48435_08h

CHAPTER 8.3

VIBRATION CONTROL

Norman J. MasonPresident, Mason Industries, Inc.,

Hauppauge, New York

8.3.1 INTRODUCTION

Chillers, pumps, blowers, cooling towers, ducts, piping, rigid electrical conduits,etc., that are rigidly bolted to a structure, transmit 100 percent of their vibratoryenergy. The introduction of properly selected vibration isolators will reduce thistransmitted energy to the point where it is completely imperceptible or so minor asto no longer be annoying to the occupants, structurally destructive, or detrimentalto critical manufacturing processes. "Vibration isolation efficiency" is defined asthe percentage of vibration that is no longer transmitted to the structure because ofthe introduction of vibration isolators.

8.3.2 THEORY

The vibration problem can be approached on a theoretical basis by using Eq. (8.3.1),the theoretical vibration efficiency equation, which represents an isolated machineas shown in Fig. 8.3.1.

UnbalancedWeight

Isolators

Machine Mass

Rigid SupportFIGURE 8.3.1 Isolated machine.

Previous Page

Copy

right

ed M

ater

ial

Copyright 1997 by The McGraw-Hill Companies Retrieved from: www.knovel.com

E = I O ['-(wrM 1^1'

where = percentage of vibration isolated (efficiency)fd = disturbing frequency of the isolated machine/ = natural frequency of the isolated machine

The disturbing frequency should be taken as the r/min of either the equipmentor the driver, whichever is lower. All equipment has some unbalance at the primaryspeed, and this approach is conservative, since any higher-frequency vibration usedin the formula would result in overly optimistic values for the primary disturbance.

This equation states that a given percentage of the vibration can be kept out ofthe structure by installing the machine which is creating the vibration at a knownfrequency (usually the machine r/min) on a system of vibration mountings. Theseshould resonate at a frequency that is very much lower than the disturbing frequencydescribed above. When this ratio of disturbing frequency to the natural frequencyis 3:1, abut 90 percent of the transmitted vibration is theoretically eliminated. Referto Fig. 8.3.1 to help visualize the mechanical system represented by Eq. (8.3.1). Amotor is shown driving a rotating component. Assume that there is one point ofunbalance in the rotating machine and none in the motor. These two componentswould be kept in their relative positions by either a steel base or a steel base filledwith concrete. The total weight of the system would be the weight of the motorplus the weight of the machine plus the weight of the base.

Assuming that the system has only one degree of freedom and that there are noexternal connections, when we push the system down it will bounce back up andcontinue to oscillate vertically at a specific frequency. This frequency / is knownas the "natural" frequency because it continues to move at this rate with no furtherintroduction of energy.

When we turn the machine on, the unbalanced force that is centrifugally gen-erated will force the system to vibrate at the operating speed of the driven machine.This frequency is known as the "forcing" or "disturbing" frequency fd. Exami-nation of the efficiency equation (8.3.1) shows us that the larger the ratio betweenthis forcing frequency and the natural frequency, the greater will be the degree ofvibration isolation.

The only unknown in the equation is /. The equation used in determining thisfrequency is a simple relationship to the static deflection, provided the isolationmaterial that is used has a straight-line deflection curve and virtually no damping.(Mass in no way enters the equation.) The frequency is dependent upon accelerationprovided by the return spring force acting against the mass. When the deflection isdirectly proportionate to increased mass, as a function of the spring stiffness, massand stiffness conveniently drop out and deflection is all that is needed to determinefrequency. Static deflection refers to the actual difference in height between theunloaded spring and the spring in the deflected condition. For example, if the springwere 6 in (150 mm) tall to begin with and 4 in (100 mm) tall under load, the staticdeflection would be 2 in (50 mm).

If the material does not have a straight-line deflection curve, as in the case ofsome rubber materials in compression, or the material has appreciable dynamicstiffness (which means that in motion it acts like a stiffer material than the deflectionwould indicate), frequency tests can be run to determine the frequency at variousloads and deflections and the curves plotted. This is normally the case for naturalrubber and other elastomers and for materials such as fiberglass, felt, cork, andsisal. When using these materials, deflection is a very poor indicator of the natural

Copy

right

ed M

ater

ial


frequency, and test curves must be referred to for any degree of accuracy wheninserting / in the efficiency equation. The same applies to more exotic deviceslike air springs where there is no deflection under load but only a change in airpressure. Frequencies can be calculated, but test data are far more reliable.

Equation (8.3.2) refers to helical steel springs which are free of damping andhave uniform deflection rates. The natural frequency / is then expressed as

/. - ^ p

where / = natural frequency, cycles/min (Hz)dl = static deflection in the spring supports, in (mm)

In looking at this as a mechanical system, it is simpler to think in terms of thependulum. The frequency of a pendulum is dependent on the length of the pen-dulum from the rotation point to the center of the bob or mass and is in no way afunction of the mass. The frequency drops as the pendulum becomes longer andincreases as it is shortened. Pendulum length is analogous to spring deflection. Thelonger the pendulum, the lower the frequency/The more static deflection, the lowerthe natural frequency of the system.

Returning to the efficiency equation, let us try a solution referencing a machinewith a disturbing frequency of 600 cycles/min. A natural frequency of about 200cycles/min would be required to attain a ratio of 3:1 and to arrive at 87.5 percentefficiency. Since exact numbers are not critical, it is easier to use an efficiency chart(Fig. 8.3.2).

This is still a theoretical discussion and not to be used for practical purposes.This is only introductory to aid in understanding.

The operating speeds or disturbing frequencies are listed across the bottom ofFig. 8.3.2. The vertical scale is a continuous solution of the natural frequencyequation, with the deflection on the left and corresponding natural frequencies onthe right. The efficiency lines run diagonally.

To use this chart, start at the bottom with the lowest operating speed of theequipment. Move up vertically and intersect the desired efficiency line and thenmove over to the left to pick up the natural frequency of the mountings requiredfor that efficiency and the corresponding spring deflection. We have outlined thisprocedure for 90 percent at 600 r/min in the dark lines. For all practical purposes,we would need mountings with about 1 in (25.4 mm) of static deflection in orderto obtain the 90 percent efficiency.

It is important to emphasize that commercial vibration control is not an exactscience. The weight of the equipment provided by manufacturers is approximate atbest, and the location of the center of gravity is even more nebulous. Using com-mercially available mountings, it is impossible to select mountings for exact de-flection. Vibration isolation commodities are not precision devices, and there cannotbe exact cataloged springs for every load. For these reasons, deflections are gen-erally specified as minimums. Deflections beyond those that are minimal willmerely lower the natural frequency and improve performance.

Transmitted vibration is that portion of the vibration that is still sensed by thestructure. At 90 percent efficiency, there is 10 percent transmission. Efficiency isbasically a salesperson's term, probably developed because of the grading systemthat we all grew up with: 85 percent fair, 90 percent good, and 95 percent excellent.In our work, these numbers can be extremely deceptive. In comparing 95 to 90percent, we assume an improvement of only 5 percent. However, the comparison

Copy

right

ed M

ater

ial


STAT

IC DE

FLEC

TION

(6) or

NA

TURA

L FR

EQUE

NCY

(/ n)

DISTURBING FREQUENCY (/d)CYCLES PER MINUTE

FIGURE 8.3.2 Theoretical isolation efficiency chart.

of 5 to 10 percent transmission shows that only half the remaining force is trans-mitted. Neither number can be considered alone as the source of the vibration, andwhat we have to eliminate is the deciding factor.

In broad terms a 125-hp (93-kW) pump will generate five times the vibratoryenergy of a 25-hp (19-kW) pump. Therefore, the isolation provided for the 125-hp(93-kW) pump must be five times more efficient to reduce the force transmissionto a similar level. An 80 percent efficiency with 20 percent transmission for the25-hp (19-kW) pump is equivalent to 96 percent efficiency with 4 percent trans-mission for the 125-hp (93-kW) unit.

8.3.3 APPLICATION

8.3.3.1 Basic Considerations

A completely arithmetical, rather than conventional, approach provides a bettermechanical visualization of what is really going on. The results can be confirmedon the efficiency chart.

There are six basic considerations once the installation is well away from res-onance (fd/fn < 3:1):

inchesCo

pyrig

hted

Mat

erial


1. Efficiency is controlled by static deflection, and transmissions reduced in directproportion to the increase in deflection.

2. If the frequency of the isolator approaches the operating speed of the equipment,resonance is approached.

3. When approaching resonance, the dynamic motion generated by the unbalancedforce is amplified.

4. Dynamic motion is controlled by the unbalanced force and its relationship tothe total mass. For all practical purposes, at higher frequency ratios the frequencyratio itself no longer influences motion.

5. Once the frequency ratio is 3:1 or greater, motion should only be reduced bythe introduction of mass.

6. Assuming the unbalanced forces act through the center of gravity, motion isreduced in direct proportion to the increase in mass.

In reality, when a rotating machine vibrates on isolation mountings, the foun-dation has a small rotary motion. Since floors are much more sensitive in thevertical direction, the other modes are usually ignored and installations are alwaysvisualized moving vertically.

A spring constant is referred to as k and is defined as the number of pounds(kilograms) required to deflect the spring 1 in (25 mm). Thus a spring with aconstant k of 1000 Ib/in (18 kg/mm) would deflect 0.5 in (13 mm) at 500 Ib (227kg) and 1.0 in (25 mm) at 1000 Ib/in (454 kg).

A system constant is normally defined as the number of pounds (kilograms)required to deflect all the supports of a system 1 in (25 mm) simultaneously. Usingfour springs with a constant k of 1000 Ib/in (18 kg/mm) each, one in each cornerof an installation, the system constant would be 4000 Ib/in (72 kg/mm). Thus 2000Ib (907 kg) of equipment would deflect the system 0.5 in (13 mm) and 4000 Ib(1814 kg), 1.0 in (25 mm).

Since the spring rate is uniform, this also means that if the upward vibratoryforce pulled the equipment 0.10 in (2.5 mm) up from the neutral position, it wouldbe reducing the spring load by 0.10 of 4000 Ib (1814 kg), or 400 Ib (181 kg). Asthe vibratory motion pushed the installation 0.10 in (2.5 mm) below the neutralposition, there would an increase in the spring load of the same 400 Ib (181 kg).Since the bottom of the spring is attached to the structure, this change in force of400 Ib (181 kg) is what the structure sees as a change in the static loading. Thisoccurs at 600 r/min. A change in static loading at a particular frequency is anotherdefinition of vibration.

This approach to the problem is best illustrated by Fig. 8.3.3.If the machine is running at 600 r/min, the vibratory force transmitted would

be 400 Ib (181 kg) at 600 cycles/min.Assume that the 400 Ib (181 kg) is unacceptable in the structure. The instinc-

tive solution would be to use a larger mass. A traditional mass ratio is three timesthe equipment weight, bringing the system to 16,000 Ib (7258 kg) (Fig. 8.3.4).Continuing with a 1-in (25-mm) deflection spring grouping, the system constantwould have to be raised to 16,000 Ib/in (286 kg/mm) by using stiffer individualsprings or clusters of four times the original spring groupings.

In basic rule 6, it was stated that the vibratory motion would be reduced in directproportion to the increase in mass. Therefore, the amplitude of 0.10 in (2.5 mm),as in Fig. 8.3.3, would now become 0.025 in (0.64 mm). Following the sketchesacross, it is found that the reduced motion is merely acting against a proportionately

Copy

right

ed M

ater

ial


OFFDead WeightOnly600 RPMUnbalance Up 600 RPMUnbalance Down

UnbalancedWeightCentrifugal ForceDownCentrifugal ForceUp

1" Static Defl.(25.4mm)

UnloadedSpring

FIGURE 8.3.3 Vibratory transmission, 4000-lb (1800-kg) load. 1-in (25-mm) deflection.

stiffer spring constant so that there is no reduction in vibration transmission butonly in amplitude.

If the machine is running at 600 r/min, the vibratory force transmitted remains400 Ib (181 kg) at 600 cycles/min.

The question then comes up as to how to actually reduce the transmitted vibra-tion. Assume we wish to reduce this transmission by 75 percent so that the endresult is 100 Ib (45 kg). Basic rule 1 states the efficiency is controlled by staticdeflection and transmission reduced in direct proportion to the increase in deflec-tion. Rule 4 also states that for all practical purposes, the frequency ratio no longerinfluences motion.

Therefore, increase the deflection to 4 in (100 mm), as in Fig. 8.3.5. Since thislarger deflection will give us a lower natural frequency, there will be no noticeable


, UnbalancedWeight( Centrifugal ForceDown

(Centrifugal ForceUp

1" Static Defl.(25.4mm)

UnloadedSpring

FIGURE 8.3.4 Vibratory transmission, 16,000-lb (7300-kg) load. 1-in (25-mm) deflection.

Copy

right

ed M

ater

ial



UnbalancedWeightI Centrifugal ForceDown

>Centrifugal ForceUp

UnloadedSpring

FIGURE 8.3.5 Vibratory transmission, 4000-lb (1800-kg) load. 4-in (100-mm) deflection.

difference in the amplitude. The example shows that the spring constant hasdropped to 1000 Ib/in (18 kg/mm). Since the amplitude remains at the original0.10 in (2.54 mm), this amplitude multiplied by the new spring constant resultsin a force transmission of only 100 Ib (45 kg) at 600 cycles/min.

The problem can now be approached on the basis of reducing both amplitudeand transmission by reusing the total weight of 16,000 Ib (7258 kg) and providing4-in (100-mm) deflection, as shown in Fig. 8.3.6. A reduction is now made in boththe amplitude to 0.025 in (0.64 mm) and in the transmitted force to 100 Ib (45kg).

The vibratory force transmitted would be 100 Ib (45.5 kg) at 600 cycles/min.This really agrees with the efficiency chart (Fig. 8.3.2), as a 600-r/min machine

isolated by 1-in (25-mm) and 4-in (100-mm) deflection springs would show effi-


UnbalancedWeightCentrifugal ForceDownCentrifugal ForceUp

UnloadedSpring

FIGURE 8.3.6 Vibratory transmittion, 16,000-lb (7300-kg) load. 4-in (100-mm) deflection.

Copy

right

ed M

ater

ial


ciencies of 90 and 97.5 percent, respectively. Transmission reduction is 10:2.5,which is the same factor 4 shown by the arithmetic.

Both the efficiency equation and the efficiency chart are based on the completelyfalse assumption that the floor stiffness or frequency in an upper story is very highas compared to the stiffness or frequency of the isolator. In reality, the floor has adeflection of its own and a natural frequency which can be low enough to mandatethe use of isolators with very much higher deflections than indicated by the chart.Figure 8.3.7 shows the actual conditions in a structure. Rather than a simple systemwith the machine or machine foundation resting on springs on a relatively unyield-ing support, the springs are supported by a spring board with a finite mass of itsown. Schematically, this is sketched in Fig. 8.3.8. The machine mass rests onsprings on the floor mass, and the floor stiffness is shown by a second set of springs.Although floors are supported by beams connected to vertical columns, ground-level vertical stiffness really exists only at the columns and not in between.

UnbalancedWeight

Machine Mass

Isolators

Floor Deflection"*" Column Supports

Assumed to be RigidFIGURE 8.3.7 Actual structure conditions (floor de-flection exaggerated).

UnbalancedWeight

Machine Mass

Isolators

Floor MassSchematic ofFloor Deflection

Schematic of Rigid ColumnsFIGURE 8.3.8 Schematic of floor deflection.

Copy

right

ed M

ater

ial


The worldwide structural limit on floor deflection is 1/360th of the span. Inmany commercial buildings design spans are at least 360 in (9144 mm), or 30 ft(9.14 m). This means that the structural engineer is allowed a floor deflection of 1in (25 mm) at the center of the span when the floor is fully laden with both liveand dead loads.

Let us make the assumption that in a particular area where a 125-hp (93 kW)pump is installed that the floor is loaded to half dead load plus live load. The floordeflection would then be 0.5 in (13 mm). If the pump is running at 1750 r/min, aquick reference to the efficiency chart shows that a mounting deflection of 0.1 in(2.54 mm) should provide 90 percent efficiency or 10 percent transmission. How-ever, the 0.5-in (13 mm) floor deflection would be five times the deflection of theisolator. The actual efficiency would be influenced by the floor's 0.5-in (13-mm)deflection, the mass of the floor, and the floor's damping characteristics. The 90percent theoretical efficiency could never be attained, and depending on the com-bination of conditions, the actual efficiency might slide down to 50 or 40 percentand not meet the requirement.

Therefore, rather than relying on the theoretical method, commercial selectionof isolators has evolved into using isolators with deflections that equal or exceedthe floor deflection to attain acceptable transmission levels. The efficiency chartshould only be used as a tool to learn the subject and gain direction.

Studies show that the floor stiffness is greatly in excess of the isolator stiffness,because the mass of the floor is much greater than the mass of the machine that isto be isolated. While this may be of importance when isolating small equipment,it is certainly not significant with large pumps and chillers.

On a day-to-day basis, the cost of an involved engineering investigation of acommercial structure's stiffness and resonance, along with the possibility of errorin these conclusions, dictates the continued use of the more conservative floor-deflection rather than stiffness approach. The cost of isolation is small as comparedto the cost of an installed air-conditioning system. Possible savings in using lower-deflection materials are in no way proportionate to complete loss of occupancy orlower rental rates in a noisy structure or recourse to very expensive retrofits.

The recommended deflections shown in the selection guide (Table 8.3.6) arebased on empirical data gathered through 40 years of installation experience, aswell as discussions with mechanical and acoustical engineers, architects, and man-ufacturers. The deflections were influenced by operating speeds, size of equipment,the equipment as a vibratory source, and the sensitivity of the floor structure interms of construction and floor span.

8.3.3.2 Isolation MaterialsAn "isolation material" can be defined broadly as any resilient material that willaccept a load on a permanent basis and produce a resonant or natural frequencythat is reasonably consistent and predictable. It is also important that any increasein this frequency is small as the material ages.

Vibration Pads. "Elastomeric" describes any rubberlike material. While naturalrubber has the best performance characteristics, it is generally not used commer-cially because of aging when exposed to oxygen, ozone, or oil. A synthetic elas-tomer similar in properties to natural rubber but lower in cost is the oil derivativeSBR (styrene-butadiene rubber). It is very commonly used where there are no spe-cific aging requirements. The neoprenes are not quite as resilient as natural rubberor SBR, but because of their very excellent aging characteristics, better grades of

Copy

right

ed M

ater

ial


pad are either all neoprene or neoprene blended with SBR or natural rubber toreduce cost.

While the selection of the proper elastomer is an important choice in terms ofperformance and aging, the physical properties of the compound are specificallycontrolled by formulation. The polymers are not used by themselves but mixed withother materials such as carbon black and clay to provide reinforcement. Largerratios of these fillers will reduce cost but with penalty to physical properties. As ageneral statement, a ratio of one-third fillers to the selected polymer produces thebest physical and dynamic properties. When aging is the primary concern, theformulation should contain only neoprene and not a blend of the neoprene withnatural rubber or SBR. Unless a specification states exactly what the physical prop-erties and ingredients must be, the material that is furnished will only follow themanufacturer's conscience. Unfortunately, there are no industry standards as towhen a pad can be called neoprene or natural rubber. Products are available thatare made with as little as 5 percent neoprene but still referred to as commercial-grade neoprene. A good guide to quality are the AASHO standards shown in Table8.3.1.

Other than foams, which do not have enough capacity or stability to be used asisolators, air-free rubber materials are incompressible. When a load is applied to apad, it changes shape but does not lose volume. The ability to change shape iscontrolled by the shape factor and the material's hardness. Since a pad used incompression can only change its volume by bulging, unconfined edges are referredto as "escape area," whether internal in the form of holes or external. The term"shape factor" is the ratio of the loaded area to the escape area. The lower theshape factor, the more deflection at a particular load.

Thus a 4-in (100-mm) square pad that is 1 in (25 mm) thick, covered completelyeither by the equipment or by a steel plate, would have a loaded area of 16 in2(100 cm2). Since the perimeter is 16 in (400 mm) and the pad 1 in (25 mm) thick,the escape area is also 16 in2 (100 cm2). The shape factor (load area divided byescape area) would be 16/16 (100/100), or 1. Assuming that the hardness remainsthe same, we could increase the load-carrying capacity of this pad by using twol/2-in (12.5-mm) thick pads with a steel plate separating the two layers. Since theloaded area would remain 16 in2 but the escape would not be 16 in (400 cm)multiplied by 0.5 in (12.5 mm), or 8 in2 (50 cm2), the shape factor would beincreased to 2 with a lower deflection for the same imposed load. These relation-ships are shown in shape-factor curves (Fig. 8.3.9). These curves are empiricalbased on test data, and small variations will be found from one publication toanother.

Hardness is measured by a durometer (in units called "duros"), a clocklikegauge with a penetration probe on the bottom. Pads are normally used in 30 to 70duro in 10-point increments. Since rubberlike materials cannot be exactly con-trolled, the normal acceptable variation is 5 of a nominal duro. To give you someidea of the feel of these durometers, common references are rubber band stock atabout 30, red erasers at about 40, white erasers at 50 to 60, and the old-fashionedhard gray erasers at 70.

Automobile tires are 50 to 70 duro. The influence of hardness on load capacityis shown in Fig. 8.3.10. A 70-duro material will handle about four times the loadthat would be carried by the same shape in 30 duro.

Unfortunately, the harder the rubber material, the less it acts like a steel spring;the introduction of viscosity is similar to the introduction of a dashpot working inparallel with a spring. When the dashpot becomes large and the fluid stiff, vibratoryforces are transmitted through the dashpot. It is for this reason that 70 duro isC

opyr

ighte

d M

ater

ial


TABLE 8.3.1 Physical Properties of Structural Bearings Made from Du Pont Neoprene

Performance RequirementsHARDNESS GRADE

706050Test MethodPhysical Property

psi (kPa)minimum% maximum

lbs. per inch,minimum

psi(kPa)

points, maxi-mum

%, maximum%, maximum%, maximum*

% maximum

7052500

(1725)30040

110(76)

+ 15

-40-15+80

No Cracks

35

6052500

(1725)30040

75(52)

+ 15

-40-15+ 80

No Cracks

35

5052500

(1725)40040

50(35)

+ 15

-40-15+ 80

No Cracks

35

ASTM D 2240ASTM D 412

ASTM D 412ASTM D 429

(Sample first prepared 96 hr. at-20 20F (-29 I0C) axialload 500 psi and strain of 20%"T" [effective thickness].)Shear resistance after 1 hr. at25% shear strain not to exceedvalues shown

ASTM D 573

ASTM D 471*

ASTM D 1149

ASTM D 395

Hardness, durometer ATensile strength

Elongation at breakAdhesion

Bond made during vulcanizationLow-temperature performance

Resistance to heatChange in original properties after 70 hrs. at

2120F (10O0C)Hardness

ElongationTensile Strength

Resistance to oil aging*Change in volume after 70 hrs. immersion inASTM Oil No. 3 at 2120F (10O0C)

Resistance to ozoneCondition after exposure to 100 pphmozone in air for 100 hrs. at 100 20F(29 I0C) (sample under 20% strain)

Resistance to permanent setCompression set after 22 hrs. at 2120F (10O0C)

This oil aging requirement is not a part of the AASHO Specification referenced. However, its inclusion is strongly recommended to assure use of a high-quality neoprenecompound.

Source: American Association of State Highway Officials Standard Specification for Highway bridges, Table B.

Copy

right

ed M

ater

ial


COMP

RESS

ION

STRE

SS P.S

.I. (kP

a)

BULGERATIO

PERCENT COMPRESSION - 40 DUROMETERFIGURE 8.3.9 Shape-factor or bulge ratio.

considered the extreme hardness for vibration isolation. Since 30 duro becomesuneconomical for large loadings and hard to manufacture, most pad materials fallinto the 40- to 60-duro range.

Figure 8.3.11 is a dynamic stiffness chart based on experimental work withneoprene compounds containing no other elastomer and minimum fillers and theactual frequency at various deflections and hardness. Increased use of fillers lowerscost and quality at the expense of performance. To see the influence of the dynamicstiffness, you need only compare these frequencies and deflections with the samedeflections but lower frequencies shown in the efficiency chart, Fig. 8.3.2.

Pad deflection is limited by thickness. For the material to remain resilient andto control permanent set and creep, pad deflection should be limited to 15 percentof the thickness regardless of the rubber configuration or the rubber material.C

opyr

ighte

d M

ater

ial


RELA

TIVEL

OADC

APAC

ITYDY

NAMIC

NA

TURA

L FRE

QUENC

Y - Her

tzDat

a Base

d on

Resona

nce Tes

ts

DYNA

MIC NA

TURA

L FRE

QUENC

Y - CPM

PERCENT DEFORMATIONFIGURE 8.3.10 Influence of hardness on load capacity.

DYNAMIC FREQUENCY CURVE

Test No. Kal-1391-1-72Dynamic StiffnessNeopreneDurometgr Compound Stiffness

70 7072 2.3260 6072 1.6350 5072 1.5040 4072 1.4330 3072 1.25

Data Based on Numbered NeopreneCompounds for EAFM Mounts

STATIC DEFLECTION - inches mmNote: Data would not apply to compounds other than those tested.

FIGURE 8.3.11 Dynamic stiffness chart.Cop

yrigh

ted

Mat

erial


Thus, the maximum deflection for a 1-in-thick (25-mm) pad is 0.15 in (3.8 mm)with corresponding reductions in deflection for thinner materials.

A vibration pad may be solid if for a given load it has the proper shape factorfor the maximum 15 percent deflection in an acceptable durometer. Deflection canonly be increased by increasing thickness. Pads may be molded thicker or made ofmultiple layers separated by steel shim plates.

In most cases, loadings per unit area are low, so rather than solid pads, additionalescape area is needed to reduce the shape factor. Most vibration pads are moldedwith round or square holes and in cross-ribbed and waffle designs, as shown inFig. 8.3.12.

Most commercial isolation pads are available in thicknesses up to 3/s in (9.5mm). They should be used at a maximum deflection of 0.06 in (1.5 mm) per layer.Dynamic frequency is 16 Hz in the best materials. A new %-in (19-mm) pad hasbeen introduced recently with deflections of 0.11 in (2.8 mm) and a dynamic fre-quency of 12 Hz.

FIGURE 8.3.12 Typical vibration pads.

Copy

right

ed M

ater

ial


Other pad materials are cork, felt, sisal, and heavy-density precompressed fiber-glass. Since the fiberglass is a fragile, spongelike material, it is normally coveredwith a neoprene or other coating to protect it against fraying and moisture. Fiber-glass-pad frequency is not as sensitive to deflection and loading, so fiberglass padsare sometimes described as flat-frequency materials. However, this frequency ishigher than neoprene or natural rubber at the same deflection. Table 8.3.2 showspublished frequencies for 0.5-in (12.5-mm) and 1.0-in (25-mm) pads. Fiberglasspads should be avoided if there are large shear forces, such as those under a hor-izontal compressor.

Rubber mountings are sophisticated rubber pads. While the rubber may beloaded in shear, most commercial rubber mountings for air-conditioning applica-tions are loaded in compression. Quality levels may be as described for pads, butthe static deflections are higher only because the mountings are thicker. Rubbermountings have the advantage of provision for bolting to the equipment and to thefloor when needed. Many of the newer designs have rubber under and over thebase and top plates so that they can be slipped into place without bolting in sta-tionary applications.

Springs. Spring mountings are generally required to provide the minimum de-flection needed to compensate for structural flexibility. The heart of any steel springmounting is the spring itself. It should be designed with a minimum diameter-to-deflected-height ratio of 0.8 so that the horizontal spring constant Kh is a minimumof 80 percent of the vertical Kv. Most designs end up with those 0.8 ratio, but thisis not an exact rule. While this chapter is not meant to be a spring design handbook,Fig. 8.3.13 will give you the criteria for checking the horizontal stiffness as com-pared to the vertical.

An allowance of 50 percent additional capacity beyond rated load is also goodpractice. This means a spring rated for 2-in (50-mm) deflection would not have thecoils touching (solid) before 3 in (75 mm). If it is rated for 1-in (25-mm) deflection,it should not go solid before IVz in (37.5 mm). Overtravel allowance is needed asit is impossible to calculate exact weight distribution. Published equipment weightis inaccurate, and center-of-gravity locations are often unavailable. A 50 percentovertravel will allow for an acceptable 20 percent overload.

Isolation springs should be designed such that when the coils are touching, theelastic limit has not been exceeded so that springs will return to full height. If thespring is designed this way, it will be stressed two-thirds the elastic limit undernormal load. Springs last indefinitely as vibration amplitudes are very small, andthe spring movement per coil is the total amplitude divided by the number of activecoils. With so little movement the stress cycle is close to zero, and these applica-tions approach static loadings.

A simple spring can be considered a vibration isolator and often built into me-chanical equipment. This low-cost method is satisfactory when thousands of springsare used in a repetitive application. When springs are used in the field, however,minimum additions to the design are normally a neoprene friction pad on the bottomto eliminate the need for bolting and to act in series with the steel coils to helpeliminate high-frequency noise transmission. Since loadings are not easily or ex-actly determined, there must be a means of leveling, and this is usually done bymeans of an adjustment bolt. The bolt is often used to attach the mountings to theequipment as well.

Air Springs. The last remaining commodity of major importance is the air spring.Air springs are made of neoprene with nylon tire cord reinforcement and shapedlike vertical bellows. Ethylene-propylene (EPDM) is also used for this purpose and

Copy

right

ed M

ater

ial


TABLE 8.3.2 Comparison of Natural Frequencies at Given Deflections: Heavy-Density Fiberglass, AASHO Neoprene, Steel Springs

Steel springsAASHO neopreneHeavy-density fiberglass

Freq.,Hz

Deflectionmmin

Dyna-mic

freq.,Hz

15%Deflection

Padthickness

mminmmin

Dyna-mic

freq.,Hz

DeflectionPad

thicknessmminmmin

14.811.410.48.17.4

1.11.92.33.84.6

0.0450.0750.0900.1500.180

18.014.513.110.59.5

1.11.92.33.84.6

0.0450.0750.0900.1500.180

7.612.715.225.430.5

0.30.50.61.01.2

22

16

1.1 to2.3

2.3 to4.6

0.045to

0.0900.09 to

0.18

13

25

0.5

1.0

Copy

right

ed M

ater

ial


8st STATIC DEFLECTIONhs S COMPRESSED HEIGHT

LATE

RAL S

TIFFN

ESS .

KAIIO

AX

IALST

IFFNE

SS

RATIO COMPRESSED HEIGHT. hsMEAN COIL DIAMETER " DFIGURE 8.3.13 Horizontal versus vertical stiffness.

SPRING STABLE IF

sometimes butyl as the inner liner. Butyl is the least permeable of the rubber ma-terials and reduces air loss.

The rolling lobe is a design variation that provides for movement by the rubberwalls literally rolling down a steel stanchion rather than flexing. Both designs areequally suitable, and it is merely a matter of selecting one over the other dependingon what frequency is needed. In general, rolling lobes have lower natural frequen-cies as compared to single-convolution and double-convolution bellows of the sameheight.

All air springs leak. The leakage rate is very low, but it is generally impracticalto set up supervisory replenishing procedures. All air spring systems should beinstalled with replenishing air lines connected to height-sensitive leveling valves.If an air spring or cluster of air springs loses air and the equipment settles, air willautomatically be added. Where air springs might be installed on a hot roof, theleveling system would respond and bleed small quantities of air should the airexpand.

Leveling valves also level equipment that goes out of level because of externalforces when the equipment is running. For example, a top horizontal fan tends torotate away from the point of discharge. It rears up on the discharge end and settlesin the back. Leveling valves automatically compensate for this and return the in-stallation to proper elevation.

Air springs have the advantages of low frequencies and low profiles. Since thereis no steel continuity, there is no noise transmission. Air spring frequency variesvery little with pressure, but since the capacity is directly proportional to increasedpressure, air springs need not be selected as carefully as steel springs. Most deviceswill handle loadings at a minimum of 25 Ib/in2 (172 kPa) and as much as 100 Ib/in2 (689 Pa), which allows for a 4:1 ratio from minimum to maximum loading ona particular mounting.

Hangers. Hangers accommodate all the above devices in modified form so thatthey fit within steel frames which are usually open-sided. Very simple high-frequency neoprene hangers could be pad hangers, but more often rubber elementsare designed for hangers by eliminating the base plates and the tapped holes ontop. The elements are molded with a projecting bushing that passes through thehanger hole to prevent the rod from rubbing. Occasionally fiberglass pads are usedin lieu of neoprene elements, but they generally have even higher frequencies. Steelsprings are generally fitted into neoprene cups, and the cup itself has a no-rub

Copy

right

ed M

ater

ial


bushing arrangement. Rubber elements are often used in series with steel springsto combine the advantages of both materials.

It is very important to isolate and provide for flexibility in the piping so that thefunction of the floor mountings or equipment hangers connected to piped equipmentis not interfered with or bypassed.

Connectors. Stainless steel connectors can be manufactured to specific lengths;they consist of a corrugated stainless steel body welded to the appropriate fittings.Even under moderate pressures, the closely spaced bellows become unstable andwould spew out as the bellows expand. Therefore, virtually all metallic connectorshave a stainless steel braid attached to the two ends to form a tube over the stainlesssteel corrugations. Stainless steel braid prevents elongation and adds to the radialstrength.

Metallic connections are designed to allow flexing and reduce fatigue. The con-nector end away from the equipment should be rigidly secured so that the connectoris forced to flex. This minimizes piping vibration after the anchor. Unfortunately,when any such connector is pressurized, there is a tremendous pull on the braidwhich makes the assembly extremely stiff. As a practical matter, it is very difficultto secure the afterend rigidly, so that in the average installation the flexible con-nector compensates for misalignment but does very little to reduce noise and vi-bration.

The next class of flexible connectors are Teflon bellows which are an improve-ment on the metallic connectors because the Teflon introduces a discontinuity inthe metallic pipe wall. While Teflon is an excellent material, temperature andpressure ratings are often too low for high-rise structures. Teflon bellows are man-ufactured with built-in control rods to prevent excessive elongation. Control rodsseverely limit vibration and noise reduction as they bypass the flexible bellows.

Hand-built, single-arch rubber expansion joints are still being manufactured andare similar in function to Teflon. The arch in the center is all that provides theflexibility, and because of the bellowslike shape these connectors also elongateunless control rods are used or the piping is anchored. Here again the control rodstend to bypass the action of the expansion joint. The stiff walls leading up to thearch have little function other than to provide room for the steel retaining rings andthe bolt heads that go between the steel rings and the arch. Applications are gen-erally reserved to industry as they can be built up to 144 in (3.7 m) in diameterand manufactured in exotic rubber materials, particularly for high temperatures andhighly corrosive chemicals.

The most recent entrant is the spherical neoprene or ethylene-propylene (EPDM)expansion joint. Unlike the three commodities described above, spherical connect-ors are designed on the principle of the automobile tire. The reinforcement fabricis nylon tire cord. The tire cord forms a suspension bridge from one flange to theother. When these connectors are pressurized, the nylon stretches until the stretchingforce equals the pressure and then the connector remains stable at that length anddiameter. In most cases they can be installed without control rods.

The volumetric expansion and contraction of the connector dissipates soundenergy. They do an excellent job in reducing sound transmission at blade passagefrequency (number of pump buckets times shaft r/min). Unfortunately, they are toostiff to handle the primary vibration at r/min, so it is still necessary to protect thestructure by isolating the pipelines with isolation hangers. The connectors take careof misalignment and virtually eliminate the high-pitched whine that normally travelsthrough a structure. Connectors are recommended in a double-sphere configuration,for equal ends or concentric reduction. Long-radius tapered elbows save space atpump connections.

Copy

right

ed M

ater

ial


Bases. When equipment is manufactured or supplied in multiple components notconnected by a common base, a base must be used to connect the two elementsbefore the vibration isolators can be applied under the entirety. A belt pulling ona flexibly mounted motor connected to a flexibly mounted blower would pull oneor the other off the isolators. Torque causes similar problems when machinery isdirectly connected. Bases can be constructed of structural steel, or, where additionalmass is required or advisable, of steel frames filled with concrete. Most air-conditioning equipment is so well-balanced that there is no base weight criteria andrigidity is the only concern. Descriptions of these bases are covered in the selectionsection.

8.3.4 SELECTION

Following is a complete guide specification written in engineering terms. The spec-ification selection guide provides the proper prescription for the complete isolationof a unit in terms of the type of mounting or hanger, the recommended deflection,the need for a base (if there is one), and the recommendation for a specific flexibleconnector.

We suggest you include all these specification paragraphs in your standard en-gineering specifications. In addition, prepare a drawing of standard details thatbecomes part of the mechanical drawing set. This eliminates constant editing oneach job since the materials are not actually used unless an extra column is addedto your equipment schedule, Table 8.3.3. It is this callout that defines the isolationthat is to be used under each piece of equipment. Table 8.3.3 refers to the appro-priate specification paragraph by letter with the notation as to the proper staticdeflection. The recommendations come right from the selection guide (Table 8.3.6),so the table containing them can be prepared very quickly.

The selections were based on a 30-ft (9.1-mm) floor span in the penthouse (9.1m) and a 20-foot (6.1-m) span in other locations. Note that for pump no. 5 noisolation is called out as it is located in the basement under the garage where noone could be annoyed by the vibration. Fire pumps (no. 6) are seldom isolated.

In preparing a schedule similar to Table 8.3.3 there is an opportunity to considerevery piece of equipment and very little possibility of skipping over some item inthe rush of completing the project.

SPECIFICATION ADouble-deflection neoprene mountings (Fig. 8.3.14) shall have a minimum static de-flection of 0.35 in (8.9 mm). All metal surfaces shall be neoprene-covered to avoidcorrosion and have friction pads both top and bottom so they need not be bolted to thefloor. Bolt holes shall be provided for these areas where bolting is required. Steel rails(Fig. 8.3.15) shall be used above the mountings to compensate for the overhang onsmall vent sets close-coupled pumps, etc.

SPECIFICATION BSpring-type isolators (Fig. 8.3.16) shall be freestanding and laterally stable without anyhousing and complete with /4-in (6.4-mm) neoprene acoustical friction pads betweenthe baseplate and the support. All mountings shall have leveling bolts that must berigidly bolted to the equipment. Spring diameters shall be no less than 0.8 of thecompressed height of the spring at rated load. Springs shall have a minimum additionaltravel to solid equal to 50 percent of the rated deflection. Submittals shall include springdiameters, deflections, compressed spring height, and solid spring height.

Next Page

Copy

right

ed M

ater

ial


Front MatterTable of ContentsPart A. System ConsiderationsPart B. Systems and ComponentsPart C. General Considerations8.1 Automatic Temperature, Pressure, Flow Control Systems8.1.1 Control Basics8.1.1.1 Control Systems8.1.1.2 Modes of Feedback Control8.1.1.3 Flow-Control Characteristics8.1.2 Control Equipment Types8.1.2.1 Sensors8.1.2.2 Controllers8.1.2.3 Final-Control Elements8.1.2.4 Auxiliary Equipment8.1.2.5 Pneumatic, Electric, Electronic Comparisons8.1.3 Control Applications8.1.3.1 Boiler Control8.1.3.2 Control of Excess Air8.1.3.3 HVAC Fan Systems8.1.3.4 Refrigeration Control8.1.3.5 Central Heating and Cooling Plants8.1.3.6 Water-Distribution Control8.1.4 Building Management Systems8.1.4.1 Building Management System Types8.1.4.2 Management System Applications8.1.5 Selection8.1.6 Total Building Function8.1.6.1 Type of Building and System Zoning8.1.6.2 Types of Occupancy and Use8.1.6.3 Accuracy Requirements8.1.6.4 Economic Justification

8.2 Noise Control8.2.1 Introduction8.2.2 The Nature of Sound8.2.2.1 Displacement Amplitude and Particle Velocity8.2.2.2 Frequency8.2.2.3 Wavelength8.2.2.4 Sound Level8.2.3 The Speed of Sound in Air8.2.4 The Speed of Sound in Solids8.2.5 The Decibel8.2.5.1 Sound Power Level8.2.5.2 Sound Pressure Level8.2.6 Determination of Sound Power Levels8.2.7 Calculating Changes in Sound Power and Sound Pressure Levels8.2.7.1 Sound Power Level8.2.7.2 Sound Pressure Level8.2.8 Propagation of Sound Outdoors8.2.9 The Inverse-Square Law8.2.10 Partial Barriers8.2.11 Propagation of Sound Indoors8.2.11.1 Direct Sound Path8.2.11.2 Reverberant Sound Path8.2.11.3 Effects of Direct and Reverberant Sound8.2.12 Sound Transmission Loss8.2.12.1 The Mass Law8.2.12.1 The Effect of Openings on Partition TL8.2.12.3 Single-Number TL Ratings: STC Ratings8.2.13 Noise Reduction and Insertion Loss8.2.14 The Effects of Sound Absorption on Receiving-Room NR Characteristics8.2.15 Fan Noise8.2.76 Cooling Tower Noise8.2.17 Duct Silencers-Terminology and Types8.2.18 Effects of Forward and Reverse Flow on Silencer SN and DIL8.2.18.1 Brief Theory of the Effects of Air-Flow Direction on Silencer Performance8.2.19 Combining Active and Dissipative Silencers8.2.20 Sound Transmission Through Duct Walls-Duct Break-out and Break-in Noise8.2.21 Noise Criteria8.2.21.1 dBA Criteria8.2.21.2 Community and Workplace Noise Regulations8.2.21.3 Noise Criteria (NC) Curves8.2.21.4 Speech Interference Levels8.2.21.5 Ambient Noise Levels as Criteria8.2.22 Enclosure and Noise Partition Design Considerations8.2.22.1 Actual Versus Predicted Sound Transmission Losses 8.2.598.2.22.2 Joints8.2.22.3 Windows and Seals8.2.22.4 Doors and Seals8.2.22.5 Transmission Loss of Composite Structures8.2.22.6 Flanking Paths8.2.22.7 Room Performance8.2.23 Sound Absorption in Rooms8.2.24 Silencer Application8.2.24.1 Specific Effects of Flow Velocity on Silencer Attenuation8.2.24.2 Interaction of DIL with Self-Noise8.2.24.3 Pressure Drop8.2.24.4 Energy Consumption8.2.24.5 Effects of Silencer Length and Cross Section8.2.24.6 Impact on Silencer p of Proximity to Other Elements in an HVAC Duct System8.2.24.7 Duct Rumble and Silencer Location8.2.24.8 Effect of Silencer Location on Residual Noise Levels8.2.25 Systemic Noise Analysis Procedure for Ducted Systems8.2.25.1 Procedure8.2.25.2 Silencer Selection8.2.25.3 Calculating the Attenuation Effects of Lined Ducts8.2.26 Acoustic Louvers8.2.27 HVAC Silencing Applications8.2.28 Self-Noise of Room Terminal Units8.2.29 The Use of Individual Air-Handling Units in High-Rise Buildings8.2.30 Built-Up Acoustic Plenums8.2.31 Fiberglass and Noise Control-Is It Safe?8.2.32 References

8.3 Vibration Control8.3.1 Introduction8.3.2 Theory8.3.3 Application8.3.3.1 Basic Considerations8.3.3.2 Isolation Materials8.3.4 Selection8.3.5 Seismic Protection of Resiliently Mounted Equipment8.3.5.1 Theory8.3.5.2 Seismic Specifications8.3.6 Acoustical Isolation by Means of Vibration-Isolated Floating Floors8.3.6.1 Theory and Methods8.3.6.2 Specification

8.4 Energy Conservation Practice8.4.1. Introduction8.4.2 General8.4.3 Design Parameters8.4.3.1 Energy Audit8.4.3.2 Design8.4.3.3 Types of Systems8.4.3.4 Chillers8.4.3.5 Boilers8.4.3.6 Waste Heat and Heat Recovery8.4.3.7 Automatic Temperature Controls (See Also Chapter 8.1)8.4.4 Life-Cycle Costing8.4.4.1 General8.4.4.2 Discounting, Taxes, and Inflation8.4.4.3 Related Methods of Evaluation8.4.5 Energy Management Systems8.4.5.1 Components8.4.5.2 Software Programs8.4.5.3 Functions8.4.5.4 Optional Security and Fire Alarm System8.4.5.5 Selecting an EMS8.4.6 References

8.5 Water Conditioning8.5.1 Introduction8.5.2 Why Water Treatment?8.5.2.1 Cost of Corrosion8.5.2.2 Cost of Scale and Deposits8.5.3 Water Chemistry8.5.3.1 Hydrologic Cycle8.5.3.2 Water Impurities8.5.3.3 Dissolved Gases8.5.3.4 Dissolved Minerals8.5.4 Corrosion8.5.4.1 General Corrosion8.5.4.2 Oxygen Pitting8.5.4.3 Galvanic Corrosion8.5.4.4 Concentration Cell Corrosion8.5.4.5 Stress Corrosion8.5.4.6 Erosion-Corrosion8.5.4.7 Condensate Grooving8.5.4.8 Microbiologically Influenced Corrosion (MIC)8.5.5 Scale and Sludge Deposits8.5.5.1 Mineral Scale and Pipe Scale8.5.5.2 Langelier Index8.5.5.3 Ryznar Index8.5.5.4 Boiler Scale8.5.5.5 Condensate Scale8.5.6 Foulants8.5.6.1 Mud, Dirt, and Clay8.5.6.2 Black Mud and Mill Scale8.5.6.3 Boiler Foulants8.5.6.4 Construction Debris8.5.6.5 Organic Growths8.5.6.6 Algae8.5.6.7 Fungi8.5.6.8 Bacteria8.5.7 Pretreatment Equipment8.5.7.1 Water Softeners8.5.7.2 Dealkalizer8.5.7.3 Deaerators8.5.7.4 Abrasive Separators8.5.7.5 Strainers and Filters8.5.7.6 Free Cooling8.5.7.7 Gadgets8.5.8 Treatment of Systems8.5.8.1 General8.5.8.2 Boiler Water Systems8.5.8.3 Treatment for Open Recirculating Water Systems8.5.8.4 Treatment of Closed Recirculating Water Systems8.5.9 References8.5.10 Bibliography

Appendices

Index

48435_08h

Documents

transmitted vibration

vibration controlnorman

system of vibration

specific frequency

asthe percentage of

vibration isolation

disturbing frequency

ratio of disturbing