example, such an undesirable state of aairs can

be comprehended by considering the following situ-

ation: if a φ5

0 mm face mill assembly is rotated at

15,000 rev min

–1

, it will produce a peripheral speed

>240 km hr

–1

, which may prove to be disastrous if it is

unbalanced!

Basically there exists, three types of unbalance con-

ditions for rotating assemblies – such as tooling, these

are:

1. ‘Static unbalance’ – single-plane. is type of un-

balance occurs when the mass does not coincide

with the rotational axis, but is parallel to it and the

force created by such unbalancing, is equal to the

magnitude at both ends of the rotating body. us,

if some relief – metal removal (i.e see Fig. 234b) –

on the toolholder body equal to the out-of-balance

mass that occurs, then a nominal static unbalance is

achieved,

2. ‘Couple unbalance’ – Under these circumstances,

the cutter assembly – mass axis – does not coincide

Machining and Monitoring Strategies 469

Figure 232. The taper tment against runout/eccentricity for a milling cutter and its associated balanced tool-

holder

470 Chapter 9

with the rotational axis, but intersects it at the cen-

tre of gravity of the ‘assembly’s body’. Under such

conditions the force vectors equalise, but are 180°

apart.

3. ‘Dynamic unbalance’ – dual-plane. Such a condi-

tion of the toolholder assembly arises when the axis

does not coincide with the rotational axis and is not

either parallel to, nor intersecting this axis (i.e. see

Fig. 232).

For any rotating tooling assembly, estimating the cut-

ter unbalance is possible using the following variables:

M = cutter/holder mass,

S = mass centre,

e = displacement of mass centre,

r = distance from centre of tooling, to the centre of

gravity of mass (m),

ω = angular velocity,

m = mass unbalance,

U = cutter unbalance,

9549 = a constant.

Determining the relative unbalance (U) of a rotat-

ing tooling assembly, can be found by the following

expression(s):

U = M × e or, alternatively: U = m × r (i).

It is usual to express unbalance in terms of the product

of the mass times distance, typically using the units:

‘g-mm’.

Finding the magnitude of centrifugal force produced

by the rotating tooling assembly with a given unbal-

ance, can be established as follows:

F = U × ω

(ii).

Where: ‘ω’ is the angular velocity in units of radians

sec

–1

e formula to nd ‘ω’ is expressed by:

ω =

 � π � rpm



(iii).

erefore, by combining formulae: (i) and (iii), in (ii),

we can obtain the magnitude of centrifugal force ‘F’ , as

follows:

F = m × r × ( 2 × π × rpm/60)

(iv).

As established in equation (iv), the centrifugal force

caused by tooling unbalance will increase by the

‘square of the speed’ , in a similar manner to the spin-

dle nose taper swelling (i.e. growth) previously men-

tioned. Nonetheless, assuming that this specic tool-

holder initially has a low unbalance, this will become a

problem if the rotational speeds are increased beyond

10,000 rev min

–1

. For example, with most toolholders

exhibiting single-plane unbalance

, research experi-

mentation has shown that the initial unbalance of a

typical tooling assembly will be of the order: 250 g-

mm. When such tooling is rotated at 15,000 rev min

–1

this 250 g-mm of out-of-balance develops a continu-

ous radial force of 642.6 N.

Unbalanced tooling can introduce considerable

detrimental eects on not only the machine tool – this

high centrifugal force causing internal bearing stresses

leading to premature spindle failure, but aects cut-

ter life and degrades workpiece surface texture. Much

of the principal tooling unbalance problems can be

traced-back to several sources, such as:

•

Toolholders of the V-ange type, which might have

dierent depth of drive/slots, these toolholder fea-

tures being part of the inherent design,

•

Toolholders for some end mills and slot-drills, hav-

ing set screws for locking the cutter securely in

place, so due to necessary clearance and the radial

application of the set screw, this creates minute cut-

ter eccentricity – causing unbalance,

•

Out-of-balance caused by an unground V-ange

base,

•

Collet and its collet nut tend to be recurring sources

of unbalance in HSM tool holders.

NB Most of these tool holding-related issues can

be eliminated by simply modifying the tooling de-

sign.

As can be seen from Fig. 232, the marginally eccen-

tric adjustable balance rings can be rotated to adjust

the degree of single-plane balance, with several of the

tooling manufacturers oering diering adjustment

methods for HSM toolholders.

Finally, consideration needs to be given to the level

of balance-quality required and in HSM applications

for example, a milling cutter is expected to withstand

25 ‘Single-plane unbalance’ , relates to the type of unbalance that

occurs in either one of two planes. Namely, the tooling assem-

bly’s single-plane unbalance will be in either its axial, or radial

directional plane.

Machining and Monitoring Strategies 471

both high rotational speeds and associated cutting

forces, thus here it can be considered as a ‘rigid ro-

tating body’. is assumption allows one to use the

ANSI S2.19-1989 Standard, for achieving balance

– see Fig. 233, which denes the permissible residual

unbalance of a rotating body relative to its maximum

speed. is Standard and its equivalents (e.g. ISO:

1940:1; ISO: 1290 G), assigns dierent balance-quality

grades termed: ‘G-numbers’ , related to the grouping of

rotating bodies (i.e. not shown), these groupings be-

ing based upon the experienced gained with a variety

of: sizes; speeds; and types. us, the balance-quality

grade ‘G’ , equals the specic unbalance ‘e’ times the

rotational speed ‘ω’ , as follows:

Balance-quality G = e × ω (mm sec

–1

Furthermore, the equation was described earlier, thus:

e =

(i)

∴ solving for ‘U’ , we obtain:

U =

 � M � G

rpm

(v).

From the Standard, the balance-quality for machine

tool drives is given as: G2.5, although in many in-

stances the value utilised should ideally approach that

of G1.0 – this being the specication for grinding ma-

chine tool drives, as today in HSM applications they

are compatible. However, if for the purposes of clari-

cation of the unbalance tooling condition the value

of G2.5 is utilised, then the following worked example

illustrates the balance-quality necessary using a tool-

holder weighing 3 kg, rotating at 25,000 rev min

–1

U (higher) =

 �  � .

, 

(g-mm)

∴ U (higher) = 2.85 g-mm.

As alluded to previously, this unbalance condition

is the ‘worst case’ and the tooling should ideally ap-

proach G1.0, this balance-quality value, gives:

U (lower) =

 �  � . 

, 

(g-mm)

∴ U (lower) = 1.14 g-mm.

is then follows that the balance is between 1.14 and

2.85 g-mm, which is toward the ‘upper-end’ for the

maximum residual specic unbalance for the G2.5,

while approaching this level for the G1.0 (i.e shown by

the graph in Fig. 233).

Even when the tooling assembly has been dynami-

cally balanced in both planes (i.e. see Fig. 234a – more

to be said on this topic shortly in Section 9.5.2), prob-

lems still exist, particularly in the t of the spindle ta-

per connection (Fig. 232). is is a result of the taper

rate accuracy requirements between both the shank

and taper socket. In fact, the situation is quite a con-

fusing one, due to the relative cone ‘Angle Tolerance’

grades: AT-1 to AT-6, that are employed using the con-

ventional tment of: 7:24 taper. Not only do dierent

countries oen have their own connection Standards,

but previously, even individual machine tool manu-

facturers within each country had adopted diering

Standards! Today, many machine tool companies tend

to utilise taper spindle connections that are compat-

ible to an appropriate Standard and complement those

of the tooling manufacturers.

9.5.2 HSM – Dynamic Balancing

Machine Application

It has been discussed in the previous sections that cut-

ting tool assemblies when combined with an HSM

strategy, can be a large contributor to dynamic unbal-

ance. For instance, in the production and manufacture

of say, the geometry of a face-mill, the tooling stock

material is: externally/internally turned on one side;

unclamped; ipped-over and rechecked; then turned

on the other side; then located onto a milling machine

tool for operations on the individual insert pockets

that must be milled; and indexed

– as appropriate

for the number of cutting edges; this necessary clamp-

ing/reclamping workpiece (i.e. face-mill) procedure,

will create a tool that is marginally-unbalanced. With

HSM, the otherwise unnoticeable unbalance at con-

ventional rotational speeds, becomes intolerable in

these high-speed ranges. Oen, the most economical

technique for achieving balanced tooling for tooling

26 ‘Insert pockets’ , are sometimes ‘dierentially-pitched’ which

means they have unequal spacing of teeth around the cutter’s

periphery. is pitching technique for cutting insert pockets,

is quite eective as a means of reducing machining vibrational

eects oen encountered with coarse-pitched face-mills.

472 Chapter 9