CHAPTER 2.3

Vacuum Blowers Jay Richman

Consultant, Stokes Vacuum Inc.

2.X1

INTRODUCTION

Roots-type blower systems comprised of blowers and backing pumps are almost universally used in vacuum systems where large gas loads are pumped in the in­termediate vacuum range and/or volumes must be evacuated in a relatively short time. They are used as primary pumps at intermediate vacuum levels, for rough pumping to the vacuum level where high-vacuum pumps can be effectively em­ployed, and, where required, as backing for high-vacuum pumps. In some sys­tems, the blower system provides all three functions. This section is intended to provide the user with an understanding of the pumping principles, operating char­acteristics, and limitations of blower systems to assist in proper application and problem diagnostics.

2.3.2

EQUIPMENT DESCRIPTION

Roots-type blowers (mechanical boosters) are positive displacement machines that employ two identical counterrotating impellers, configured to run with close clearances between the impellers and between the impellers and the enclosing casing. The impellers may have two or three lobes with cycloidal, involute, or

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97

98 Chapter 2.3: Vacuum Blowers

Fig. I. TIMING

TYPICAL VACUUM BLOWER

CONSTRUCTION

Vacuum seals may be lip or face type. Labyrinth and shaft seals can be replaced by four face seals to isolate the pumping chamber from the bearings and gears.

compound curve flank geometry. Figure 1 illustrates the typical construction of a two-lobe blower. The impellers are mounted on bearings, are well balanced, and normally run at rotational speeds in the range of 1800 to 3600 rpm. Lip or face type seals are used for the atmospheric shaft seals. For critical applications, blow­ers are available with hermetically sealed direct drive motors to replace the shaft seals. In one version of these motors, the rotor is mounted directly on the drive shaft and isolated from the stator by a hermetic enclosure (Figure 2).

Pumping action is obtained by the impellers trapping pockets of gas at inlet pressure between the impellers and casing, and transferring the gas to a higher-pressure discharge port (Figure 3). The blowers are valveless, run without sealing fluids (except in special cases), and close clearances are required to maintain pumping efficiency by minimizing backflow. Clearance between impellers is main­tained by minimum backlash timing gears, and clearances between impeller ends and casing end plates are controlled by accurate positioning of the shafts in the bearings.

2.3.2 Equipment Description 99

Fig. 2. Sealess blower drive.

STATOR

Rotor isolated from the stator and atmosphere by hermetically sealed, thin-walled shell. Blower has a bypass valve to facilitate starting from atmosphere. Courtesy: Stokes Vacuum Inc.

Fig. 3.

SUCTION GAS

Pi PRESS

T i TEMP s-i M V H R

Blower pumping principle.

DISCHARGE

P2 T? S2

GAS

PRESS TEMP

M V H R

Two lobe impellers discharge four trapped volumes (per rotation of drive shaft) to higher-pressure region.

100

Fig. 4.

Chapter 2.3: Vacuum Blowers

Typical blower-backing pump performance.

1200

800

600

400

200

~ —r -^~r

CUT IN PRESSURE

I FOREPUMP #1

Pumping speed curves for the same blower backed by two different-capacity backing pumps.

Blower systems consist of one or more blowers in series with dry-running or liquid-sealed mechanical vacuum pumps that discharge to atmosphere. The blow­ers provide high pumping capacities at low pressure and extend the operating range of the backing pumps by a decade or more. Staging ratios (relative dis­placements of blower and backing pump) range from 2 to 10 or more, depending on operating pressure. The blowers also provide vapor compression that reduces the amount of vapor backstreaming from liquid-sealed backing pumps.

Blank-off pressures of two-stage blower-mechanical pump systems range from 1 X 10 "" to 1 X 10"^ torr depending to a large extent on backing pump blank-off. Capacities range from 170 m^/hr to 30,000 m^/hr (and higher). Figure 4 il­lustrates typical pumping speed relationship to pressure level of two systems that employ the same blower with different capacity backing pumps.

2.3.3

BLOWER OPERATING PRINCIPLE

As noted, pumping action is obtained by the transfer of low-pressure pockets of gas to a higher-pressure regime. At the instant that the pocket is exposed to the blower exit port, high-pressure gas diffuses into the displacement volume to equal-

2.3.4 Blower Pumping Efficiency 101

ize pressure. Recompression is complete when the impeller has rotated 90 de­grees and interstage volume between the blower and backing pump has been re­duced to its original size. This action takes place four times per revolution of the drive shaft when two lobe impellers are employed. Volumetric displacement (max­imum theoretical pumping speed), m^/hr, is obtained by multiplying the four vol­umes displaced by the rotational speed per hour (rph).

2.X4

BLOWER PUMPING EFFICIENCY

As noted, the blower has no valves and depends on close clearances to minimize backflow from discharge to inlet pressure. The amount of gas that flows back through the clearances reduces the net effective pumping speed. Efficiency is measured by the ratio of the actual volume of gas pumped to the gross displace­ment or theoretical maximum pumping speed.

E = ^ (1)

£ = 1 - ^ (2)

where E = blower pumping efficiency SD = blower displacement, mVhr

SfB ^ total backflow, m Vhr S^ = net pumping speed, m Vhr P2 = interstage pressure, ton-Pi = blower suction pressure, ton-

Total backflow, 5FB, (mVhr) is comprised of flow through clearances, canyback of gas on impeller surfaces, and reexpansion of minor volumes trapped between impellers. Efficiency is influenced by the following factors.

2.3.4.1 Pressure Level and Compression Ratio

Impeller clearances vary with blower size but typically are on the order of 0.25 mm. At high pressures, with a compression ratio of 2 (or higher), backflow is sonic and turbulent and is a function of upstream pressure. As suction pressure decreases, flow may become viscous and backflow will be a product of the differ­ential pressure times clearance conductance. With a further reduction in pressure.

102 Chapter 2.3: Vacuum Blowers

flow becomes transitional and finally molecular; flow resistance increases, con­ductance backflow becomes nearly constant and compression ratio increases. The results of this phenomenon can be observed in Figure 4, where, at pressures be­low 10 millitorr the pumping speed of the blower is essentially the same despite the difference in pumping speeds of the backing pumps.

2.3.4.2 Gas Species

Conductance through clearances varies not only with the pressure regime but also with the type of gas being pumped. In the turbulent and viscuous flow ranges, flow is inversely proportional to gas viscosity. As an example, the viscosity of air = 180 and hydrogen = 87 micropoise, and conductance and backflow for hydrogen would be twice that for air. In the molecular flow range, conductance is propor­tional to the square root of the ratio of the molecular weights, and conductance backflow for hydrogen would be 3.79 times that for air. The effect of the in­creased backflow (and compression ratio) on pumping efficiency when pumping air and hydrogen at two different compression ratios is illustrated in Figure 5.

Rg.5. Blower pumping speed efficiency curves.

100

90

5 80 o

an

S ° o 60 2 UJ o 50 UJ

UJ UJ

d) 30

i 20 Q.

10

10 10 10 10 10 10 10 10

INLET PRESSURE (TORR)

Effect of staging ratio (R) and gas species (air and hydrogen) on pumping speed efficiency.

2.3.4.3 Rotational Speed

Theoretical pumping speed increases with rpm, but clearances and conductance backflow remain constant for a given compression ratio. As RPM increases, effi­ciency increases because backflow becomes smaller relative to displacement.

2.3.5 Blower Pumping Speed Calculations 103

2J.4.4 Gas Carryback

Gases adsorbed on impeller surfaces and trapped in minor impeller volumes are carried from the high-pressure interstage and released at low inlet pressure. The effect of carryback at pressures above 100 mtorr are minimal. Polishing the im­peller surfaces can significantly reduce carryback.

2.3.5

BLOWER PUMPING SPEED CALCULATIONS

The net effective pumping speed of the first-stage blower in a blower system is a function of blower efficiency and the efficiency of the backing pumps. In stan­dard two-stage systems, the pumping speed of the system is usually published by the manufacturer. For nonstandard, two-stage, and multistage systems, the pub­lished pumping speed of the backing pump is the starting point for calculations. The pumping speed of the blower adjacent to the backing pump is calculated, and in multistage systems this progresses until the speed of the first-stage blower is determined.

The information usually supplied by the blower manufacturer for purposes of pumping speed calculations is a zero-flow maximum compression ratio (KQ) curve. Figure 6 illustrates typical compression ratio curves at three rotational

Fig. 6. Typical KQ zero flow compression ratios.

50

o P <

O CO Ul UJ Od Q_

o o X

<

45 h

40

35

30

25

20

15

10

'—r 2764 RPM

2215 RPM

"1 _L

10 10 10 10 10

P2 (TORR)

10 10 10

Data taken with three blower rotational speeds. Backflow (slip) decreases as a proportion of displacement with increased rotational speed.

104 Chapter 2.3: Vacuum Blowers

speeds. These compression ratio curves are obtained by blanking off the blower inlet (zero flow) and, with the blower and backing pump running, admitting air or other gas into the interstage volume between blower and backing pump. The interstage and resulting blower suction pressures are measured, and a curve is plotted of KQ (compression ratio ^2/^1) ^s a function of the interstage pressure P2. Following Equation (2) and letting Pj/Pi = K

E=l- KS^^/S^ (3)

When flow at the inlet is zero, backflow is balanced by pumping capacity, net pumping speed is zero, efficiency is zero, and

Ko = 5D/5FB (4)

The general equation for blower efficiency is derived from Equations (3) and (4).

E=l~ K/KQ (5)

Relating the throughput of the backing pump to blower pumping speed:

S2P2 = {ES]))Pi (ignoring temperature effects)

K=(ESi^)/S2 and K = EK^) (6)

(5D/'^2 — ^ D ~ theoretical compression ratio)

Combining Equations (5) and (6),

E = —^^- (7)

Sample pumping speed calculation:

SD = blower displacement = 600 m- /hr ^2 = backing pump pumping speed at P2 = 200 m^/hr P2 = 1 torr KQ = 44 From Figure 6 at P2 and 1800 rpm, for air KD = SD/S2 = 600/200 = 3 | 5, = £ 5 D = 0.94 * 600 = 563 mVhr

Ko 44 , 52P2 200*1 E = '— = = 0.94 P^=-^-^ = = 3.6*10~4orr

ATo + ^D 44 + 3 ' ^ 5i 563

2.3.6

POWER REQUIREMENTS

Power requirements in blower pumping systems are somewhat unusual. Many systems run at low operating pressure for most of their operating cycle. Time at

2.3.6 Power Requirements 105

Fig. 7. Power comparison for single- and two-stage blower systems.

1 STAGE SYSTEM PV=150,000

(TORR)

P3 200

(TORR)

P2 100

STAGE 2

2 STAGE SYSTEM PV TOTAL=100,000

PV = 100 X 500

50.000 STAGE 1

PV=50 X 1000=50.000

0 1000

V, (INLET)

1 STAGE BLOWER SYSTEM

Staging ratio of 2 assumed for this illustration.

0 500 1000

V, (INLET)

2 STAGE BLOWER SYSTEM

high pressure may be relatively short, and the large motor required to handle the high gas loads may be underutilized. Also, large motors running at low loads can cause power factor problems. To reduce initial and operating costs, systems have been designed with motors that have been much overloaded at peak loads. These have performed satisfactorily, but to prevent equipment damage safe time at over­load must be carefully predetermined. High-temperature windings and thermal overload protection must be used.

Power savings can be obtained by the use of two blowers in series or a blower-backing pump combination to obtain a given pressure rise. Contributing to this phenomenon is the pressure-volume relationship of Roots-type blowers.

The rotary positive blower follows a "square card" diagram. That is, if pres­sure rise is plotted against the constant displacement volume, the resulting curve throughout one complete cycle would be a rectangle under ideal conditions. In Figure 7, with a single blower, and a pressure rise of 200 torr, PV = 150,000. If two stages are used with an assumed compression ratio of 2, PV would equal 100,000. Power consumption is a direct function of PV, and the savings would be significant. The cost of using two blowers in series may outweigh the power sav­ings, but such systems have been built for factors other than economy and power savings have been achieved.

The power required by a blower is the sum of the power required for gas com­pression plus the power required to overcome the friction of bearings, seals, gears, etc. The values for friction losses must be obtained from the blower manufacturer. Power for compression is calculated as follows:

/ny = 3.7 * 10-5 ^ 5^ ^ PdiPilPi) - 1)]

PilPx^r, P\[r-\]=P2-P^ and

inv = 3.7 * 10-5 ^ ^^ ^ (p^ _ p^) (8)

106 Chapter 2.3: Vacuum Blowers

Note that power for compression is dependent on blower displacement and delta P but independent of gas species. In the absence of specific information, approxi­mately 20% may be added to the calculated power for the power to overcome fric­tion but it would be best to contact the manufacturer for this value. In addition to these power requirements, each blower has lower and (as might be expected) upper power limits. The lower limit is established by the power required to over­come friction and impeller inertia when accelerating the impellers from a stand­ing start to operating speed with no gas load. The upper limit is dictated by the mechanical strength of the operating components: impellers, gears, shafts, bear­ings, etc. These data must also be obtained from the manufacturer.

2.3.7

TEMPERATURE CONSIDERATIONS

2.3.7.1 Maximum Allowable Temperature Rise

A maximum allowable continuous gas temperature rise is established by the blower manufacturers to prevent seizure from excessive differential expansion be­tween the impellers and casing. The maximum rise allowed varies with blower size and manufacturer but a fair average may be 100°C, based on a maximum in­let temperature of 37°C. The differential expansion occurs because both the cas­ing and impellers absorb significant amounts of the heat of compression, but the impeller is partially insulated by vacuum and expands at a greater rate than the casing. Gas temperature rise may be calculated as follows:

T2 = degrees Kelvin discharge temperature Ti = degrees Kelvin inlet temperature

k = c^lc^ =1 .4 air F = temperature rise factor

The temperature rise factor, F, is obtained empirically and accounts for heat loss by radiation and convection. Figure 8 illustrates typical (varies with the blower) factors related to inlet pressure, compression ratio, and rotational speed. Note that as compression ratio decreases and rotational speed increases, the fac­tor increases. These trends increase efficiency, which tends to decrease the theo­retical temperature rise. However, the increased efficiency also results in less thermal losses, hence an increase in the factor.

When inlet temperatures are above 37°C, the maximum allowable temperature

2.3.7 Temperature Considerations 107

Rg.8.

o <

Q_

UJ

Temperature rise factor.

10''

10'

1 0 - ^ h

10"

i- R=2 < 2300 RPM^ '!!l^--^sr^'*!r"'^ ! 11800 RpM-rr^:^-'::: ' -^—'

,^^'' ^^^..^^^^^'^ ^ - 2 3 0 0 RPM

i y ^ ^ — 1 8 0 0 RPM

1

TYP TEMPERATURE RISE FACTOR (BASED ON AIR) VS INLET PRESSURE. BLOWER HANDLING AIR OR DIATOMIC GASES. _

.. 1 1 1 i 1 1 1 ; 10" 10" lO'- 10' 10^ 10-^

INLET PRESSURE (TORR)

Empirical data used to modify the theoretical temperature rise.

rise, Ta, must be derated as follows:

la corrected = Ta - 0.66(rin - ITQ) (10)

in = degrees C inlet temperature

Because thermal expansion is time dependent, in pumpdown situations it may be possible to exceed the maximum allowable temperature rise provided that the time from blower start pressure to the safe operating pressure does not exceed a certain limit. The blower manufacturer should be consulted for this limit. As a word of caution, if time between pumpdowns is very short and the cycle is repeti­tive, it would be best not to exceed the continuous allowable limit.

2.3.7.2 Maximum Allowable Discharge Temperature

The maximum allowable discharge temperature is based on the limits established by materials of construction such as seals, lubricants, etc. The manufacturer's rec­ommendations should be followed.

2.3.7.3 Maximum Allowable Compression Ratio

The curve in Figure 9 depicts a typical range of allowable pressure ratios versus inlet pressure based on the allowable temperature rise and average temperature

108 Chapter 2.3: Vacuum Blowers

Fig. 9.

Maximum allowable compression ratio.

10'' 10 10 10

INLET PRESSURE (TORR)

Typical values based on "average" temperature rise factor.

rise factors. It can be used for a first approximation to establish blower-backing pump relationships subject to a more accurate analysis.

2.3.7.4 Discharge Gas Cooling

Coolers that employ water-cooled surfaces in close proximity to the impellers at the discharge port may be used to cool the interstage gas that surges into the pumped gas pockets. The resulting cooling effect allows a higher compression ratio.

2.3.8

FLOW AND COMPRESSION RATIO CONTROL MECHANISMS

Several arrangements are available to keep power requirements low and the heat of compression within allowable limits.

2.3.8.1 Sequenced Start Systems

In sequenced start systems, the backing pumps are started first, and blowers are brought on line (starting with the first upstream blower) at safe operating pres-

2.3.8 Flow and Compression Ratio Control Mechanisms 109

sures by pressure switches or timers. In some variations, with several stages, two or more blowers may be switched into parallel at low pressure to provide more pumping speed.

2.3.8.2 Inlet Throttling

In inlet throttling systems, the blower and backing pump run at full speed contin­uously. A throttling valve at blower inlet controls blower inlet pressure to main­tain an allowable temperature rise or, if power limited, the differential pressure. In the former case, over-temperature switches are still required. The valve is par­tially open at high pressure and may be fully open at low pressure.

23.B.3 Blower Rotating Speed Control

Blower rotating speed control systems maintain allowable compression ratios by gradually increasing blower rotating speed as the inlet pressure decreases. Fol­lowing are some typical speed control systems.

1. Variable-speed AC or DC motors and drives. 2. Hydraulic or magnetic slip-type coupling between the blower and drive motor.

2.3.8.4 Bypass Valves

In a bypass valve type of system (Figure 10), an integral bypass valve with spring or deadweight loaded poppet or external bypass throttling valve, feeds a portion of the blower discharge gas back to the blower inlet. The valve is adjusted to pro­vide a fixed safe differential pressure across the blower. When the throughputs of the blower and backing pump are equal, the valve closes. The valve closes at pres­sure P = (P2 - PijIiPilPx) - 1. Figure 10 illustrates a stock blower with integral bypass. The blower and backing pump run at full speed from atmosphere to ul­timate pressure. Blower power required is a function of the pressure difference selected.

Inlet-throttling, blower rotating speed control, and bypass valve systems can produce higher pumping speeds from atmosphere to safe operating pressure than a sequenced start arrangement; see Figure 11. Assuming that the systems are ar­ranged to provide the same safe differential pressure, the gain in pumping speed would be the same for all three systems.

Fig. 10.

Blower with integral bypass valve.

-BYPASS VALVE

Pi T l Si

GAS

PRESS TEMP MVHR

DISCHARGE GAS

P2 PRESS T2 TEMP S2 MVHR

Safe differential pressure established by spring selection.

Fig. 11.

Typical blower-backing pump speed performance with compression ratio control. 3 0 0 0 I >—,

01

Q UJ UJ CL {/)

O Z CL

:s Z) CL

2500

2000

1500

1000

500

10

"1 ^n r- 1 n ^

10 10 ' 10 10

PRESSURE (TORR)

Both pumps run from atmosphere. Shaded area indicates increase of pumping speed (using ratio control).

23.8 Flow and Compression Ratio Control Mechanisms 111

2.3.8.5 Cool Gas Feedback Blowers

In cool gas feedback blowers, the blower is arranged to permit the admission of cool gas into the trapped volume of the blower where it mixes with and reduces the temperature of the discharge gas. These blowers are generally used at pres­sures of 200 torr and higher but they have also been used at lower pressures. The advantage of these blowers is that at higher pressues a compression ratio of 4:1 may be maintained, compared to 2.2:1 for uncooled blowers. This feature may make it possible to eliminate one blower stage when used alone to discharge to at­mosphere or as backing pumps in chain blower systems. The penalty paid is high power consumption, as discussed in Section 2.3.6.

Atmospheric air may be used for feedback or if special gases are being pumped that must be preserved, a portion of the discharge gas may be fed back through a cooler (see Figure 12).

Fig. 12.

COOLGAS FEEDBACK

BACKING PUMP

-ATMOSPHERIC AIR

BLOWER BACKING PUMP (WHEN REQUIRED)

Cool gas feedback systems: (a) Pumped gas recirculation system, (b) Atmospheric air cooled.

112 Chapter 2.3: Vacuum Blowers

2.3.9

LIQUID-SEALED BLOWERS

Blowers can be arranged to use water or other fluids (such as oil) injected into the inlet port to act as a cooling and sealing medium. With water injection, oper­ating pressure with a single-stage blower is approximately 100 torr and with a two-stage blower, 50 torr. Pressure limit is generally based on the water vapor pressure at a given temperature. Lower pressures are attainable with lower-vapor-pressure fluids. These blowers compete with liquid ring pumps. They generally use less water and power than liquid ring pumps, but their initial cost may be higher. As with liquid ring pumps, consideration must be given to the cost of dis­posal of contaminated liquids and the cost, when required, of a closed loop sys­tem to cool and recover expensive liquids or to conserve water.

2.3.10

SELECTED SYSTEM ARRANGEMENTS

The broad range of blower and backing pump types available provides the sys­tem designer with an opportunity to optimize systems for application require­ments. Following are some interesting examples of systems developed for spe­cific applications.

2.3.10.1 Multistage (chain blower) Systems

Multistage systems are generally employed when large volumes must be pumped down in a relatively short time. The multiple stages are used to maintain the al­lowable compression ratio across each stage when starting from atmosphere. An example is a 3-stage chain blower system (Figure 13) built to evacuate a 1308-m' chamber from 760 torr to 0.01 torr in two hours with a 20 torr L/S air leak. Be­ginning with the last stage, each stage is sequentially started with a few seconds time delay between stages to reduce initial power surge. With an approximately 2:1 staging ratio between the three stages, the interstage relief valves maintained a maximum allowable differential pressure of 400 torr. Excess air was discharged through the valves until the throughput of each backing stage equaled that of the preceding stage. The interstage heat exchangers removed the heat of compres­sion, maintained an approximately constant compression ratio, and limited the in­let temperatures to the downstream stages. With this type of system, the first-stage pumping speed of 10,270 mVhr was essentially maintained from atmosphere

2.3.10 Selected System Arrangements 113

Fig. 13. Multistage blower pumping system.

V _ _ ^ ^""^ HX HX

-INTERSTAGE-RLF VALVES SILENCER

EKl

BLO-1

760 TO 2 TORR 10,272 M^HR

BLO-2

5.246 M^/HR

MECH PUMPS

3.056 M^/HR

2 TORR TO BLANKOFF / 15,518 M^/HR

Three-stage, series-parallel operation. Pumpdown of 1,308 m^ space chamber from 760 to 0.01 torr in 100 minutes with 20 torr L/s air leak. Blowers are parallel at 2 torr to handle air leak at 0.01 torr. Courtesy: Stokes Vacuum Inc.

down to 2 torr. At 2 torr the first and second stages were valved in parallel to pro­vide 15,518 mVhr to handle the air leak, which at 0.01 torr equaled 7200 mVhr. Actual time to 0.01 torr was 100 minutes.

Peak power requirements for stages 1 and 2 were 159 kw and 75 kw respec­tively. Mean average power over a 20-minute period was approximately 60% of peak. Motors selected had ratings of 100 kw and 56 kw. Time at overload was ap­proximately 3 minutes. High-temperature insulation was used for the windings, and special over-temperature protection was provided by sensors buried in the windings.

2.3.10.2 Slush Hydrogen Pumping System

The slush hydrogen pumping system (Figure 14) was designed to pump 10,200 m-Vhr of pure hydrogen at 52 torr from a slush hydrogen source at 14° Kelvin. Through heat gain, the temperature at the inlet was approximately 249° Kelvin. To handle the cold gas, a three-stage liquid-sealed blower sys­tem was selected, using ethylene glycol for the sealing liquid. The glycol was circulated through a closed-loop system, where it was heated to 40°C to main­tain the blowers at a reasonable operating temperature. Liquid sealing increased blower efficiency, permitting the use of smaller blowers and direct discharge to atmosphere.

114 Chapter 2.3: Vacuum Blowers

Fig. 14.

Slush hydrogen pumping system.

STAGE 1

15,300 M V H R

• * •

STAGE 2

10.200 M^/HR

56 KW 224 KW

START AT-75 TORR

PERFORMANCE SPECS-

STAGE 3

5.100 M V H R

CLOSE AT 400 TORR

• 10,200 M^/HR AT 52 TORR.

• PUMP 249X H2 CONTINOUS IN RANGE FROM 40 TO 60 TORR.

• 25 TO 760 TORR TOTAL PRESSURE RANGE.

(—1 AIR/GLYCOL U SEPARATOR

DISCHARGE CHECK VALVE

SILENCER

Heated glycol injected into the suction ports of the blowers for sealing and to maintain blow­ers at reasonable operating temperature. Courtesy: Stokes Vacuum Inc.

At start of pumpdown, stages 2 and 3 were in operation with a portion of the discharge from stage 2 bypassed around stage 3 through a pressure-sensitive check valve directly to the air-giycol separator and the silencer. At approxi­mately 400 torr, the check valve closed, and the discharge of stage 2 was handled entirely by the third stage. At 75 torr, the first stage was started and hydrogen flow commenced. Pressure decreased to 52 torr as pumping progressed.

2.3.10.3 Product Dehydration System

The product dehydration system (Figure 15) was designed to hold 6 torr with 11.4 kg/hr air plus 100 kg/hr water vapor. A blower-condenser-mechanical pump system was employed to provide the net effective pumping speed of 18,500 m^/hr. The blower was used to increase the pressure at which the condensers would op­erate— in this case, from 6 to 18 torr. Two condensers, each with receivers, were required to handle the entire water load, with a chiller furnishing 10°C cooling water for the condensers. An automatically operated pressure control valve lo-

2.3.10 Selected System Arrangements 115

Rg.l5. Product dehydration system.

6 TORR 18 TORR

z::

IQ-C

WATER CHILLER

T

30 KW

23.700 M V H R

rPERFORMANCE SPECS-

• 11.3 KG/HR AIR

• 1 0 0 KG/HR H2O VAPOR AT 27'C 8c 6 TORR

PRESSURE CONTROL VALVE

EXHAUST

i I 1 VALVL /-->^

^ I r i O wc hH I 22 KW ^^ a I—I c FT AT

wcv SET AT

82*C 1,240 M V H R

13*C ^

High water vapor load (100 kg/hr) plus small air load (11.3 kg/hr) to be pumped at 6 torr. The backing pump handles the air load plus partial pressure of water vapor at 10°C condenser tem­perature. Courtesy: Stokes Vacuum Inc.

cated between the condensers and the mechanical pump maintained a pressure of 18 torr in the condensers to prevent reflux of the condensed vapor. Most of the water was removed by the condensers, and the backing pump was small because it was only required to handle the air load plus the small water vapor load corre­sponding to the partial pressure of the water vapor in the condenser. The backing pump was operated at 82°C.

The condensing system was designed to operate for eight hours. With the addi­tion of automatic valving to drain the receivers while operating, the system could have been made to operate on a continuous basis.