Liquid pressure amplification with bypass
RELATED APPLICATIONS 
     This is a Continuation-in-Part of an application filed on Mar. 7, 1994 entitled LIQUID PRESSURE AMPLIFICATION WITH SUPERHEAT SUPPRESSION U.S. patent application Ser. No. 08/207,287, now U.S. Pat. No. 5,386,700, which is a divisional of Ser. No. 07/948,300 filed on Sep. 21, 1992 (U.S. Pat. No. 5,291,744), which in turn is a divisional of Ser. No. 07/666,251, filed Mar. 8, 1991, (U.S. Pat. No. 5,150,580). 
    
    
     BACKGROUND AND SUMMARY OF THE INVENTION 
     This invention relates generally to refrigeration and operation and more particularly to a method and apparatus for boosting the cooling capacity and efficiency of air-conditioning systems under a wide range of ambient atmospheric conditions. 
     In air conditioning, the basic circuit is essentially the same as in refrigeration. It comprises an evaporator, a condenser, an expansion valve, and a compressor. This, however, is where the similarity ends. The evaporator and condenser of an air conditioner will generally have less surface area. The temperature difference DT between condensing temperature and ambient temperature is usually 27° F. with a 105° F. minimum condensing temperature, while in refrigeration the difference DT can be from 8° F. to 15° F. with an 86° F. minimum condensing temperature. 
     I have previously improved the cooling capacity and efficiency of refrigeration systems. As disclosed in my U.S. Pat. No. 4,599,873, this is accomplished by addition of a liquid pump at the outlet of the receiver or condenser. Operation of the pump adds 5-12 p.s.i. of pressure to the condensed refrigerant flowing into the expansion valve, a process I call liquid pressure amplification. This suppresses flash gas and assures a uniform flow of liquid refrigerant to the expansion valve, substantially increasing cooling capacity and efficiency. The best results are obtained when such a system is operated with the condenser at moderate ambient temperatures, usually under 80° F. As ambient temperatures rise above the minimum condensing temperature, the advantages gradually decrease. The same thing happens when the principles of my prior invention are applied to air conditioning, except that the minimum condensing temperature is higher. 
     While conventional air-conditioning systems can benefit from my prior invention, the greatest need for air conditioning is when ambient temperatures are high, over 80° F. Conventional air conditioning becomes less effective and efficient as ambient temperatures rise to 100° F. or more, as does use of my prior liquid refrigerant pressure amplification technique. 
     I have since found that in large refrigeration or air conditioning systems, high refrigerant flow rates require multiple pumps in parallel or a larger single pump. The use of a larger single pump is often preferred for simplicity of design. In such systems the large electrically-driven compressors typically operate on a separate electrical circuit from the liquid pressure amplification pump motor. Should the power circuit to the liquid amplification pump motor be turned off or disconnected while the compressor motor circuit is still operable, the compressor will work to drive refrigerant through the pump. At high flow rates, the pressure drop through a centrifugal pump, ordinarily fitted with an output restrictor, will become higher than acceptable. In order to preserve all the available capacity of the partially disabled system under those circumstances, pressure drops in the system must be minimized where ever possible. Unfortunately, it is not possible to entirely eliminate the pressure drop through the idle liquid pressure amplification pump. If a positive displacement pump is used as the liquid pressure amplification pump, in place of the preferred centrifugal pump, the pump can block flow completely when its motor loses power. This, too, is unacceptable. 
     It is, therefore, an object of the invention to improve the efficiency of refrigeration and air-conditioning systems. 
     Another object of the invention is to increase the cooling capacity of such systems when operated at high ambient temperatures. 
     A further object of the invention is to enable the aforementioned objects to be attained economically and by retrofitting existing systems as well as in new systems. 
     A third object of the invention is to minimize the pressure drop imposed on the operating refrigeration or air conditioning system by the liquid pressure amplification pump when idle. 
     The present invention is an improvement in the structure and method of operation of an air-conditioning or refrigeration system which includes a compressor, a condenser, an expansion valve, an evaporator, and conduit means interconnecting the compressor, condenser, expansion valve and evaporator in series in a closed loop for circulating refrigerant therethrough, and optionally may include a receiver between the condenser and expansion valve. The conduit means includes first conduit means coupling an outlet of the compressor to an inlet to the condenser to convey superheated vapor refrigerant from the compressor into the condenser at a first pressure and temperature. A liquid pump means has an inlet coupled to an outlet of the condenser (or to the receiver outlet) for receiving condensed liquid refrigerant at a second pressure less than said first pressure and boosting the second pressure of the condensed liquid refrigerant by a substantially constant increment of pressure within a predetermined range to discharge the condensed liquid refrigerant in a forward direction from an outlet of the pump means at a third pressure greater than said second pressure. A second conduit means couples the outlet of the pump means to an inlet to the expansion valve to transmit a first portion of the condensed liquid refrigerant from outlet of the pump means at said third pressure through the expansion valve into the evaporator to vaporize and effect cooling for air conditioning or refrigeration. A third conduit means couples the outlet of the pump means to an inlet to the condenser to transmit a second portion of the condensed liquid refrigerant from outlet of the pump means into the inlet of the condenser to vaporize therein. The portion of the condensed liquid refrigerant injected into the condenser inlet cools the superheated vapor refrigerant entering the condenser to a reduced temperature, thereby reducing said first pressure. 
     The first and second conduit means are preferably proportioned so that the second portion of refrigerant is sufficient to reduce the first temperature to a reduced temperature close to a saturation temperature of the refrigerant, preferably within 10° F. to 15° F. above saturation temperature, and so that the second portion of refrigerant is substantially less than the first portion, preferably less than about 5% of the first portion and typically in the range of 2%-3% of the first portion. Suitably, the first and second conduit means are proportioned with a cross-sectional area ratio of about 16:1. The system preferably further includes means responsive to a temperature of the evaporator for modulating the expansion valve. 
     The system further includes a bypass conduit connected between the intake and outlet of the liquid pressure amplification pump, and a flow control means in the bypass conduit, through which refrigerant flows in the forward direction responsive to a predetermined pressure differential, and which blocks refrigerant flow in a reverse direction responsive to a reversal of the pressure differential. The flow control means preferably includes a check valve, or can include an electrically operated solenoid valve. 
     In the improved method of operation, superheated vapor refrigerant is transmitted from the compressor to an inlet to the condenser at a first temperature and pressure. The vapor refrigerant is condensed and discharged as liquid refrigerant at a second temperature and pressure less than said first temperature and pressure. The pressure of the liquid refrigerant discharged from the condenser (or receiver) is boosted to a third pressure greater than the second pressure by a substantially constant increment of pressure. Then, in accordance with the invention, a first portion of the liquid refrigerant is transmitted at said third pressure via the expansion valve into the evaporator and a second portion thereof is transmitted into the condenser inlet so that the first temperature of the superheated vapor refrigerant is reduced toward said second temperature, thereby reducing said first pressure. 
     The first and second portions of liquid refrigerant at said third pressure are proportioned so that the first portion is substantially greater than the second portion. Preferably, the added increment of pressure is 8 to 10 p.s.i. and the second portion has a flow rate less than 5% of the flow rate of the first portion. The flow of the first portion through the expansion valve can be modulated in response to a temperature in the evaporator. 
     Prior art ammonia-refrigeration systems are known in which a portion of liquid refrigerant is injected from the receiver to the condenser inlet to suppress superheat. This has not been done, however, in combination with adding an incremental pressure, for example by means of a centrifugal pump, to the pressure of the liquid refrigerant flowing into the expansion valve. 
     Operation with an added incremental liquid refrigerant pressure preferably includes allowing the first pressure to float with an ambient temperature. This reduces overall system pressures, thereby increasing system efficiency at moderate ambient temperatures. The present invention desuperheats the compressed refrigerant vapor as it enters the condenser, lowering its temperature and further reducing the first pressure, even when ambient temperatures are high. The invention thus raises the temperature range over which benefits can be obtained from adding an increment of pressure to the liquid refrigerant. This further improves efficiency and enables effective operation in very high ambient temperature environments. 
     The foregoing and other objects, features and advantages of the invention will become more readily apparent from the following detailed description of a preferred embodiment of the invention which proceeds with reference to the accompanying drawings. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic diagram of a conventional air-conditioning system, with the condenser and evaporator shown in cross section and shaded to indicate regions occupied by liquid refrigerant during condensation and evaporation. 
     FIG. 2 is a view similar to FIG. 1 showing the system as modified to include a liquid pump in accordance with the teachings of my prior patent. 
     FIG. 3 is a graph of certain parameters of operation of the system of FIG. 2 with the liquid pump ON and OFF. 
     FIG. 4 is a view similar to that of FIG. 2 showing the system as further modified for superheat suppression in accordance with the present invention. 
     FIG. 5 is a chart of test results comparing three parameters for each of the systems of FIGS. 1, 2 and 4 operating under like ambient conditions. 
     FIG. 6 is a view similar to that of FIG. 4 showing the system as further modified for bypassing the liquid pressure amplification pump in accordance with the present invention. 
    
    
     DETAILED DESCRIPTION 
     To understand how we can improve the refrigeration cycle we must first analyze the components of a conventional air-conditioning system and understand where the inefficiencies exist. 
     FIG. 1 depicts the conventional air-conditioning circuit 10. The circuit of FIG. 1 consists of the following elements: a compressor 12, condenser 14, expansion valve 16, and evaporator 18 with temperature sensor 20 coupled controllably to the expansion valve, connected in series by conduits 13, 15, 17 to form a closed loop system. Shading indicates that the refrigerant within the condenser passes through three separate states as it is converted back to a liquid form: superheated vapor 22, condensing vapor 24 and subcooled liquid 26. Similarly, shading in the evaporator indicates that the refrigerant contained therein is in two states: vaporizing refrigerant 28 and superheated vapor 30. Pressures and temperatures are indicated at various points in the refrigeration cycle by the variables P1, T1, P2, T2, etc. 
     In the evaporator, only the refrigerant changing from a liquid state 28 (P4, T3) to a vapor state 30 (P4, T4, assuming DP small) provides refrigerating effect. The more liquid refrigerant (state 28) in the evaporator, the higher its cooling capacity and efficiency. The ratio of liquid to vapor refrigerant can vary. The determining factors are the performance of the expansion valve, the proportion of &#34;flash gas&#34; entering the evaporator through the valve, and the temperature T3 and pressure P4 of the entering liquid refrigerant. As can be seen in FIG. 1, only superheated vapor (state 30) enters the compressor 12. The term &#34;superheat&#34; refers to the amount of heat in excess of the latent heat of the vaporized refrigerant, that is, heat which increases its volume and/or pressure. High superheat at the compressor inlet can add considerably to the work that must be performed by other components in the system. Ideally, the vapor entering the compressor would be at saturation, containing no superheat and no liquid refrigerant. In most systems using a reciprocating compressor 12 this is not practical. We can, however, make significant improvements. 
     The discharge heat of the vapor exiting from the compressor includes the superheat of the vapor entering the compressor plus the heat of compression, friction and the motor added by the compressor. At the entrance of the condenser, all of the refrigerant consists of superheated vapors at pressure P1 and temperature T1. The portion of the condenser needed to desuperheat the refrigerant (state 22) is directly related to the temperature T1 of the entering superheat vapors. Only after the superheat is removed can the vapors start to condense (state 24). 
     The superheated vapors 22 are subject to the Gas Laws of Boyle and Charles. At a higher temperature T1, they will tend to either expand (consuming more condenser area) or increase the pressures P1 and P2 in the condenser, or a combination of both. The rejection of heat at this point is vapor-to-vapor, the least effective means of heat transfer. 
     As the vapors enter the condensing portion of the condenser they are at saturation (state 24) and at a pressure P2 and temperature T2 which are not greater than P1 and T1, respectively. At this stage, further removal of latent heat will convert the vapors into the liquid state 26. The pressure P2 will not further change during this stage of the process. 
     As the refrigerant starts to condense, the condensation will take place along the walls of the condenser. At this point, heat transfer is from liquid-to-vapor, and produces a more efficient rejection of unwanted heat. 
     The condensing pressures are influenced by the condensing area (total condenser area minus the area used for desuperheating and the area used for subcooling). The effect of superheat can be observed as both a reduction in condensing area (state 24) and an increase in the overall pressure (both P1 and P2). 
     In an effort to suppress the formation of flash gas entering the expansion valve, many manufacturers use part of the condenser to further cool or subcool the liquid refrigerant to a lower temperature T3 (state 26). If we consider only the subcooling of the liquid without regard to decreased condensing surface, then we can expect a gain of 1/2% refrigeration capacity per degree (F.) of subcooling. If we consider the reduction in condensing surface, however, then there is a net loss of capacity and efficiency due to increased condensing temperature T2 and higher head pressure P1. 
     Analysis of the refrigeration cycle shows that several factors that can be improved. Combining these factors, as described with reference to FIG. 4, can dramatically improve the overall capacity and efficiency of performance. 
     FIG. 2 illustrates, in an air-conditioning system, the effects of liquid pumping as taught in my prior U.S. Pat. No. 4,599,873, incorporated herein by reference. The system is largely the same as that of FIG. 1, so like reference numerals are used on like parts. The various states are indicated by like reference numerals followed by the letter &#34;A.&#34; Temperatures and pressures are also indicated in like manner with the understanding that the quantities symbolized by the variables differ substantially in each system. 
     The principal structural difference is that a liquid refrigerant centrifugal pump 32 is installed between the outlet of the condenser 14 (on systems that do not have a receiver) and the expansion valve 16. The pump 32 increases the pressure P2 of the liquid refrigerant flowing from the condenser outlet by a DP of 8 to 15 p.s.i. to an incrementally increased pressure P3. This is referred to as the liquid pressure amplification process. The pressure added to the liquid refrigerant will transfer the refrigerant to the subcooled region of the enthalpy (i.e., P3&gt;P2, T3 same, and will not allow the refrigerant to flash prematurely, regardless of head pressure. By eliminating the need to maintain the standard head pressure, minimum head pressure P1 can be lowered to 30 p.s.i. above evaporator pressure P4 in air-conditioning and refrigeration systems. Condensing temperature T1 can float rather than being set to a fixed minimum temperature in a conventional system, e.g., 105° F. in R-22 air-conditioning systems. If ambient temperature is 65° F., using a pump 32 in an R-22 air-conditioning system lowers condensing temperature T1 to about 86° F. at full load. Additionally, head pressure P1 is lowered, as next explained. 
     For the evaporator 18 to operate at peak efficiency it must operate with as high a liquid-to-vapor ratio as possible. To accomplish this, the expansion valve 16 must allow refrigerant to enter the evaporator at the same rate that it evaporates. Overfeeding or underfeeding of the expansion valve will dramatically affect the efficiency of the evaporator. Using pump 32 assures an adequate feed of liquid refrigerant to valve 16 so that the exhaust refrigerant at the intake of compressor 12 is at a temperature T4 and pressure P4 closer to saturation. 
     FIG. 3 graphs the flow rate of refrigerant through the expansion valve 16 in laboratory tests with and without the liquid pump 32 running. The upper trace indicates incremental pressure added by pump 32 and the lower trace graphs the flow rate of refrigerant through the expansion valve. The test begins with the system running in steady state with centrifugal pump 32 ON. At 131 min. the pump was turned OFF. The flow rate of refrigerant entering the evaporator 18 through the expansion valve 16 (TXV) shows a decided decrease in flow compared to the flow when the pump is running. An increase in head pressure only partially restores refrigerant flows. The reduced flow of refrigerant to the evaporator has several detrimental effects, as shown in FIG. 1. Note the reduced effective evaporator area 28 as compared to area 28A in FIG. 2. 
     At 150 min., the liquid pump 32 is turned ON. With the pump 32 again running, the flow rate is consistently higher, with an even modulation of the expansion valve, because of the absence of flash gas. It can be seen that running the pump increases the amount of refrigerant in the evaporator yet the superheat setting of the valve controls the modulation of the expansion valve at a consistent flow rate. The net result is a greater utilization of the evaporator 18 as shown in FIG. 2 (note state 28A). 
     The efficiency of the compressor 12 is related to a number of factors, most of which can be improved when the liquid pumping system is applied. The efficiencies can be improved by reducing the temperature in the cylinders of the compressor, by increasing the pressure P4 of the entering vapor, and by reducing the pressure P1 of the exiting vapor. With the vapor entering the compressor at a higher pressure, the compressor capacity will increase. With cooler gas (T4) entering the cylinders, the heat retained in the compressor walls will be less, thereby reducing the expansion, due to heat absorption, of the entering vapor. 
     With these improvements on the suction side of the compressor, the condensing temperature T1 can float with the ambient to a lower condensing temperature in the system of FIG. 2. This reduces the lift, or work, of the compressor by reducing the difference between P4 and P1. The increased capacity or power reduction, due to the lower condensing temperatures, will be approximately 1.3% for each degree F. that the condensing temperature is lowered. As explained earlier, the liquid pump&#39;s added pressure DP maintains all liquid leaving the pump 32 in the subcooled region of the enthalpy diagram. For this reason, it is no longer necessary to flood the bottom part of the condenser (See 26 in FIG. 1) to subcool the refrigerant. This portion of the condenser can now be used to condense vapor (Compare state 24A of FIG. 2 with state 24 in FIG. 1). This increased condensing surface can further lower the condensing temperature T2 and pressure P2. The temperature T3 of the refrigerant leaving the condenser will be approximately the same as if subcooled, but with little or no subcooling (state 26A). 
     With the application of the pump 32, the evaporator discharge or superheat temperature T4 and compressor intake pressure P4 have been reduced considerably from the corresponding parameters in the system of FIG. 1. 
     The best results are obtained when such a system is operated with the condenser at moderate ambient temperatures, usually under 80° F. As ambient temperatures rise above the minimum condensing temperature, the advantages gradually decrease. At a typical ambient temperature of around 75° F., a typical improvement in efficiency of the system of FIG. 2 over that of FIG. 1 is 7%-10%, declining to negligible at 100° F. ambient temperature. 
     I have discovered, however, that, by using the present invention, next described, an additional 6% to 8% savings can be achieved under typical ambient conditions. Moreover, we can obtain very substantial improvements of efficiency and effectiveness at ambient temperatures over 100° F. 
     FIG. 4 shows an air-conditioning system 100 in accordance with the present invention. The general configuration of the system resembles that of system 10A in FIG. 2. In accordance with the invention, however, a conduit or line 34 is connected at one end to the outlet of pump 32 and at the opposite end to an injection coupling 36 at the entrance to the condenser. This circuitry enables a portion of the condensed liquid refrigerant to be injected at temperature T3 from the pump outlet into the entrance of condenser. As this liquid refrigerant enters the desuperheating portion of the condenser, it will immediately reduce the temperature of, and thereby suppress, the superheated vapors entering the condenser at pressure P1 and temperature T1. 
     The amount of refrigerant injected at coupling 36 should be sufficient to dissipate the superheated vapors and preferably reduce the incoming temperature T1 to a temperature close (within 10° F.-15° F.) to the saturation temperature T2 of the refrigerant. In a 10 ton, 120,000 BTU air-conditioning system, line 15 has an inside diameter of 1/2 inch and line 34 has an inside diameter of 1/8 inch, for a cross-sectional ratio of line 34 to line 15 of 1:16 or about 6%. Due to flow rate differences and variations (e.g., due to modulation of valve 16 by sensor 20) the flow ratio is less than about 5%, probably 2%-3%, in a typical application. 
     Suppression of superheated vapor will have four effects: 
     (1) By reducing the superheat temperature T1, the pressure P1 and volume of the superheat vapors will both be reduced. 
     (2) The vapor will be very close to or at saturation point (T2, P2). 
     (3) Condensing will occur closer to the inlet of the condenser. 
     (4) Heat transfer will be higher because of liquid-to-vapor heat transfer over a greater area of the condenser (compare state 24B with state 24A). 
     The injection of liquid refrigerant into the condenser 14 is accomplished using the same pump 32 that is installed for the liquid pressure amplification process. By reducing the work required to desuperheat the refrigerant vapor, the pump can perform a substantial portion of the work required to recirculate the liquid through the condenser. Although some gain can be seen at low ambient temperature, with this process of superheat suppression, the best gains will be realized at higher ambient temperature. This is just the opposite of the benefits noted with liquid refrigerant amplification alone. For example, at over 100° F., the system of FIG. 2 gives little if any increase in efficiency and capacity over the system of FIG. 1. Tests have shown that the system of FIG. 4, on the other hand, will provide efficiency increases of 10%-12% at 100° F. and as much as 20% at 113° F., and add capacity to allow air conditioning to be run effectively in the desert. 
     FIG. 5 is a graph of actual results achieved in a test of a 60 ton Trane air-conditioning system comparing operation of system 100 of FIG. 4 with operation of systems 10 and 10A of respective FIGS. 1 and 2. All readings were taken at 86° F. ambient temperature. The readings are: A. standard system without modification (FIG. 1), B. same system adding the pump 32 only (FIG. 2), and C. the same system modified in accordance with the present invention to include both pump 32 and superheat suppression circuitry 34, 36 (FIG. 4). For each parameter--head pressure P1 (p.s.i.), condensing temperature T1 (°F.) and liquid temperature T3 (°F.) entering the evaporator--configuration C, the present invention, demonstrated lower readings. Such performance characteristics enable a system 100 according to the present invention to provide a greater cooling capacity as well as greater efficiency. These advantages continue to higher ambient temperatures, including temperatures at which configurations A and B would no longer be effective. 
     FIG. 6 shows an alternative embodiment including bypass conduits 50, 52 connected around liquid amplification pump 32, and valve 54 to control refrigerant flow through bypass conduits 50 and 52. I have discovered that the high refrigerant flow rates of large refrigeration or air conditioning systems necessitate multiple liquid pressure amplification pumps in parallel or a larger single liquid pressure amplification pump. The use of a larger single pump is often preferred for simplicity of design. In such systems the large electrically-driven compressors typically operate on a separate electrical circuit from the liquid pressure amplification pump motor. Should the power circuit to the liquid amplification pump motor be turned off or disconnected while the compressor motor circuit is still operable, the compressor will work to drive refrigerant through the pump. In order to preserve all available cooling capacity of the partially disabled system under those circumstances, unnecessary refrigerant pressure drops in the system should be minimized where possible. Unfortunately, it is not possible to entirely eliminate the pressure drop through the idle liquid pressure amplification pump. In the case of an idle centrifugal pump, the convoluted flow path through the idle pump, along with the throttling of the pump outlet required to minimize cavitation, together cause a pressure drop through the idle pump which cannot be eliminated. In the case of an idle positive displacement pump, refrigerant flow is likely be blocked entirely, other than seepage of fluid through clearances within the pump. I have solved this problem by providing bypass conduits 50 and 52 around pump 32, which is preferably a centrifugal pump but which could alternatively be a positive displacement pump. 
     Refrigerant flow through bypass conduit 50 and 52 is controlled by valve 54 (FIG. 6). In one embodiment, valve 54 is a check valve of standard design, such a swing check valve, a lift check valve, or a tilting-disk check valve, which remains closed during normal system operation to prevent backflow of refrigerant around pump 32. In an alternate embodiment, valve 54 can be an electrically operable valve, such as a solenoid-actuated valve which is spring-biased to a normally open position to permit flow through the bypass conduit, and electrically biased to a closed position, from the pump motor circuit. Whenever power is removed from the pump motor, the power to the solenoid is turned off, allowing the valve to move to its normally open position to open the bypass line. In yet another embodiment, valve 54 can be a solenoid-actuated valve in which the power is turned off to open the valve responsive to a loss of pressure downstream of pump 32. 
     In each of the foregoing instances, if pump 32 is idled while the compressor continues to operate, valve 54 opens permitting refrigerant to bypass pump 32 in a forward, i.e. downstream, direction and limits the pressure drop to less than about 5 psi, and preferably to 1/2 to 1 psi. When pump 32 is restarted and downstream pressure increases, valve 54 closes again to prevent backflow. 
     Having described and illustrated the principles of the invention in a preferred embodiment thereof, it should be apparent that the invention can be modified in arrangement and detail without departing from such principles. I claim all modifications and variation coming within the spirit and scope of the following claims.