CA2172843A1

CA2172843A1 - Apparatus for maximizing air conditioning and/or refrigeration system efficiency

Info

Publication number: CA2172843A1
Application number: CA 2172843
Authority: CA
Inventors: Marc D. Sandofsky; David F. Ward
Original assignee: Individual
Current assignee: JDM Ltd
Priority date: 1993-09-28
Filing date: 1994-09-28
Publication date: 1995-04-06
Also published as: WO1995009335A3; JPH09506162A; AU7962594A; WO1995009335A2

Abstract

The invention entails the use of a positive displacement pump (4) magnetically coupled to a drive motor (42) located in a conduit arrangement (60) this is parallel to the liquid line of the refrigeration system as in the figure. This parallel conduit arrangement also includes a pressure regulating valve that will regulate the amount of pressure added to the liquid line by the parallel pump and piping arrangement In addition a check valve (47) is located in the liquid line to maintain the pressure differential added to the liquid line. This parallel piping arrangement (60) is desirable in order to allow a constant, predetermined pressure to be added to the liquid line regardless of variations in flow rate of the liquid refrigerant. In addition, the parallel piping arrangement allows the system to operate without liquid line obstruction in the event of pump failure. Also a/any method of introducing a controlled, fine mist of water into the entering air stream of air cooled condensers to simulate the performance of an evaporative condenser by reducing the entering air temperature from dry bulb temperature to wet bulb temperature resulting in a reduction in system condensing temperature/pressure. The pump of the invention consists of an outer driving magnet (200), a stationary cup (201), and an O-ring seal (202). The pump further includes an inner driven magnet (203), a rotor assembly (204) and vanes (205). The pump further includes an O-ring seal (206) and a brass head (207).

Description

APPARATUS FOR MAXIMIZING AIR CONDITIONINC~
AND/OR REFRIGERATION SYSTEM EFFICIENCY

1. Field of the Invention This invention generally relates to the field of mechanical refrigeration and air S conditioning and more particularly to improving efficiency of colllpression-type refrigeration and air conditioning systems.

2. Background of the Invention In the operation of commercial freezel~, refrigel~tol~, air conditioners and other colllpression-type refrigeration systems, it is desirable to m~ximi7e refrigeration capacity while minimi7.ing total energy con~ulllplion. Specifically, it is nPcçs~ry to operate the systems at as low a collll)lession ratio as possible without the loss of capacity that normally occurs when colllplessor colllplession ratios are reduce 1 This is accomplished by ~uppressillg the formation of "flash gas". Flash gas is the spontaneous fl~hing or boiling of liquid refrigerant res hing from pressure losses in refrigeration system liquid refrigerant lines. Various techniques have been developed to eli...i-~te flash gas. However, collvellLional m~-fhotlc for ~up~lessillg flash gas can subst~nti~lly reduce system efficiency by increasing energy co~ ion.

Fig. 1 represents a conventional m~ch~nic~l refrigeration system of the type - typically used in a supermarket freezer. Specifically, col.lples~or 10 colllpresses refrigerant vapor and discharges it through line 20 into con~len~er 11. Con~1en~er 11 con~len~es the refrigerant vapors to the liquid state by removing heat aided by circnl~ting fan 31. The liquid refrigerant next flows through line 21 into receiver 12.
From receiver 12, the liquid refrigerant flows through line 22 to counter-flow heat exc}l~nger (not shown). After passing through exchanger 13, the refrigerant flows via line 23 through thermostatic exran~ion valve 14. Valve 14 ~xp~n~ls the liquid refrigerant to a lower pres~ule liquid which flows into and through evaporator 15 S~ E S~E~ (RU! E 26) WO95/09335 2 ~ 7 2 8 4 3 PCT/US9-1/11116 1--where it evaporates back into a vapor absorbing heat. Valve 14 is c- nn~cted to bulb 16 by capillary tube, 30. Bulb 16 throttles valve 14 to regulate te~ elaLules produced in evaporator 15 by the flow of the refrigerant. Passing through evaporator 15, the e~p~nded refrigerant absorbs heat relul ..;ng to the vapor state aided by circulating fan 32. The refrigerant vapor then returns to co~ ,lessor 10 through line 24.

In order to keep the refrigerant in a liquid state in the liquid line, the refrigerant pr~s~ure is typically m~int~ined at a high level by keeping the refrigerant temperature at condenser 11 at a i.~ini---~"" of approximately 95o F. This minim~lm conde.n~ing ~elllpel~lulc; ",~inli.i,lc pressure levels in receiver 12 and thus the liquid lines 22 and 23 above the flash or boiling point of the refrigerant. At 95 F.
condencing l~lllpe~ l~, this pl`eS~Illt; for example would be; 125 PSI for refrigerant R12, 185 PSI for refrigerant R22 and 185 PSI for refrigerant ~502. These temperature and ples~ul~ levels are snfflcient to ~upples~ flash gas formation in lines 22 and 23 but the coll~/elllional means of m~int~ining such levels by use of high colllpressor discharge pleS~Illes limits system efficiency.

Various means are used to m~int~in the telll~lalure and pressure levels stated above. For example, Fig. 1 shows a fan unit 31 connected to sensor 17 in line 21.
Controlled by sensor 17, fan unit 31 is r~;s~ol~si~e to contl~ncP.r telllpel~lure or plt;s~ul~ and cycles on and off to regulate condenser heat ~licsir~tion. A plc;s~ule le~pollsi~/e bypass valve 18 in con~1encer output line 21 is also used to m~int~in pr~ure levels in receiver 12. Normally, valve 18 is set to enable a free flow ofrefrigerant from line 21a into line 21b. When the pl~iSult:; at the output line of condencer 11 drops below a predetermined minimllm, valve 18 operates to permit co~ re~sed refrigerant vapors from line 20 to flow through bypass line 20a into line 21b. The addition to vapor from line 20 into line 21b increases the pres~ure in receiver 12, line 22 and line 23, thereby ~u~pres~ih~g flash gas.

The foregoing system elimin~tes flash gas, but is energy inefficient. First, m~int~ining a 95~F. con~-n~er telllpel~lule limits colll~ressor capacity and increases energy consu~ Lion. Although the 95 F. ~ el~Lule level m~int~in~ sufficient WO 95/09335 217 2 ~ ~ 3 3 PCT/US94/11116 IJleS~ul`e to avoid flash gas, the r~s~ nt elevated pressure in the system produces a back ~les~ule in the conden~er which increases compressor work load. The operation of bypass valve 18 also increases back pl~S~Ulc in the conden~er. In addition, the release of hot, co,ll~)lc~sed vapor from line 20 into line 21 by valve 18 increases the S refrigerant specific heat in the receiver. The added heat necessitates yet a higher pressure to control flash gas formation and reduces the cooling capacity of the refrigerant, both of which reduce efflciency.

Another approach to ~u~Jles~ing flash gas has been to cool the liquid refrigerant to a ~ ;lalule sulJ~ y below its boiling point. As shown in Fig.
1, a subcooler unit 40 has been used in line ~2 for this purpose. However, subcooler unitsrequireadditional m~cl.i~.GJy andpower, increasing e4ui~lllentandoperating cost and reducing oveMll operating efficiency.

Other methods for controlling the operation of refrigeration ~y~l~uls are disclosed in US. Pat. Nos. 3,742,726 to F.ngli~h, 4,068,494 to Kramer, 3,589,140to Osborne and 3,988,9()4 to Ross. For example, Ross liscloses the use of an extra colll~lessor to increase the pressure of gaseous refrigerant in the system. The high plc;S~ure gaseous refrigerant is then used to force liquid refrigerant through various parts of the system. However, each of these ~y~le~ls is complex and requires extensive purchases of new equipment to retrofit t~ ting systems. The expenses involved in the purchase and operation of these methods usually outweigh the savings in power costs.

A more recent method of controlling the formation of flash gas in the iiquid line was disclosed in US Pat. No. 4,599,873 by R. Hyde. This method involves theuse of a m~gnetic~lly coupled cellllifugal pump placed in the liquid line as seen in Fig. 2. Fig. 2 shows a vapor line 114, a condenser 116, a fan unit 118, a liquid line 120, a receiver 122, a pump 124 and 125, a liquid line 126, a heat exchanger 128, aliquidline 129, avalve 130, aline 131, acontrol 132, anevaporator 134, afanunit138, and a vapor line 140. The purpose of this method is to illl~)l.ve system effiçi-~nf,y by allowing system condensing ~l~ssules and temp~ldlul~s to be reduced WO 95/09335 PCT/US9 1/11116 ~
2172~3 as ambient tempel~lwes reduce. The cellllifugal pump 124 adds pl~;s~ule to the liquid line 126 at the point where the liquid line exits from the Condens(~.r 116 or receiver 122 without the use of colllplessor horsepower. This method of using a ce~ irugal pump to add plt;S~UIt; reduces the amount of flash gas that forms in the liquid line, S but does not ~-,limin~te it altogether.

Purthermore, e~ ;on of the centrifugal pump curve in Fig. 3 shows that as flow increases, the ples~ule added by the c~nl~irugal pump decreases. However, as flow of refrigerant liquid through the liquid line increases the pres~ul~ drop in the liquid line increases by the square of the velocity. This colllbination of effects as shown in Fig. 4. causes the cellllirugal pump to only reduce the formation of flash gas during certain low flow con-lition~, below point A in Fig. 4. As refrigerant flow increases at high load Con~liti(m~ and the l,res~ule added by the centrifugal pump decreases, the formation of flash gas begins to increase again and system capacity is lost when it is needed most.

Ahother deficiency of the previously described centrifugal pumping method is that the centrifugal pump is located within the liquid line itself. If the c~llllirugal pump fails to operate L,ropelly for any reason, it becollles an obstruction to flow of refrigerant liquid seriously i...lu;.;,~E the operation of the refrigeration system.

The most serious deficiency of the previously described centrifugal ~ulllping method however, is caused by the state of the refrigerant at the outlet of the condenser 116 or receiver 122. The liquid refrigerant at this location in ~e system is commonly at or very near the saturation point. Any vapor ~at forms at the inlet of the cellllirugal pump due to incomplete conden~tion or slight drop in pr~ssule caused by the pump suction or any other reason will cause the cellL.i[ugal pump to cavitate or vapor lock and lose prime. This renders the cellllirugal pump ineffective until the system is stopped and restarted again, and is very detrim~nt~l to pump life and reliability. Due to the constantly varying conditions of operation of the refrigeration system this can occur with great regularity.

Wo gs/09335 2 1 7 2 8 ~ 3 5 PCT/US94/11116 A further development pertaining the fields of mechanical air con-litinning and refrigeration relating to system o~l;",i~i~lion is ~ close~l by U.S. Pat. No. 5,150,580 also by R. Hyde. This development, seen in Fig. 2., involves the transfer of some small amount of liquid refrigerant from the outlet of the ce~ irugal pump 124 in the S liquid line 126 to be injected via conduit 136 into the collll"~ssor discharge line 114 by means of the added prt;s~ure of the cellLIirugal pump 124 in the liquid line. The purpose of injection this liquid into the discharge line is to desuperheat the compressor discharge vapors before they reach the conden~çr to reduce condenser pressure and thereby reduce the co",p,~ssor discharge ples~u,e. This development is said to improve system efficiency at high ambient te",pe,~lu,es when air conditioning systems work the hardest and system ples~u,es are the highest Again, however, as system p,~s~ures increase and refrigerant flow rates increase at higher loads, the increased flow rate of reffigerant causes more y~es~ùl~;
loss through the con-l~.ns~.r. However, this same increased flow rate causes less ples~ule to be added to the liquid by the cel~l~irugal pump 124 in the liquid line 126.
Thus, less liquid is by~assed via conduit 136 into the colllplessor discharge line and less superheat is ~li",in~l~l at the time when more reduction is needed. And at some point the lJres~ule loss through the condenser is greater than the p,es~u,~ added by the cel,LIirugal pump and the effect is lost entirely.

Obviously, there lel"aills a need to provide a stable pl~ iUIe increase in the liquid line 126 to completely t~li---i-~5~1e the forrnation of flash gas, and likewise a stable ~,es~u,e increase in the liquid injection line 136 to completely desuperheat the co"~ essor discha,ge vapors if the i",l.,oYe,l-ent in system efficiency is to be realized on a COQ~I~ull and reliable basis regardless of system configuration or refrigerant flow rate or vapor content.

In addition, since the energy conmmrtion of the compressor reduces with a reduction in con-le.n~ing ~"~pe~lu~e, and it is the object of the aforementi~nedtechnology to operate the system at as low a con~l~n~ing telllpt;l~lule/discharge plessult; as possible, a further need exists to reduce con~1en~ing ~ el~lule~, 3!!i PCT/US9-1/11116 ~72~3 6 particularly during times of high ambient tempeMtures. This has previously been achieved by using a water cooled conc~Pn~-r in place of an air cooled condenser.Water cooled conden~Prs are much more expensive and weigh more than air cooled condensers when used as original equipment. The use of water cooled condensers also involves water tr~oAtmçnt and disposal concerns. These weight, cost and water quality concerns are even more restricting when it is desired to exrhAnge air cooled con-~çn~ers with water cooled condensers. There is a need then to develop a simple and effective way to modify air cooled cont~en~çrs into water cooled condensers.
The objectives of the present invention are to:

1) Reliably and con~tAntly increase the ~ ;S:jUIe in the liquid line to ~Up~lt;Ss the formation of flash gas without l)nnece~Arily IllAi~ \g a high system pressure, and without the possibility of obstructing the flow of refrigerant through the liquid line.

2) To reliably and constantly inject the correct amount of liquid into the co""),essor discl~e line to IllAx;llli~P the heat L.~n~rel in the condenser.

3). To illlploYe the operating efficiency of coll,pression-type refrigeration and air con~li1~oning ~y~Lellls in a con~L~ul, controlled and reliable basis regardless of system configuration or refrigerant flow rate.

4). To ,.l~xil-.i,e the refrigeration capacity of refrigeration and air conditioning systems in a con~L~Ill, controlled and reliable basis regardless of system configuration or refrigerant flow rate.

5). To economically and con~LA~ y suyyless the formation of flash gas in refrigeration and air conditioning systems without illlpAil ;.~g refrigeration capacity and efficiency regardless of system configuration or refrigerant flow rate.

WO 95/09335 PCT/US9~/11116 ~72~3 6). To economically and reliably modify air cooled cQnd~on~ers to achieve the benefits of water cooled con-le.n~çrs without the high weight, cost and water quality problems, thus reducing the con~on~ing telllpeldlure/compressor discharge pressure during periods of high ambient lelll~e,dlules for the purpose of extenlling the benefits of low head pres~ e operation.

7). To provide a way to inexpensively retrofit existing refrigeration systems to attain the foregoing objects on a reliable and controllable basis regardless of the system configuration or refrigerant flow rate.

8). To provide a method of dulo,~,AIic~lly reducing the flow rate of the pulll~ g ~pald~US to match the refrigerant flow rate in large refrigeration or air cl n~itioning systems that have some unloading capability to match the load.

9)Further, the previous objects must be met in a way that will not be detrim~.n~l to the system in the event of failure of the installed pumping mechanism or conden~er cooling mech~ni~m 9). Still further, the above objects must be reliably met regardless of the presence of some vapor in the liquid at the inlet of the pumping arrangement since the liquid is at or near saturation.

10). Moreover, the above objects must be met in a way that can be adjusted to satisfy a majority of the wide range of system configurations found in the field.

This invention provides for the refrigeration or air conditioning system to be operated in a way which "~ ";~.e~ energy efficiency and ~u~)p-~sses flash gas formation regardless of system configuration or refrigerant flow rate.

This invention further provides for the modification of air cooled c-)nrlen~ers to achieve the benefits of water cooled con~en~çrs without the weight, cost and water quality co~ normally ~so~ d with water cooled conden~e.rs.

WO 95/09335 ............. . ' . PCT/US9~/11116 ~
~728~ 8 The foregoing and other objects, features, and advantages of the invention will become more readily apparellt from the following description of a preferred embodiment, which proceeds with reference to the figures.

3. S~mm~ry of the Invention S The invention entails the use of a positive displacement pump magnetically coupled to a drive motor located in a conduit arrangement that is parallel to the liquid line of the refrigeration system as in Fig. S This parallel conduit arrangement also inc~ e$ a pl`t;S`7ule reg~ ting valve that will regulate the amount of pl~7Ult; added to the liquid line by the parallel pump and piping arrangement. In addition, a check valve is located in the liquid line to ~ the ple,sule diflerelllial added to theliquid line. This parallel piping arrangement is desirable in order to allow a constant, pre-d~L~ ed pres,ule to be added to the liquid line regardless of variations in flow rate of the liquid refrigerant. In addition, the parallel piping arrangement allows the system to operate without liquid line obstruction in the event of pump failure.

Fur~er, a pump is added to ~e liquid injection line that is connect~ between the liquid line and the collll~ressor discharge line for the purpose of desuperheating the compressor dischal~e vapors. This pump insures a constant flow of liquid refrigerant to the coluyl~ssor disch~ge line to fully de-,u~ellleat the collll)ressor discllarge vapors.
The preferred method would entail the use of a positive displacement pump, but any suitable ç,ulllyillg method can be used.

Also, the above pump can be controlled by a variable speed drive mech~nism The variable speed drive mechanism is controlled by two (2) temperature sensors. One temperature sensor is located on the condenser to sense ~ led temperature of therefrigerant in the condenser. The other l~n~pelalule sensor is located at the inlet of the condenser d~wlls~ of the point of liquid injection into the compressor dischargeIine to sense amount of superheat in the discharge line. The speed of the pump located in the liquid injection line is varied by the ~tt~checl variable speed drive by means of the sensed Lt;lllpel~Lure differelltial to provide just the proper amount of wo gs/09335 ~17 2 8 ~ 3 PCT/US94/11116 liquid injection into the discharge line to adequately desuperheat the compressor discharge vapors for optimum heat transfer in the c5)nden~er regardless of the refrigerant flow rate or amount of superheat present in the co~l~ylessor discharge vapors.

S In addition, this invention entails the use of a precisely controlled ultrasonic fogger to produce a regulated fine spray of water that evaporates quickly as it enters the condenser entering air stream. The fine mist of water is controlled to reduce the condenser entering air temperature by evaporation of the mist, and to wet the surface of the condenser coil without producing any excess that would run off the conden~er.
The purpose of the mist is to lower the condenser entering air temperature to the wet bulb tellly~ ule lller~y reducing the con-ien~ing temperature/l,les~ule of the system.

4. Description of the Drawings Figure 1 is a schem~tic diagram of a typical refrigeration system, as previouslydescribed.

lS Figure 2 is a schematic diagram of a refrigeration system including the prior art as previously described. including the liquid injection for de~ul)ellle~ g.

Figure 3 is a diagram of a typical ce,lllirugal pump curve showing yles~ul~ added vs.
flow rate.

Figure 4 is a diagram of yleS:,ule loss through a piping system vs. flow rate with the cellllirugal pump curve superimposed over it.

Figure S is a schem~tic diagram of a refrigeration system inchlding the present invenhon.

WO 95/09335 217 2 8 ~-~ PCT/US91/11116 1--Figure 6 is a more ~Pt~ilecl diagram of the parallel piping arrangement with positive displ~cernent pump, ~les~ule differential regulating valve and check valves of the present invention.

Figure 7 is a more det~ile~ diagram of the preferred method of adding yle~sule to the liquid injection line including tne optional ~,~;re,-ed control method.

Figure 8 is a diagram of the duplex l~u~lping arrangement used to match ch~ngin~refrigerant flow rate in larger systems with unloading capabilities.

Figure 9 is a diagram of the ultrasonic fogger arrangement in the condenser entering air stream of a previously air cooled condenser.

Figure 10 is a blown up depiction of a l.rek~ d embodiment of the pump(s) of thepresent invention.

5. Detailed D~scli~lion of the Preferred Embodime~t Referring now to Fig 5, a closed circuit co~ ession-type refrigeration system includes a colllplessor 10, a condenser 11, an optional receiver 12, an expansion valve 14 and an evaporator 15 connPctPcl in series by conduits defining a closed-loop refrigerant circuit. Refrigerant gas is compressed by compressor unit 10, and routed through discharge line 20 into conden~er 11. A fan 31 facilitates heat rli~siration from condenser 11. Another fan 32 aids evaporation of the liquid refrigerant in evaporator 15. The colllpressor 10 receives warm refrigerant vapor at yles~,ule Pl and colll~lesses and raises its ~res~uie to a higher ple~sure P2. The condenser cools the col~ ;ss~d refrigerant gases and con~ n~es the gases to a liquid at a reduced ~les~ule P3. From condenser 11, the liquefied refrige~nt flows through line 21 into receiver 12 in cases where there is Cu~ y a receiver in the system. If there is no receiver in the system the con~3en~eA refrigerant flows directly into the liquid line 22.Receiver 12 in turn discharges liquid refrigerant into liquid line æ.

WO 95/09335 ~17 2 8 ~ 3 PCT/US9~11 1 1 16 Figure 6 shows a positive displ~ement pump 41, driven by electric motor 42 m~gn~tic~lly coupled to the pump head is positioned in conduit arr~ngemPnt 60 parallel to the liquid line 22 at the outlet of the receiver or con(len~er to p~ uli~e the liquid refrigerant in the line to an increased pressure P4. This parallel piping- S arrangement 60 also inclu~es the pl~;S~iUle differential regulating valve 45 and a check valve 46 arranged as shown in Fig. 6 to provide for a constant added pressure (P4 -P3) regardless of refrigerant flow rate or vapor content. A check valve 47 is added to the liquid line 22 to m~int~in the ples~ule differential between line 22 and line 23 (see FIG. 7). An adjustable ~les~ure regulating valve 45 can also be used to more accurately match the ples~ul~; differential required or to f~cilit~te changes that may be needed in the ~ S~iUIe dirr~r~ al added. The yres~ule dirrele~llial of the regulating valve 45 (FIG. 6) determines the amount of ples~ule that is added to the system. Dirrert;lll amounts of pres~ule can be added to the liquid line 22 as neceS~ry for each difrelel~l system configuration by using dirrel~ l yl~ule dirrel~lllial valves or by adjusting the valve to a specific ples~ule as n-~lecl. As the flow rate of the system varies in conduit 22, more or less refrigerant flows through parallel conduit 22a (FIG. 6) and yres~ule regulating valve 45 so the refrigerant flow into and out of the parallel piping all~lgelllent 60 always m~t~hçs the flow rate through conduit 22 and 23 and the pres~ul~ differential (P4 - P3) remains comlz...l From parallel piping all~lgel~ent 60, the liquid refrigerant flows into the liquid line 23 (FIG. 7). Some of the liquid refrigerant flows through conduit 25 and through pump 43 into c~ lessor discharge line 20 to ~e~-.pe.l-t-~t the co,llylt;~sor discharge vapor. Pr~rel~bly pump 43 would be a yOSiliVt; displ~cemt~.nt pump controlled by variable speed drive 44. The speed of the pump is d~lerl,lined by the tempeMture dirrt;l~lllial between the con~-n~ing l~lllpel~lule of the refrigerant in contl~.n~er 11 as sensed by bulb 49 and the t~;lllpel~lule of the superheated refrigerant in line 20 as sensed by bulb 48.

The rem~in(l~r of the liquid refrigeMnt from the paMllel piping arrangement 60 flows llllou~ll the line and through an optional c~,ulller-current heat exch~n~er (not shown) to thermost~tic exp~n~ion valve 14. Thermostatic t~xp~n~ion valve 14 exp~n(l~

WO 95/09335 PCT/US9~/11116 1--~28~3 the liquid refrigerant into evaporator 15 and reduces the refrigerant pres~u,e to near P1. Refrigerant flow through valve 14 is controlled by temperature sensing bulb 16 positioned in line 24 at the output of evaporator 15. A capillary tube 30 connects sensing bulb 16 to valve 14 to control the rate of refrigerant flow through valve 14 S to match the load at the evaporator 15. The expanded refrigerant passes through evaporator 15 which, aided by fan 32, absorbs heat from the area being cooled. The rxr~nd~A, warmed vapor is returned at ples~u-t; Pl through line 24 to colllp,essor 10, and the cycle is repeated.

Pump 41 and pressul~e regulating piping arrangement 60 is preferably located as close to leceiver 12 or the outlet of condenser 11 as possible, and may be easily in~t~ll~ in existing systems without extensive purchases of new equipment. Pump 41 must be of sufflcient capacity to increase liquid refrigerant pressure P3 by whatever p,es~u,e is necec~-y to elimin~te the formation of flash gas in the liquid line 23 (FIG. 7). The pump must also be capable of adding a Coh~l~ull plt;s~ule to the liquid line regardless of the presence of some vapor in the incoming liquid refrigerant in line 22. A positive displ~rement pump and pressure regulating valve located in a parallel piping arrangement 60 most effectively, economically and reliably provides this capability.

Pump 41 must also be capable of adding a co~ l pressure to the liquid line undercon-lition~ of variable refrigerant discharge rates from valve 14, including conditions in which valve 14 is closed.

In systems where the refrigerant flow rate varies signifi~ntly~ the ~ùnlpillg arrangement must be able to vary its flow rate by a similar amount. In these cases, a duplex pumping arrangement, Figure 8, is used. The duplex pumping arrangement 2~ consists of two pumps piped in parallel each with either a single speed, two speed or variable speed motor and a control mech~ni~m capable of adjusting the speed of one or both pumps to match the flow rate of the refrigerant in the refrigeration circuit.
This duplex pumping arrangement is typically used in systems that have multiple c.,lllyressors or coll,pre~ors with the capability of llnlo~ling to signifir~ntly reduce I Wo ss/o933s ~ 1 7 2 S ~ 3 PCT I Sg4/~ 6 the refrigerant flow rate The duplex pumping arrangement controls tie into the system controls to adjust the pump or pumps speed to match the compressor loading thereby - ~c11ing the refrigerant flow rate During the higher outside ambient temperature conditions, the ultrasonic fogger S al)l)dldlus as shown in figure 9 would be activated by an adjustable l~11,pe1dture control set to ambient te""~)elatu,e The fogger would reduce the condenser entering air temperature of the conventional air cooled conden~r from the dry bulb temperature to the wet bulb temperature to sim~ te the performance of an evapo,dLi~re con-len~er. This reduction in entering air temperature would result in a loweredconden~ing ~e111pe1dlurt;/~les~u1e, thus reducing co.--p.essor energy co1~u-11~lion during higher outside ambient conditions Operation Referring to Fig 5, co",p,essor 10 co",piesses the refrigerant vapor which then passes through discharge line 20 to con-len~er ll. In the con-~enser ll, atpressure P2, heat is removed and the vapor is liquefied by use of ambient air orwater flow across the heat exchange.. At contlen~er ll, l~---pe-dlu-e and p.~ssll-e levels are allowed to fl~ ct 1~t~ with ambient air tempe-dlu,c;s in an air-cooled system, or with water te.. pe~dlures in a water-cooled system to a .. in;.. 1 conden~ing pressure/le..1pe,~lu,~e that has previously been set at about 95 P. This previously set minim11rn contlen~ing temperature has been neces~ y to prevent the formation of flash gas in the liquid line 22. The previously set n.ini.. n. was n.~;nl;.in~l by reducing air or water flow across the heat exchanger of conci~n~er l l to reduce heat transfer from the con~enser. Further decreasing the con~ n~ing te,--~e,~lu-es increase system efficiency in two ways: l) The lower ~)-es~u-e dir~.~1.Lial of the compressor lOincreases the col--E)-essor volumetric efficiency according to the formula Vc= 1 +C-C*(VI/V2) where V~ is volumetric efficiency, C is the cl~ nce ratio of the compressor, Vl is the specific volume of the refrigerant vapor at the beginning of co,..~.ession~ V2 is the specific volume of the refrigerant vapor at the end of WO 95/09335 PCT/US91tllll6 ~
~728~3 .

col"p,ession, and 2) The lower liquid refrigerant temperature at the outlet of the condenser results in a greater cooling effect in the evaporator.

The negative effect of reducing condensing temperatures below this previously set ..-i-~i...~.,.. has been the formation of flash gas in the liquid line 23 (FIG. 7), which when passed through expansion valve 14 reduced the net refrigeration effect of the evaporator 15. The net result was a reduction of energy consumption per unit time by the compressor, but a ~im~llt~neous reduction capacity of the system causing an increase in compressor run time resulting in no net energy savings.

When the refrigeration or air conditioning system is morlified with the present invention as in Fig. 5, the .. ini.. --.. conden~ing temperature and pressure can be reduced .~ignific~ntly without the loss of capacity mentioned above due to the pressure added to the liquid line by the pump 41 and parallel piping arrangement. As the ambient air e"lpe,alu,e or water temperature used to cool the condenser becomes lower, the efficiency of the compressor improves, and the capacity of the evaporator increases, since no flash gas has been allowed to form in the liquid line. This is most beneficial with refrigeration systems that operate year around and can take advantage of the cooler ambient te",pel~tures.

As ambient air ~ell,pe,alure or cooling water telllpelaLule increases the con(len~ing lel"pe,~lule and prt;s~u,e of the refrigeration or air con(litioning system also increases and efficiency is reduced. In order to im~luve efficiency at these higher ambient con-lition~ when air conditioning and refrigeration systems are at or near maximum capacity, liquid refrigerant is bypassed from the liquid line 23 (Fig.
7) into the colllplessor discharge line 20. Since there is some amount of p,es~u,e lost as the refrigerant passes through the condenser 11, making cnnden~r exit pressure P3 lower than e"l~ ce pies~ule P2, a pump is needed to add enough pressule to insure flow of liquid from the liquid line 23 intû the discharge line 20. The preferredmethod is to use a positive displ~cem~nt pump, driven by a variable speed drive,controlled by the lelllpelalul~ differential between the superhP~ted colllplessûr discharge vapor tel~til~lure T2 and the conden~ing lel,lpe,alule T3. As the ~ WO 95/09335 PCT/US9~/11116 21~2~4~

temperature differential becomes greater, the variable speed drive would cause the positive di~pl~rement pump to pump more liquid into the discharge line 20 to decrease the superheat. When the superheat LG~ eldlule and the conden~inE
temperature were the same, the refrigerant vapor entering the c~ n-l-on~çr would be at 5 the saturation point and the speed of the positive displ~rem~nt pump would stabilize to a pre-set speed to ~ the condition.

This method of superheat suppression insures that the refrigeMnt vapor is entering the condenser at saturation res~ ing in the o~ulilllulll conditions for heat ll~rel thereby ol)!i---i~i.-g the efficiency of the condçn~er. This portion of the invention is most beneficial at higher ambient temperature.

In addition, a method of further i"lpruving air cooled con(~en~er efflciency at higher ambient tempe,dLulGs is by the use of the previously described ultrasonic fogger a~dlus to reduce the condenser entering air IGIll~eldlulG from dry bulb to wet bulb temperature. The ultrasonic fogger appdldlus would dispense a controlled fog of 10 micron sized water particles into the entering air stream of the previously air cooled con-l~n~er to sim~ te the performance of an G~oldlivG cnndçn~çr. The amount of this fog would be varied with the condenser load, entering air temperature and/or wet bulb depression to just dispense enough fog to saluldte the entering air stream. This would achieve the benefits of reducing the elllGrillg air ~GlllpGldlulG from the dry bulb lelllpeldture to the wet bulb temperature without the need for other water pumps, drain pans, sumps, water tre~tmçnt or blowdown ~;ullGIllly associated with evaporative con~l~n~çrs.

Referring to Figure 10, the pump(s) of the present invention con~istc of an outer driving magnet 200, a stationary cup 201, and an O-ring seal 202. The pumpfurther includes an inner driven magnet 203, a rotor assembly 204 and vanes 205.The pump further includes an O-ring seal 206 and brass head 207.

Taken together, both parts of the invention i"lpro~e system performance and ~ffleit~ncy over the full range of operating contlition~ and lG"lpG,dtules.

WO9~/09335 2~ 728~3 PCT/US9~/11116 ~

The use of m~gnetically-coupled rotary-vane pumps as positive displacement pumps for pumping refrigerants has been found to be startlingly effective and they have been found to exhibit a surprisingly long life. Once the vanes are worn to the extent that they are properly seated and sealed, subsequent wear is almost negligible.
S This discovery has resulted in very effective use of these m~gn~tic~lly-coupled rotary-vane pumps as positive displacement pumps for ~u-l,phlg refrigerants in non-compressor-type refrigeration cycles. This application is particularly effective when a compressor-type refrigeration cycle (preferably with the help of the present invention) is used to store refrigeration, for exarnple, in the form of ice, during low energy cost periods and then the col~ s~or is turned off during peak energy costperiods. During the peak period, the magnetically-coupled rotary-vane pump of the present invention (ideally the same pump used to increase the efficiency of the co",l~re~sor cycle) is used to circulate the same refrigerant through the ice, through the same conduits, and through the same cooling coils (ev~o.ator), to cool the conditioned space during peak energy cost periods.

Another aspect of the present invention is the use of starting torque control means for the positive displacement pump. Typically, when a positive displacement pump is placed into the liquid line of an air conditioning or refrigeration system, the electric motor driving it is energized when the compressor is energiæd. This creates two problems when the pump head is full of refrigerant upon start-up, as it is normally the case. First, excessive torque is required to bring the pump head up to speed while it is adding pressure to the liquid. Second, the rapid acceleration of the pump rotor will cause temporary, but signific7/nt, cavitation that may damage the pump.

The solution to both problems is to ramp the motor and pump up to operating speed slowly. This can be accomplished by using a device called a "soft starter".
This device will bring the motor up to full speed over a period of 1 or more seconds, depending on its design.

WO 95109335 PCT/US9-1/llllG
~ ~172813 Upon normal start-up, a standard electric motor will go from O R.P.M. to its full speed of 3450 R.P.M. in less than 1 second. This causes excessive torque re~luire-lle lls and cavitation when such a motor is coupled to a positive displacement pump that is full of a liquid near saturation. If the acceleration rate of the motor and pump head is slowed down so that it comes up to speed in preferably between 2 and 8 seconds, for example, the excessive torque and cavitation problems are avoided.

The variation in start-up acceleration can be accomplished by several means:
1. using an induction coil in series with the electric motor, or 2. redecigning the motor windings to give less start-up torque, therefore slower starting speed, or 3.
in~t~lling a separate "soft start" electronic component to a standard motor that varies the voltage to the motor.

Having described and illustrated the principles of the invention in a preferred embodiment thereof, is should be apparent that the invention can be modified slightly in arrangement and detail without departing from such principles. In that regard, this patent covers all modific~1ion~ and variations falling within the spirit and scope of the following claims:

Claims

1. Any refrigeration, air conditioning or process cooling system using a reciprocating screw, scroll, centrifugal or other similar type of compressor and any type of refrigerant, the improvement including a first positive-displacement pump used in a parallel piping arrangement which arrangement is parallel to a conventional conduit between a condenser and an expansion valve, and parallel with a differential pressure regulating valve and a check valve.

2. A system as recited in claim 1, wherein the system includes a second pump in a liquid injection line between the output of the first pump and the output of a compressor, used for desuperheating the compressor dischargevapor, a control mechanism that controls the speed of the second pump and thereby results in the desuperheating of the compressor discharge vapor to a saturated or near saturated condition at the inlet to the condenser, said control mechanism including a temperature sensor adapted to sense the temperature of the refrigerant at the condenser.

3. A system as recited in claim 2, wherein the system includes a control system which sets the minimum condensing temperature setting of refrigerant exiting the condenser to a lower-than-conventional value when the first pump is functioning properly and reverts the air conditioning or refrigeration system back to the higher minimum condensing temperature setting in case of failure of the first pump.

4. A vapor-compression heat transfer system having fluid refrigerant, a compressor, a condenser, an expansion valve, an evaporator, a refrigerant conduit between the condenser and the expansion valve, and a refrigerant pump in the conduit adapted to increase the pressure of the refrigerant between the condenser and the expansion valve, the improvement comprising (a) the fact that the said pump is a positive displacement pump, and (b) a first bypass conduit is provided in parallel around the pump, said first bypass conduit including a differential pressure regulating valve which imposes an upper limit on the pressure increase caused by the pump, and (c) a second bypass conduit is provided in parallel around the pump, said second bypass conduit including a check valve adapted to stop flow of refrigerant through the said second bypass conduit from the expansion valve to the condenser, but to allow flow of refrigerant through the said second bypass conduit from the condenser to the expansion valve, and (d) said pump, and bypass conduits being adapted to increase the said pressure of the refrigerant sufficiently to avoid the formation of refrigerant flash gas in said conduit between the pump and the expansion valve, while still allowing flow of refrigerant from the condenser to the expansion valve if the pump fails to operate.

5. A vapor-compression heat transfer system as recited in claim 4, wherein a liquid injector conduit is provided between an output side of the pump to an output side of the compressor, and adapted to deliver pressurized liquid refrigerant from the output of the pump to the output of the compressor to de-superheat the refrigerant which exits the compressor.

6. A vapor-compression heat transfer system as recited in claim 5, wherein the liquid injector conduit includes a variable-speed injector pump, and a control system is provided and adapted to monitor the difference in temperature of the refrigerant going into the condenser and within the condenser and to adjust the speed of the injector pump to minimize the difference in temperature, which in turn minimizes superheat in the refrigerant going into the condenser and, in turn, maximizes the efficiency of the condenser.

7. A vapor-compression heat transfer system as recited in claim 4, wherein a control system is provided to cause reduction in the minimum condensing temperature at the outlet of the condenser when the pump is effectively reducing flash gas, but the control system is adapted to raise the minium condensing temperature to a point which reduces flash gas, if the pump fails to operate.

8. A compression type refrigeration system, comprising:
an evaporator, a compressor, a condenser, a refrigerant receiver and conduit means interconnecting the same in a single closed loop for circulating refrigerant therethrough, the conduit means including;
a first conduit for circulating a flow of refrigerant from the receiver to the evaporator and;
a second conduit for circulating a return flow of refrigerant gas from the evaporator to the receiver solely through the compressor and the condenser for condensation by the condenser at a first pressure directly related to the head pressure at the compressor;
a variable flow expansion valve in the first conduit adjacent the evaporator for expanding the flow of refrigerant into the evaporator;
a third conduit which provides a parallel path around a section of said first conduit adjacent an outlet port of the receiver;
a positive displacement pump in the third conduit adjacent the receiver, the pump being adapted, continuously during operation of the compressor, to increase the pressure of the condensed refrigerant in the first conduit by a generally constant increment of pressure of at least five pounds per square inch to provide the refrigerant with a second pressure greater than the first pressure by the amount of said increment, the second pressure being sufficient to suppress flash gas and feed a completelycondensed liquid refrigerant to the expansion valve, the first conduit circulating the refrigerant solely through the pump;
motor means for the pump; and a magnetic pump drive connecting the motor means to the pump to drive the pump.

9. A system as recited in claim 8, which includes a fourth conduit which provides a parallel path around the said section of said first conduit, and includes a pressure regulating valve, in said fourth conduit, said pressure regulating valve being adapted to regulate the amount of pressure added to the first conduit by the pump.

10. A system as recited in claim 9, which includes check valve in said section of said first conduit, said check valve being adapted to maintain the pressure differential added to the first conduit by the pump while allowing full and uninterrupted flow of refrigerant in the event of pump failure.

11. A system as recited in claim 8, which includes check valve in said section of said first conduit, said check valve being adapted to maintain the pressure differential added to the first conduit by the pump while allowing full and uninterrupted flow of refrigerant in the event of pump failure.

12. A system as recited in claim 8, which includes a pressure regulating valve in a by-pass around the pump to control the effect of the pump, and a check-valve in a by-pass around the pump to allow refrigerant flow if the pump fails.

13. A system as recited in claim 8, which includes a liquid injector conduit between (a) the first conduit after said section, and (b) a point in said second conduit between the compressor and the condenser, said liquid injector conduit including a variable speed pump, the speed of which is controlled by a first temperature sensor adapted to sense the temperature of the refrigerant in the condenser, and a second temperature sensor adapted to sense the temperature of the refrigerant going into the condenser, the speed of the variable speed pump being controlled by said temperature sensors so that just the proper amount of liquid refrigerant is injected into the second conduit at a point after the compressor to desuperheat the compressor discharge refrigerant for optimum heat transfer in the condenser regardless of the refrigerant flow rate through the condenser and regardless of the amount of superheat present in the compressor discharge refrigerant.

14. A method of introducing a controlled, fine mist of water into the entering air stream of air cooled condensers to simulate the performance of an evaporative condenser by reducing the entering air temperature from dry bulb temperature to wet bulb temperature resulting in the reduction in system condensing temperature/pressure.

15. The system of claim 1 wherein the first pump consists of an outer driving magnet 200, a stationary cup 201, an O-ring seal 202, an inner driven magnet 203, a rotor assembly 204, vanes 205, an O-ring seal 206, and a brass head 207.

16. The system of claim 2 wherein the second pump consists of an outer driving magnet 200, a stationary cup 201, an O-ring seal 202, an inner driven magnet 203, a rotor assembly 204, vanes 205, an O-ring seal 206, and a brass head 207.