gas powered liquid recirculation compared to mechanical pumps

8/9/2019 Gas Powered Liquid Recirculation Compared To Mechanical Pumps

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One of the first critical decisions in the design of an ammonia refrigeration system is the choiceof liquid feed to the evaporators. Overfed liquid recirculation is usually the system of first choicefor three reasons:

Increased coil efficiency achieved as a result of wetting the entire coil surface: coil ratings are

approximately 18% higher than for thermal expansion valve feed.

Inherent compressor protection afforded by the suction line accumulator.

Simplified maintenance with only a liquid solenoid and hand expansion valve as the control

devices.

The next decision, then, is which type of liquid recirculation system to use. The choice is either a

mechanical pump, or a gas pressure recirculation system. The obvious questions are:

How much energy does it take to recirculate refrigerant using gas pressure, and

How does it compare with a mechanical pump?

In fact, the launching pad for this paper was a corporate energy manager who asked this question

about a design that was proposed incorporating a Constant Pressure Liquid type of gas pressure

recirculation system. The application was a small cooler/freezer/dock addition, far removed from the

compressor room, and the plant wanted a simple system without cavitation or gas binding problems

or potential seal leaks an ideal application for constant pressure type recirculation. The energymanager, however, was convinced that if a Constant Pressure Recirculation system was used, the

entire plant would have to be operated at an artificially high condensing pressure, and would thus

penalize the entire system with higher operating costs. I knew this was not necessarily the case, but,

that this misconception is common in the industry. In fact, the system he was concerned about was

installed using a constant pressure recirculation system, and is now in its second year of operation.

During winter conditions, this system operates at 70 psig condensing pressure the lowest of all

their ammonia refrigeration plants.

When looking for references to estimate operating costs, essentially no information was found in the

published literature to serve as a guide. The definitive work by Lorentzen (1965) dealt primarily with

piping design and system efficiencies, but did not provide any model for estimating the operating

costs. Tests performed by Lorentzen on a Single Pumper Drum type of system provide unique insight

into the dynamics of gas pressure recirculation. The measurements of energy efficiency for this type

of system, however, are not at all a good model for Constant Pressure liquid recirculation for several

reasons:

Hot Gas pressure on the Liquid Transfer Unit was not regulated, exposing the system to

increased losses.

Gas Powered Liquid Recirculation Compared To Mechanical Pumps


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Transfer time is a function of liquid demand of the system. This can expose the Liquid Transfer

Unit to warm transfer gas for long periods of time during low load conditions, increasing ther-

mal losses. In a Controlled Pressure Liquid system, the transfer time is constant, and should be

relatively short.

Refrigerant flow to the coil is not constant, but surging, in a Single Pumper Drum type system.

The intent of this paper is to present a simplified mathematical model used to analyze the energy

required to operate a Constant Pressure Liquid recirculation system. The basis for the analysis pre-

sented here grew out of an actual system design, in which there was both a 120 ton two-stage, -40F

load, and a 75 ton, single stage, 0.F load. The design criteria for each type of recirculation system

will be explained separately, and the calculation procedures clearly defined. The analysis has been

expanded to two include two other common suction temperatures, -20F two stage and +20F single

stage. All the energy calculations have been normalized to a per-100-ton basis for convenient ref-

erence. Besides comparing operating energy with that of a pump, the analysis reveals some critical

observations that result in a list of recommendations for design and operation of an efficient

Constant Pressure Liquid recirculation system.

System DescriptionsThe basic elements of a mechanical pump recirculation system and the Constant Pressure Liquid recir-

culation system can be compared in Figure 1 for a simplified single stage system.

In a mechanical pump system, the refrigerant liquid level is controlled in an accumulator at a height

above the pump sufficient to meet the pumps net positive suction head (NPSH) requirements. The

pump boosts the liquid pressure, typically 25 to 40 psig above suction pressure, making it subcooled,

and the liquid refrigerant is circulated out to the coils, where the flowrate is regulated by a hand

expansion valve or an automatic flow regulator. The heat from the evaporator coil vaporizes a por-

tion of the liquid, and that gas flows with the overfed liquid back to the accumulator, where the sat-

urated gas is returned to the compressor, and the liquid falls back down to the controlled level in thevessel. The ratio of liquid fed to the coil divided by the liquid boiled off in the coil is referred to as

the recirculation rate.

In a Constant Pressure Liquid recirculation system (referred to as a CPR system), also shown in Figure

1, the high pressure liquid is fed to a Controlled Pressure Receiver (CPR) where it is flash-cooled down

to the liquid supply pressure, which is regulated to the minimum appropriate supply pressure for the

coils. Just like the pump system, this pressure is usually 25 to 40 psig above suction pressure, but can

be adjusted merely by the setpoint of the CPR relief regulator. And liquid is circulated through the

piping and coils, just like a pump recirculation system: subcooled liquid from the CPR is fed out to

the evaporators where the flowrate is adjusted by a hand expansion valve, with the overfed liquidand vapor from the evaporator returned to the accumulator.

The accumulator in a CPR system will remain essentially empty as the overfed liquid is drained by

gravity from the accumulator into the transfer vessel, referred to as a Dump Trap or Liquid Transfer

Unit, LTU for simple reference. Liquid flows into the LTU through a low pressure drop inlet check

valve, while the displaced gas is vented back to the accumulator through a 3-way solenoid valve.

When the LTU is full, a float switch initiates the transfer cycle by switching the 3-way valve from its

vent position to the pressurize position, connected to a higher pressure source of transfer gas.


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The transfer gas is regulated to a minimum pressure adequate to push the liquid refrigerant out of

the LTU and over to the CPR through the outlet check valve. The cold liquid is returned to the bottom

of the CPR where it mixes with a portion of the make-up liquid and is recirculated back out to the

evaporators. The transfer process is terminated by a timer, adjusted for low level in the LTU, which

switches the 3-way valve from the pressure port back to the vent port, allowing the LTU to refill for

the next cycle.

An overview of the process on a pressure-enthalpy diagram is helpful. Figure 2 shows a simplified

diagram for a single stage pump recirculation system where the liquid is separated from the gas

stream, returning only saturated gas to the compressor, point 6. Saturated liquid at point 1 drains

from the condenser to the receiver and is fed to the accumulator. The high pressure liquid expands

across the hand expansion valve into the accumulator at point 2, then is flash cooled to saturated liq-


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uid at suction pressure, point 3. The pump raises the liquid pressure to point 4. The subcooled liquidis circulated to the coils, is throttled across the hand expansion valve back to essentially saturated

liquid at point 5. Some of the liquid is boiled off in the evaporator to saturated gas at point 6, which

returns in two-phase flow with the overfed liquid at point 5 to the accumulator. The compressor

takes saturated suction gas at point 6 and compresses it up to condensing pressure at point 7.

Figure 3 shows a similar simplified pressure-enthalpy diagram for a CPR pressure recirculation system.

Saturated liquid from the condenser at point 1 expands to CPR pressure at point 2 where it is flash

cooled down to saturation, point 3. Mixing with return liquid from the LTU subcools the liquid being

fed out to the coils to point 4. The subcooled liquid expands across the hand expansion valve to coil

pressure at point 5. As with the pump recirculation system, the saturated refrigerant gas, point 7,boiled off in the evaporator is mixed with overfed liquid, point 6, in two-phase flow in the return line

to the accumulator, where it is separated from the liquid and then compressed up to condensing

pressure, point 8.

Saturated liquid at suction pressure, point 6 drains into the LTU. Transfer gas, point 4a, regulated

from condensing pressure down to 10 to 15 psig above CPR pressure, is used to pressurize the LTU,

sending the saturated cold liquid over to the CPR vessel where it mixes with make-up liquid, saturated

at CPR pressure, to provide subcooled liquid for the evaporator feed, point 4.


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Analysis Criteria

The design assumptions used in this analysis are:

Single Stage Load: 75 tons at 0.F and +20F Suction

Two Stage Load: 120 tons at -40F and -20F Suction, 20F Interstage.

Annual Average Condensing Temperature: 80F

Recirculation Rate: 2 to 1 and 3 to 1

Pump liquid supply pressure: Suction Pressure+ 30 psi

CPR liquid supply pressure: Suction Pressure + 30 psi

Transfer gas pressure: CPR pressure + 15 psi

Vertical Liquid Transfer Unit (LTU)

Suction and Condensing Temperatures

The analysis of the real system design conditions, 75 tons at 0.F and 120 tons at -40F, was expanded

to 20F and -20F using the same mathematical models. Compressor energy consumption was based

on manufacturers data for existing installed compressors; reciprocating compressors at 0.F suction,

screw compressors at -40F and +20F high stage. An overall annual average condensing tempera-

ture of 80F was assumed.

Recirculation Rate

Coil recirculation rates for this analysis were selected at 2:1 and 3:1 for both mechanical pumps and

gas pressure recirculation. The issue of optimum recirculation rate for mechanical pumps will always

be one of great and varied discussion. The primary variables that contribute, however, include:

System piping sizes, both supply and return

Pump characteristics

Coil circuiting Oil return

Several observations from the literature:

Lorentzen (1965) stated,

To be on the safe side, it is customary to design for a circulation ratio of approximately

4 to 5 . . . and it is advisable to play it safe. As a result, overcirculation is probably the

more frequent occurrence . . .

One large industrial ammonia evaporator manufacturer has recently published their recommen-

dations for pump recirculation rates to be a function of suction temperature as follows:+20F 4 to 1

-25F 3 to 1

-40F 2.5 to 1

For the particular system (piping, coil, and pump) tested by Lorentzen (1965), he determined for

mechanical pumping, the optimum circulation ratio is approximately 5.

For the same system, Lorentzens tests with a single gas pump system showed, the optimum

circulation ratio was in this instance 2.


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Lorentzen (1965) plotted the overall heat transfer coefficient (U) of an air cooler coil as a

function of recirculation rate (n), at coil circuit loadings (q) between values of 200 and 1400

Btuh per square foot - degree F. The original data is shown in Appendix A. It can be seen that,

for circulation rates between 1:1 and 4:1 (circulation ratio of 1 to 4), the overall heat transfer

coefficient varied a maximum of only 6% at q = 200, down to 2% at q = 1400.

Assuming that coils are designed for their maximum loads with high circuit loading rates, it is obvious

that as the actual load on the coil decreases, and the circuit loading (q) drops below design, the recir-culation rate (n) will increase because the coil flowrate (GPM) will not change, being set once for the

maximum coil load.

This data all seems to indicate that the overall heat transfer coefficient for an air evaporator is rela-

tively insensitive to recirculation rates larger than 1:1. As a practical matter, specific coil circuiting

may be considered by a manufacturer when low recirculation rates are specified.

The coil recirculation rates selected for this study, 2:1 and 3:1, were used for both mechanical pumps

and gas pressure recirculation, because a common reasonable rate needed to be selected. In all likeli-

hood, a mechanical pump system would be field set to operate nearer twice these recirculation rates,

because as Lorentzen states,. . . it is advisable to play it safe. However, the selection of recircula-tion rate is much more critical for a gas pressure recirculation system than it is for a mechanical pump

when considering the energy required to operate, as this analysis will indicate.

Pump Supply Pressure

The pump supply pressure was selected as 30 psi over evaporator suction, a very common specifica-

tion for refrigerant pumps. The Control Pressure Receiver would be set for the same liquid supply

pressure.

Transfer Gas Pressure

The pressure of the transfer gas is regulated from its source to be 15 psi above the CPR pressure.

For a single stage system, the source is condensing gas pressure, but for a two-stage system flash gas

generated in another CPR, a Flashcooler, or the Intercooler may also be used, effecting significant

energy savings, as will be demonstrated.

Liquid Transfer Unit Orientation

Vertical LTUs were analyzed because of limited space available in the compressor room.

Constant Pressure Liquid Recirculation System

Several important initial observations are made about the energy required to circulate liquid refriger-

ant using gas pressure: All the energy comes from the transfer gas, which is at higher pressure and warmer tempera-

ture than the liquid being transferred

The energy consumed all shows up as direct heat load on the refrigeration system, or equivalent

tons of refrigeration

The work is done by the compressor(s) at their operating pressure conditions, rated in BHP per

ton. This is converted to total KW of electrical consumption in this study, to allow direct com-

parison with a mechanical pump.


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This analysis assumes that all the energy required to operate the CPR system is expended in thermodynamic

losses in the Liquid Transfer Unit, transferring cold liquid refrigerant from the accumulator to the CPR, andthat losses in the CPR vessel itself and interconnecting piping are either negligible or ar similar to losses in

a pump recirculation system.

Three distinct thermodynamic losses have been considered as the most significant in the operation of the

Dump Trap, or Liquid Transfer Unit:

1. Transfer gas - at the end of each transfer or dump cycle, the additional suction gas created by

venting the compressed transfer gas to the accumulator,

2. Gas condensation on cold liquid - during the transfer cycle the amount of warm transfer gas that

is condensed into the surface of cold liquid,

3. Vessel warming - during the transfer cycle, the mass of metal in the vessel which is warmed up bycondensation of transfer gas.

Each of these three components will be discussed separately, and its mathematical model established.

1. Transfer Gas Load

In order to determine how much transfer gas is added to the normal suction gas load, an up-close look at

the construction and operating levels for a Liquid Transfer Unit, or Dump Trap vessel, shown in Figure 4, is

required.


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The amount of liquid transferred in a dump cycle is the difference between the volume in the vessel

at the start of the cycle, at the initiate level, and when the cycle is terminated. The volume of

pressurized transfer gas that is added to the compressor load is equal to the amount of liquid trans-

ferred, plus an inefficiency; the volume of transfer gas in the top of the LTU between the initiate liq-

uid level and the top of the vessel head. This additional dead head volume represents compressed

transfer gas which does no useful work, but is added to the compressor suction after each transfer

cycle.

The percentage of this dead head volume to the total volume of liquid transferred per cycle can be

thought of as a measure of volumetric inefficiency. Conversely, the volume of liquid transferred as a

percentage of total transfer volume capability (dead head plus volume transferred) has been

defined in this study as a volumetric efficiencyfor the vessel.

For vertical LTUs the initiate level, as shown in Figure 4, has traditionally been set for pragmatic rea-

sons - there needs to be enough space to connect a side-mounted float switch; the transfer pres-

sure/vent nozzle connected to the three-way valve needs space and piping details to prevent high

velocity gas from jetting into the cold liquid and to insure uniform pressurization; and the level eye

needs to be welded in.

There is 2 of straight length on the skirt of the semi-elliptical heads used in fabricating the vessels.

Because the level eyes are approximately 2 OD and it is best to keep welding about 1 away from

the seam weld, and because this also allows space for the initiate float switch and vent/pressure

nozzle, the initiate level is normally placed about 2 below the top vessel seam. While the lower level

eye is usually placed 2 above the bottom seam, the transfer timer should be set so that the vessel

clears almost completely of liquid but does not allow transfer gas to blow into the CPR. The assump-

tion in this study is that termination level is set 2 below the bottom seam.

Conventional construction for a 20 x 48 vertical LTU relates to a volumetric efficiency of 86%,and for a 24 x 42 vessel, volumetric efficiency of 81%, as shown in Table 1.

The volumetric efficiency of the vertical transfer vessel can be improved significantly by moving the

transfer initiation level up and, thus, minimizing the dead head volume of gas. Without major fab-

rication problems, the initiate level can be changed to 2 above the top seam, rather than 2 below,

resulting in an 11% increase in volumetric efficiency from 86% to 97% for the 20 vessel; from 81%

to 94% for the 24 vessel.

The volumetric efficiency of a horizontal vessel is inherently greater due to the geometry of the vessel.

Conventional construction places the initiate level 2 from the top of the shell, resulting in volumetric

efficiencies of 93% for the 20 , and 95% for the 24 vessel.

The performance of the high efficiency vertical vessel configuration has been calculated along with

that of the conventional vessel to determine its effect on operating energy consumption.

Although the transfer cycle isperiodic, the effect of the transfer gas added to the compressor suction

is based on an average calculated mass flowrate per hour, and is equated to compressor load by the

specific mass flowrate for the system. The mass flowrate through the evaporator coils is based on


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the design load (tons), recirculation rate or overfeed ratio (n), and latent heat of evaporation (H fg)

at suction conditions,

Coil Massflow = Tons n 12,000/60 Hfg @ suction temp (lb/min)

Using the density of liquid () at saturated suction conditions, the flowrate in GPM is,

Coil GPM = Coil Massflow 7.48 (gal/cu ft) /liquid (gal/min)

The mass flowrate in the Liquid Transfer Unit has a recirculation rate equal to 1 less than the coil

recirculation rate, which will determine its average GPM liquid flowrate.

LTU GPM = Coil GPM (n-1) / n (gal/min)

The transfer gas volume flowrate, on an average basis, then, is the liquid flowrate divided by

the LTUs volumetric efficiency (). The mass flowrate is calculated using the transfer gas specific

volume, ().

Transfer Gas Massflow = LTU GPM / 7.48 (gal/cu ft) (lb/min)

The amount of superheat in the transfer gas is also calculated. It is estimated using Hg from

saturated hot gas to saturated suction gas temperature.

Superheat = Massflow (Hg sat, transfer press - Hg sat, suction) 60 (Btu/hr)

The load on the refrigeration system is calculated by use of the specific suction gas flowrate per ton

of refrigeration; using saturated suction gas density (), Hf at makeup liquid conditions and Hg at

saturated suction temperature,

Specific Suction Massflow = 200/ (Hg sat suct - Hf liquid makeup) (lb/min-ton)

Transfer Gas Load = Superheat + (Massflow/Specific Suction Massflow) 12,000 (Btu/hr)

2. Gas Condensation to Cold Liquid

The liquid surface in the Liquid Transfer Unit during the transfer cycle will be very stable in a properly

designed vessel. Extreme agitation or sloshing indicates that the transfer gas stream is jetting

directly into the cold liquid, the result of improper internal design and the source of much wasted

energy. Dump Traps manufactured by the few specialty system manufacturers will be observed to rise

in pressure rapidly and transfer with a very quiescent gas/liquid interface.

The introduction of transfer gas into the LTU during the transfer cycle creates an interface between

the warmer, higher pressure gas and the horizontal plane of cold liquid refrigerant, which results in a

complex transient heat/mass transfer process during the time duration of the dump cycle. The liter-

ature for this type of laminar heat/mass transfer is modeled primarily around vertical flat metal sur-

faces and horizontal tubes, both conditions where the condensed liquid runs off of the colder sur-

face; not at all like a horizontal flat liquid surface where the condensed liquid will collect and form a

film of liquid at the temperature of the saturated transfer gas. The warm liquid will in turn, impedethe heat transfer from the cold liquid as it both raises the surface temperature and adds thermal

resistance in the liquid film.

It seems apparent that the heat transfer coefficient at the gas-liquid interface during the dump

cycle will peak rapidly as the vessel first pressurizes, but will drop off rapidly as the cycle progresses,

and refrigerant is leaving the vessel.


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The most reasonable assumption available from the literature for the heat transfer coefficient at this

horizontal flat interface would seem to be the Nusselt analysis for condensation on a horizontal tube,

considering values for hm as tube diameter approaches infinity. A nomograph solution for the

Nusselt equation is published in the Chemical Engineers Handbook and is included as Appendix B.

The logarithmic reference axis nDt has been extrapolated to large numbers to indicate limits as

D , and the film temperature is estimated as an average between saturated suction and saturated

transfer gas temperatures. For values of t in the range of 40F to 60F, and D , the value of nDt

will be large rapidly, and hm will approach small values. At 10,000 the value for hm on a horizontalsurface would be 210 230. A conservative estimate of 250 has been assumed for all cases in this

study.

For the vertical LTU of diameter, D, the heat transfer from transfer gas to cold liquid surface is then

calculated as,

Q = hmA T (Btu/hr)

where: hm = 250 (Btu/hr-ft2-F)

A = D2 /4 (ft2)

T = (Tsat transfer gas - Tlow temp liquid), (F)

3. Vessel Warming

The dump cycle takes place as the transfer gas pressurizes the LTU vessel from the top, displacing

the cold liquid out through the check valve to the CPR vessel. In the process, the warmer transfer gas

condenses on the inside of the metal vessel walls and warms up the vessel as the liquid is displaced.

The condensed liquid is added to the transferred liquid as it is returned to the CPR. Anyone who has

observed a transfer cycle on an uninsulated LTU vessel would remember seeing the frost line move

down the vessel during the transfer cycle.

The heat gained by the vessel mass during the transfer cycle is removed by the cold liquid refrigerant

as it fills up the LTU for the next transfer cycle. And all this heat shows up as an added refrigerationload to the system.

The heat transferred to warm up the vessel on each cycle may be considered as,

Q = M Cp Fw T (Btu/cycle)

where: M = vessel mass exposed to transfer gas, (lb)

Cp = specific heat of steel, 0.12 Btu/lb F

Fw = Warming Factor; a judgment estimate of the average percentage of

full warming or cooling of the vessel mass exposed to transfer gas.

T = (Tsaturated transfer gas - Tcold liquid), (F)

Table 1, the Vessel Analysis Summary, includes the mass of the vessel, plus the percentage of the ves-

sels metal mass that is exposed to transfer gas in each cycle. If liquid is drained completely from the

LTU each cycle, then the full vessel mass is exposed to transfer gas; in reality, the termination level is

usually set near the lower sight glass level. This study assumes termination of the dump cycle will be

2 below the bottom vessel seam, leaving the bottom of the head submerged in liquid refrigerant at

the end of the cycle.

The Fw warming factor represents a judgment estimate between 0 and 1.0 to reflect the observa-


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tion that all the vessel mass exposed to transfer gas does not reach an equilibrium temperature with

the warm saturated transfer gas during the dump cycle, or with the cold saturated liquid during the

fill cycle. For purposes of this study, a warming factor of 0.8 was assumed.

Lorentzen (1965) plotted temperature profiles of both the metal vessel and the gas or liquid inside of

a vertical Liquid Transfer Unit with an extended transfer cycle of approximately 65 seconds. The tran-

sient profiles are included as Appendix C, and reflect the following:

Metal at the top of the vessel not submerged in liquid never varies far from the saturatedtransfer gas temperature maximum variation of 5F.

Metal at the bottom of the vessel which stays submerged never varies far from the saturated

liquid temperature less than 1F.

Very little warming of the liquid leaving the vessel only rises 2F during the entire 65 second

transfer cycle.

Metal walls continue to cool during the transfer cycle until exposed to the warmer transfer

gas, and then never reached the transfer gas temperature.

Inspection of these temperature profiles indicates that a warming factor of something less than 1.0

is very realistic.

The average system load for vessel warming is then calculated using the heat gain per transfer cycle

multiplied by the number of transfer cycles per hour. It is assumed that a transfer cycle takes 30 sec-

onds.

Electrical Consumption

With the three heat load components calculated, the total heat gain to the system is known and may

be represented as a percentage of the total evaporator load, a very important number. However, this

does not tell the whole story. The actual cost of operation, or energy consumption, must be deter-

mined to provide a meaningful standard of comparison with a mechanical pump.

Single Stage Systems

The calculated heat load is a direct refrigeration load, removed by the compressor(s). The compressor

manufacturers specific performance/power consumption ratings (BHP per ton) are used to determine

the compressor input in brakehorsepower, then converted to KW electrical energy consumption,

assuming large motor efficiency of 92%.

Two Stage Systems

Gas pressure recirculation applied in two-stage refrigeration systems allow for much more creative

system design and energy management. One example is if the transfer gas is taken from the

condenser gas and regulated to the appropriate pressure, then the energy consumption would show

up as a low stage load and would be calculated the same as any low stage load, considering the low

stage compressor load plus the booster balance additional high stage compressor load.

However, when the transfer gas is flash gas taken from a higher pressure CPR, a Flashcooler or an

Intercooler, as in the case of the -40F load in the real design system, the useful work has already

been done by the evaporation process in flash-cooling the liquid refrigerant. As can be seen in the

simplified 2-stage CPR system diagram, Figure 5, the mass flowrate of flash gas used for transfer gas


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represents load taken away from the High Stage compressor suction. So when the electrical consump-

tion for this transfer gas is calculated, the low stage compressors see the full refrigeration load, butthe high stage compressors only account for the additional heat due to low stage heat of compres-

sion. Thus, for the -40F two stage system with Low Stage Compressors at 1.18 BHP per ton and High

Stage Compressors at .895 BHP per ton, the total compressor energy expended is calculated as 2.30

BHP per ton using condenser gas for transfer,

BHP/TON = 1.18 + (1 + 1.18 2,545/12,000) .895

= 2.30

However, when using flash gas as the source of transfer gas, this results in only 1.40 BHP per ton.

BHP/TON = 1.18 + (1.18 2,545/12,000) .895

= 1.40

Pump Recirculation System

The total electrical consumption for a mechanical pump recirculation system has two components:

the electrical energy required to operate the pump motor, and the total pump energy input to the liq-

uid, represented as a heat load to the refrigeration system.

In this analysis, a centrifugal open drive pump was used, perhaps one of the most popular or com-

mon models used for refrigerant recirculation systems. It is the standard provided on many compa-

nies packaged recirculators. The pump was selected by the manufacturer for the flowrate and pres-

sures specified. The pump performance curve is shown in Figure 6.


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The pump flowrate is equal to the refrigerant flowrate through the evaporators at the design recircu-

lation rate, plus any by-pass liquid, which is recommended by the manufacturer to maintain a mini-

mum flowrate through the pump to avoid cavitation or gas binding. In this case, the manufacturer

recommends against extended operation at less than 7 GPM, so this amount was assumed for by-pass

flow. Brakehorsepower ratings were taken directly from the pump curve at 30 psi of total head.

The KW electrical consumption of the pump motor is calculated based on total flow at 30 psi, assum-ing a small motor efficiency of 85%. The total pump electrical input, excluding motor inefficiency, is

calculated as a refrigeration system load, and the compressor motor power to handle that load is

added to the pump motor power input for the total electrical energy required for the mechanical

pump system, expressed in KW.

One observation is that because ammonia has such a high latent heat of vaporization, which makes

it such an ideal refrigerant, the GPM flowrate requirements are so small for recirculation systems less

than approximately 200 tons that most pump selections, including this one, are not in their most


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efficient operating range. Thus, the particular system requirements (120 tons and 75 tons) do not

show the pump performance in its best light.

DISCUSSION

The mathematical models just described were integrated with ammonia properties and vessel charac-

teristics into a spreadsheet analysis and system summary for the 0.F single stage and -40F two stage

loads, and were then expanded to include +20F and -20F operating conditions. Some summary

sheets and sample calculations are included in Appendix D through G. A hand calculation for eachequation is shown for reference in Appendix H.

Figure 7 summarizes the results for 75 tons load, 0.F and 20F Single Stage systems for recirculation

rates of 2:1 and 3:1. It can be seen that at 2:1 recirculation rates, the gas pressure recirculation

system will operate at slightly lower energy consumption than the pump, but at 3:1 recirculation rate,

the pump has lower energy consumption. The critical observation is that for 3:1 recirculation rate,


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the energy consumption for a gas powered recirculation system is 100% greater than 2:1 recirculation

rate. The LTUs simply transfer twice the amount of liquid at 3:1 recirculation rate. The mechanical

pump, however, changes very little in total energy of operation, because the pump efficiency is

increasing with larger flowrates.

Figure 8 shows the comparison of electrical consumption for 120 tons at -40F and -20F suction,

two stage systems. It is observed that for 3:1 recirculation rates, the mechanical pump systemmay consume only 25% to 50% of the energy of a gas pressure recirculation system, but at 2:1

recirculation rate it is much closer, and, in fact, if designed and operated correctly, a gas pressure

recirculation system at -20F suction may operate at lower energy cost than a mechanical pump at the

same recirculation rate, and even lower if the pump is operated at their more common recirculation

rates of 4:1 to 6:1.

It is apparent that the differences between mechanical pumps and gas pressure recirculation become

greater at lower suction temperatures. This is attributed to two factors that increase simultaneously


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and make operating costs higher:

The refrigeration system compressors consume more energy per ton as suction temperatures

decrease, and

The vessel warming losses also increase as the suction temperature decreases.

The percent contribution of each calculated heat loss component to the total energy consumption for

a CPR gas pressure recirculation system is shown in Figure 9.

It is apparent that the load calculated for condensing gas into the cold liquid is the smallest compo-

nent. As suction temperature increases, the larger portion of load comes from the transfer gas load

added to compressor suction, due to the higher mass density of the gas at the elevated transfer gas

pressure. As suction temperature decreases, however, the larger portion of the load shifts to the

vessel warming effect. This is attributed to the greater heat gain in the LTU vessel when the refriger-

ant temperature is much lower.

The tabulated data shown in Appendix D to G also indicates that the effect of superheat on the

transfer gas load is small, representing only 4% of the transfer gas load.


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Figure 10 shows the heat load calculated for operation of a CPR gas pressure recirculation system as a

percentage of total evaporator load. With a conventional LTU, 3:1 recirculation rate, the percentage

varies from 3.2% to 3.8%. With a high efficiency style, the percentage drops to 2.8% to 3.4%, or

an 11% reduction. At 2:1 recirculation rate, the percentages are just half, or in the range of 1.4% to

1.8% of evaporator load.

It is also observed that the High Efficiency configuration for the vertical LTU has a significant impact

an operating energy. The effect is linear with the volumetric efficiency 11% higher efficiency trans-

lates to 11% lower operating costs, as 11% fewer dump cycles are used to transfer the same amount

of liquid.

In Figure 11, the total KW energy consumption calculated for each suction temperature and gas pres-

sure recirculation system was normalized, to be expressed in terms of KW per 100 tons of evaporator

load. This may prove helpful in estimating the energy consumption of systems of different sizes, con-

figuration and recirculation rates. Figure 11 also illustrates some of the most important observations

made as a result of this study: At 2:1 recirculation rate the operating cost would be half that at 3:1.

The volumetric efficiency of the vessel directly affects the operating cost; an 11% difference is shown

here. The effect of using flash gas as the source of transfer gas is tremendous in a two-stage or

economized system: 40% savings at -40F and 51% savings at -20F.


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OBSERVATIONS

The most critical observations made as a result of this study are:

1. Operating cost - A gas pressure recirculation system may operate at the same cost, or below

that of a mechanical pump, but it must be properly designed, adjusted, and controlled.

2. Suction temperature and system size. The lower the suction temperature (-30F and below),

and the larger the system size (200 tons and above) the more likely that a mechanical pump system

will operate at lower cost.3. Recirculation rate - the objective for a gas pressure recirculation system should be the minimum

acceptable; the upper target should be 2:1. Remember, at 3:1, the operating cost is doubled. Higher

recirculation rates will be very inefficient.

4. Volumetric Efficiency - initiate the Dump cycle as high up in the vessel as practical. Savings of

11% is realized by moving from 2 below the top seam to 2 above the top seam in a vertical LTU.

5. Transfer Gas - should always be regulated to the minimum pressure necessary. Use flash gas

where possible. Operating cost savings of 40% is realized at -40F suction and 51% at -20F suction.


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RECOMMENDATIONS

Whether maintaining an existing or designing a new CPR gas pressure recirculation system to operate

most economically, the objectives should be

Minimum recirculation rate - consider 2:1 at design loads.

Optimum volumetric efficiency of LTU vessel.

Regulated Transfer Gas - use flash gas where possible in 2-stage or economized systems.

Some practical recommendations are:

1. Adjust hand expansion valves - to set the system to minimum recirculation rate requires setting

individual hand expansion valves at each coil. Use hand expansion valves whose manufacturer pub-

lishes pressure drop and capacity ratings - typically, tons of refrigeration based on number of turns

open and pressure drop across the valve. It helps to have a feel for where you are.

2. Use thermostat control always shut off liquid supply to coils when temperature conditions

are satisfied. There will be a direct reduction in operating cost of the CPR system.

3. Adjust LTU Controls Set the transfer gas pressure as low as practical, usually 10 to 1 psig

above CPR pressure. With transfer gas pressure set properly, adjust the transfer timer so the vessel

clears almost completely of liquid before terminating the cycle.

4. Determine transfer time With transfer pressure properly set, the transfer time should not

exceed approximately 30 seconds. Excessive transfer time causes increased losses and indicates the

need to change to a larger size outlet check valve and/or transfer line size. For a new system, specify

30 seconds maximum transfer time from the system manufacturer.

5. Specify high volumetric efficiency for LTUs Look for the manufacturer to provide an efficient

vessel design, particularly if a vertical vessel is to be used. Current conventional designs can be

enhanced to provide initiation at a higher level.

6. Know the transfer volume have the vessel manufacturer provide the estimated gallons of

liquid transferred in a dump cycle, or calculate it yourself using basic vessel dimensions.

7. Count the number of transfer cycles digital counters costing less than $50 can be installed

on the timer control panels. This will allow you to track the specific gallons per minute, or gallons

per day being transferred and relate it to overfeed ratio. In an automated control system, excessive

circulation rate can be set as an alarm. Digital counters are now offered as a standard option by

some system manufacturers.

SUMMARY

A Control Pressure Liquid recirculation system may operate with the same energy consumption or less

than a mechanical pump, but not without careful attention to the design of the system, adjustment

of controls and setting of the recirculation rate.

Excessive recirculation rates, high transfer gas pressure, and low volumetric efficiency, particularly at

lower suction temperatures can combine to make a gas pressure recirculation system very inefficient.


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Gas Powered Liquid Recirculation Compared To Mechanical Pumpsby James D. Wright, P.E.


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gas powered liquid recirculation compared to mechanical pumps

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