Patent Publication Number: US-RE39625-E

Title: Boosted air source heat pump

Description:
BACKGROUND AND CROSS REFERENCE 
     This invention relates to air-source heat pumps. More particularly, it relates to new and improved air source heat pumps especially suitable for use in normally colder climates. This invention is an improvement on my U.S. Pat. No. 5,927,088, the entire contents of which are incorporated herein by reference. 
     The following discussion of typical prior art heat pumps refers to air-source heat pumps other than the heat pumps disclosed in my U.S. Pat. No. 5,927,088. 
     The air-source heat pump system is the most prevalent type of heat pump used in the world today. This is the case whether one is discussing room units, residential central type, ductless splits, or rooftop commercial systems. 
     Although the air-source concept in general has a high application potential worldwide, its popularity in the United States and elsewhere has been greatest in mild climate areas. This is because the compressor-derived heating capacity of typical prior art units declines rapidly as the outdoor ambient falls, due, in most part, to the large increase in specific volume (i.e., decrease in density) of the outdoor coil generated refrigerant vapor as the ambient (outdoor) temperature falls. This fall in compressor-derived heating capacity is obviously opposite to the heating requirement, which increases as the outdoor ambient temperature falls. When a typical prior art heat pump operates below its balance point (about 35° F.-40° F.), supplemental heating is required. The most prevalent form of supplemental heat used is electric resistance. In other than mild climates, this use of supplemental electric resistance heat puts the air-source heat pump at a serious economic disadvantage to a consumer as compared with other forms of heating (such as natural gas, oil, and propane), because of the high cost of electric resistance heating. 
     Electric utilities are also very concerned because of the associated large transformers and distribution systems that are required for any large populations of typical prior art heat pumps whenever high electric resistant (KW) heat backup is required on a regular basis for large population. 
     When operating at low outdoor ambient temperatures such as 0° F., homes heated by typical vapor art heat pumps require as much as KW input from the utility as does a home heated by electric resistance KW alone. This is not acceptable to the utility as they would have to increase their generating capacity to supply the demand. In other words, a Northern utility that was summer peaking would now become winter peaking because much more KW output is required for electrically heated Northern homes than what is required for cooling those same homes. 
     As discussed in my U.S. Pat. No. 5,927,088, one of the areas for capacity and efficiency improvement of air source heat pump systems lies in the recovery of significant heat energy currently remaining in the condensed liquid refrigerant leaving the system condenser. If this remaining energy is recovered and returned directly to the heating side of the system before being thermally degraded and sent to the system evaporator as low density vapor (as is now the case in present day systems), significant increases in compressor derived heating capacity and C.O.P. can be made at lower outdoor ambient temperatures. 
     The basic problem here is that after the refrigerant has been fully liquefied in the heating condenser, there is still a large amount of energy left in the leaving warm liquid. This remaining energy evaporates a large portion of the heating liquid itself during the normal pressure reduction process that is required to develop the necessarily low evaporating temperatures. Depending on the refrigerant utilized, and the degree of temperature existing between the evaporator and the condenser, as much as one-half of this liquid can be evaporated during this normal pressure reduction process across the system expansion device when operating at the lower outdoor ambient temperatures. 
     Obviously, if this liquid has already evaporated, it cannot be again evaporated in the system evaporator, and thus cannot absorb energy from the outside air. However, the net resulting vapor must pass through the system evaporator anyway, creating additional pressure drop along its way, and then must be induced and fully compressed to the condensing level by the compressor, thus requiring the necessary power to accomplish the compression. The C.O.P. of this portion of the heating process is only one (1) because no energy has been absorbed from the outside air. 
     Since the compressor must induct this previously evaporated vapor, the compressor can only induct a correspondingly smaller amount of vapor that has been derived from the cooling of outside air (by evaporating the refrigerant liquid that does enter the evaporator along with the previously mentioned vapor). This is not a reasonable process for typical prior art air-source heat pumps operating in other than the milder ambient temperatures because only under those conditions is the relative amount of liquid to vapor (by weight) sufficient to result in a good system C.O.P. 
     The heating energy output of any heat pump system is also closely proportional to the weight flow of refrigerant vapor entering the system condenser. Approximately 4 times the amount of heat energy is required of 0° F. than is required at 50° F. This means that approximately a 400% increase of entering condenser refrigerant vapor is required at 0° F. ambient as compared to 50° F. ambient in order to adequately match the heating energy requirement. However, the density of the refrigerant vapor generated in the system evaporator when operating at 0° F. ambient is only about 32% of that generated when the outdoor temperature is 50° F. Therefore, when approximately four (4) times the weight flow is required when only 32% of the vapor density is generated, it becomes very obvious that significant changes must be made in order to make an air source pump viable for colder Northern climates. 
     In addition, if the entire space heating requirement at 0° F. outdoor ambient is to be supplied by compressor derived heating capacity, the air flow across the heating coil of the condenser must be such that the indoor delivered air temperature will be around 105° F. in order to provide adequate freedom from a sensation of cool drafts. This in turn will cause the system condensing temperature to rise to around 115° F. considering a reasonably sized indoor coil surface. 
     The end result of all this (even if the necessary compressor displacement could be obtained somehow with a present day single compression stage system) is to cause overall system operating compression ratios to rise to the point where it becomes unrealistic to continue the use of typical prior art heat pump technology in normally colder climates. This is exactly what has happened in the marketplace of today. Air source heat pumps are no longer purchased for use in cold climate areas for reasons of both poor comfort as well as the very high cost of the electric energy requirement. 
     SUMMARY OF THE INVENTION 
     In this invention, a system and method are presented that achieve a great increase in refrigerant pumping capacity as related to large fall in the outdoor ambient temperature, combined with a method of extracting energy from warm liquid leaving the system condenser where this is needed as well. The heat pump system itself is made to be sufficiently flexing in order to accomplish the necessary goal of both sufficient and efficient heating for a wide range of outdoor ambient temperatures. 
     In order for anti-source heat pumps to become serious contenders for use in colder climates, significant changes must be made for them to realize their true potential. Fundamental Carnot Theory thermodynamic principles unquestionably show that electric powered air source heat pumps indeed to have significant potential in cold climates. In fact, the theoretical Carnot C.O.P. (coefficient of performance) limit for a sink (room) temperature of 70° F. and a source (outside air) temperature of 0° F. is 7.57 units of energy delivered to the sink for every 1 net unit of energy supplied to the compression process. 
     Carnot C.O.P.=(T 2 ΔS)+(T 2 −T 1 )ΔS where T 2  is the delivered energy sink temperature (room temperature in absolute degrees) and T 1  is the supplied energy source temperature (outside air temperature in absolute degrees) and ΔS is the constant change in entropy for this theoretical cycle. Therefore, executing the equation, the ΔS&#39;s cancel out and the final equation is simply Carnot C.O.P.=T 2 +(T 2 −T 1 ) where only T 2  need be in absolute degrees. 
     It is expected that with this present invention, an actual heating C.O.P. of at least 2.0 will be reached at this 0° F. outdoor ambient condition. This represents a Carnot efficiency level of only [(2.0+7.5)×100]=26%, which clearly is within the bound of rational achievability. Accordingly, the system and method of this invention allows the basic electric utility service size to be only about 50% of that which is required for today&#39;s heat pumps selected for cold climates. 
     In typical prior art systems, the actual delivered C.O.P. is only about 1.0 at the condition of 70° F. room temperature and 0° F. outside air temperature, because most of their delivered energy comes from electric resistance coils which obviously operate with a C.O.P. of 1.0 (1 unit of delivered energy for every 1 unit of supplied energy). Also, with these typical prior art systems, the delivered energy that does come from the refrigeration circuit may even come with a C.O.P. of less than 1 at these low outdoor ambient temperatures because of a significant percentage energy loss between the outdoor unit and the indoor condenser when operating under low refrigerant flow rate condition. Further to this problem, most residential type compressors also operate rather inefficiency at this 0° F. condition. Some system manufacturers even shut down their compressors at temperatures of 0° F. and rely entirely on electric resistance for a variety of different reasons as well. 
     Typical Embodiment of a New System 
     An embodiment of the present invention is directed to a refrigeration circuit which comprises one first stage booster compressor, one second stage primary compressor (always the first to run), a condenser, an economizer, an evaporator, and conduit means bearing a compressible refrigerant working fluid and connecting the first stage compressor, the second stage compressor, the condenser, the economizer, and the evaporator, in series and in that order, in a closed loop. An important feature of the present invention is a novel apparatus and method for oil equalization between the primary and booster compressors. 
     The closed loop further comprises a bleed line for bleeding a portion of the condensed refrigerant from the closed loop downstream of the heating condenser and expanding it within the economizer for highly subcooling the liquid refrigerant within the closed loop being fed to the evaporator. The expanded refrigerant from the economizer is then delivered to a point between the outlet of the first stage compressor and the inlet to the second stage compressor. The subcooling of the liquid refrigerant in the economizer greatly increases the ability of the refrigerant to absorb heat energy in the evaporator. Also, the vapor created by this subcooling process significantly increases the refrigerant weight flow into the heating condenser as it is directly added to the flow coming form the first stage compressor. 
     In accordance with this invention, at least the primary compressor is a multi-cylinder unloadable compressor, such as a Bristol Twin-Single, or similar type compressor (TS), preferably a 40/100 TS or a 50/100 TS compressor. A Bristol twin-single compressor has two cylinders and pistons. In one direction of rotation of the drive shaft, both cylinders/piston are operating (full capacity operation). In the reverse direction of rotation of the drive shaft, only one piston/cylinder is operative, and the other piston/cylinder is idle (partial capacity operation). A TS compressor is a preferred type of unloadable positive displacement compressor for use in this invention. In a preferred embodiment, the booster compressor is a one-speed compressor (although a two speed or variable speed compressor could also be used). In another embodiment, both the primary compressor and the booster compressor are Bristol BS compressors. 
     While the invention will be described as using the Bristol twin-single compressor, it will be understood that other steps of unloadable positive displacement compressors or other unloadable multi-cylinder positive displacement compressors could be used. For example, using the designation of 2/1 for a Bristol twin-single compressor, a 4/2 multi-cylinder compressor (i.e., four cylinders for fully capacity/two cylinders for partial capacity) or a 6/3 multi-cylinder compressor (six cylinders for full capacity/three cylinders for partial capacity), etc. could be used. 
     An important point related to maximum system pumping capacity is that the first stage (booster) compressor has a larger displacement (by about 10% to about 50%) than the second stage (primar) compressor. 
     The discharge pressure of the booster will rise to the point where the density (pounds per cubic foot) of the vapor entering the primary compressor times the primary compressor pumping capacity (in cubic feet per minute) exactly equals the pounds per minute of vapor existing the booster compressor plus the pounds per minute of vapor exiting the economizer. The increased displacement of the booster (compared to the primary) along with a very high volumetric efficiency of (because of the low booster discharge pressure) results in a very high booster flow rate. This very high refrigerant flow rate multiplied by the increased energy pickup per pound of refrigerant flowing through the evaporator (because of the low liquid refrigerant temperature entering the evaporator due to the economizer), results in a very large increase in the total amount of energy per minute absorbed from the outside air into the operating system. This increase also comes about when it is most needed, i.e., at the lower outdoor ambient air temperatures. 
     A typical control system for the present invention includes a transducer for directly sensing the outdoor ambient temperature, preventing excess system capacity (through utilization of a micro-processor) until the outdoor temperature reaches a predetermined low enough value to allow or enable more system capacity, if called for by the indoor thermoset. The control system (on heating) also responds to a preferred, three step indoor thermoset which will step the system heating capacity to various levels (upon indoor temperature demand) that, in turn, are allowed or enabled by the various outdoor ambient temperature ranges that are encountered. The control system, on cooling, also responds to the same indoor thermoset, which has two cooling steps, which will step the system cooling capacity to various levels (upon indoor temperature demand) that, in turn, are allowed or enabled by the various outdoor ambient temperature ranges that are encountered. 
     The preferred displacement ratio of the booster to the primary (at 100% primary compressor flow) is only about 1.3 to 1. This keeps the pressure ratio across the booster relatively low (whenever it needs to be low), thus resulting in higher booster volumetric efficiencies. This displacement ratio thus keeps the economizer boiling temperature low (again, whenever needed), thus allowing the extraction of as much energy as possible from the warm liquid leaving the system condenser. These facts combined with the utilization of Refrigerant R-410A keep the total displacement required very low considering the heating capacity level that is obtained during low ambient heating. It also allows utilization of relatively low cost smaller reciprocating or other type compressors. 
     The above-discussed and other features and advantages of the present invention will be appreciated and understood by those skilled in the art from the following detailed description and drawings. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic diagram of a heating mode operation of a closed loop boosted air source heating pump of the present invention. 
         FIG. 2  is a schematic diagram of a cooling mode operation of a closed loop boosted air source heat pump of the present invention. 
         FIG. 3  is a schematic showing of the shaft, lobe and pistons of a Bristol Twin-Single compressor. 
         FIG. 4  is a schematic diagram of lubricant management in accordance with the present invention when only the primary compressor is operational. 
         FIG. 5  is a schematic diagram of lubricant management in accordance with the present invention when both the primary and booster compressors are operational. 
         FIG. 6  is a heating capacity chart for a typical system of the present invention. 
         FIG. 7  is a chart showing a typical operating sequence in the heating mode for a system of the present invention. 
         FIG. 8  is a cooling capacity chart for a typical system of the present invention. 
         FIG. 9  is a chart showing a typical operating sequence in the cooling mode for a system of the present invention. 
     
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     Referring to  FIGS. 1 and 2 , there is shown a closed loop heat pump system forming an embodiment of the present invention. Referring first to  FIG. 1 , the closed loop system includes a first or booster stage compressor  22 , a second or high stage primarily compressor  24 , an indoor coil or condenser  26  which delivers heated air to a space to be heated, an economizer  28 , and an outdoor coil or evaporator  30  which, together with conduit means interconnecting these elements in a closed loop circuit, are basic components of the closed loop heat pump system. High stage or primary compressor  24  is normally operating whenever the heat pump system is delivering energy, but booster compressor  22  and economizer  28  are operated only when operation is allowed by the control system depending on outdoor ambient temperature. Warm output vapor of the primary compressor  24  is fed to the inlet of indoor coil  26  via 4 way valve  80  and conduit segment  32 , thus heating air flowing over indoor coil  26  for delivery to the indoor space to be heated. A variable speed fan  27  normally controls the flow of air over indoor coil  20 . The warm refrigerant vapor is, of course, cooled and condensed in indoor coil  26 . The outlet of indoor coil  26  delivers the condensed refrigerant to flow via conduit segment  34  and check valve  35  to the economizer  28 . A bypass or bleed line  38  may permit a portion of the liquid refrigerant to be bled from the primary closed loop circuit and to expand via an expansion valve  40  within economizer  28 . However, expansion valve  40  is normally closed, and it is opened upon receipt of an operating signal from a microprocessor  54  to allow operation of economizer  28 . With expansion valve  40  in its normally closed position, the refrigerant passes directly through economizer  28 , but without any economizer action or effect, to conduit segment  42 , and then through expansion valve  76  to outdoor coil or evaporator  30 . A fixed or variable speed fan  31  delivers the air flow over outdoor coil  30 . The refrigerant then flows from evaporator  30  through conduit segment  46 a, and through four-way valve  80  to conduit segments  46 b and  46 c, and then to the inlet to primary compressor  24 . 
     When operation of the booster compressor  22  is allowed by microprocessor  54 , the refrigerant flows from conduit section  46 b to the inlet to booster compressor  22 , and then via conduit  48  from the discharge from booster  22  to the inlet to primary compressor  24 . 
     Each of components  22  and  24  has its own internal motor, indicated at  23  and  25 , respectively. Each motor is connected to microprocessor  54 , and the operation of each compressor is allowed only by the presence of an activating signal from microprocessor  54  to the compressor motor to operate the compressor. 
     The system also includes a temperature transducer  56 , such as a thermistor, at outdoor cell  30  to sense the temperature of the outdoor air flowing over outdoor coil  30 , a temperature transducer  58 , such as thermistor, at indoor coil  26  to sense the temperature of air leaving indoor coil  26 , and an indoor thermostat  62  which senses the temperature of the air in the space to be heated and sends signals to microprocessor  54  when heat is required or when the desired temperature has been attained. Transducers  56  and  58  and thermostat  62  are connected to deliver signals to microprocessor  54 . The signals received at microprocessors  54  from outdoor ambient temperature transducer  54  are used to allow various combinations of operation of the primarily compressor  24 , booster compressor  22  and economizer  28  as a function of outdoor ambient temperature to meet heating requirements; and the signals received at microprocessor  54  from transducer  58  are used to reduce the speed of fan  27  to reduce or eliminate the “cold blow” problem common to heat pump systems (by reducing air flow at higher ambient temperatures). 
     In the present invention, thermostat  62  has three operating points or stage for heat operation and two operating points or stages for cooling operation. A thermostat of this type may, be e.g., a Minneapolis Honeywell thermostat type T8611M2005 or T8511M1002, available from Minneapolis Honeywell. 
     In the embodiment now being discussed of the present invention, both primary compressor  24  and booster compressor  22  are Twin-Single compressors available from Bristol Compressors of Bristol, Va. Referring to  FIG. 3 , the Bristol Twin-Single is a compressor reciprocating compressor having two pistons  202  and  204 , mounted on a shaft  206 . Shaft  206  can be rotated either clockwise or counterclockwise. Rotatable eccentric lobe  208  is also mounted on shaft  206 . When shaft  206  is rotating, one of the pistons, e.g., piston  202 , is always reciprocating, regardless of the direction of rotation of shaft  106 . When shaft  206  is rotating in the direction shown in  FIG. 3A , lobe  208  is positioned off center of the axis of shaft  206  so that piston  204  also reciprocates (along with piston  202 ). However, when shaft  206  is rotated in the opposite direction, as indicated in  FIG. 3B , lobe  208  is repositioned on the center axis of shaft  206 , whereby piston  204  is idle, i.e., does not reciprocate, and only piston  202  reciprocates. 
     It will be understood that the cylinders in which the pistons  202  and  204  reciprocate are not shown. 
     The fluid flow capacity of a Bristol TS compressor can be split, i.e., allocated, as desired between the two pistons/cylinders. For example, the capacity can be split between 40%/100% to 50%/100% or somewhat larger ratios, where 100% is the flow capacity when both cylinders are reciprocating, and where the lower number is the percentage of total flow capacity when only one piston is reciprocating. For the present invention, a 50/100 split or a 40/100 split is preferred. A split of 40/100 provides adequate heating capacity at higher ambient temperatures where operation of only one cylinder of the primary compressor is required. 
     In a second embodiment of the present invention, only primary compressor  24  is a Bristol TS of similar type compressor. In this second embodiment, booster compressor  22  is either a two speed compressor, or a single speed (fixed displacement) compressor, the latter resulting in the lowest manufacturing cost for the system. In this case, the booster compressor can be any type of positive displacement compressor. 
     Also, is preferred that the compressors be sized so that (1) 100% of the capacity of primary compressor  24  be equal to the rated capacity normally required for cooling by Air Conditioning and Refrigerant Institute (ARI) standards, and (2) the displacement ratio of booster compressor  22  to primary compressor  24  be in the range of 1.1:1-1.7:1, preferably about 1.3:1. 
     Operating sequences for heating and cooling will be set forth and discussed below. At various points in the operating sequences, valve  40  will be operated by a signal from microprocessor  54  to bleed and expand refrigerant fluid from point  36  in line  34  to economizer  28 . The expansion of the refrigerant in economizer  28  results in significant subcooling of the main body of liquid refrigerant which flows in a closed conduit through economizer  28 . This subcooled liquid refrigerant then passes directly to evaporator  30  via conduit segment  42 . This highly subcooled liquid refrigerant expands via expansion valve  44  into and within the evaporator  30  to perform the function of absorbing energy from the outside air flowing over outdoor coil  30  and vaporizing in evaporator  30 . The amount of energy absorbed within evaporator  30  is greatly increased because of the highly subcooled refrigerant delivered from economizer  28  to the evaporator. The refrigerant vapor from evaporator  30  then flows via conduit segment  46 a,  46 b and  46 c and check valve  47  to point  52  and via conduit segment  48  to the suction or low side of primary compressor  24  to complete the closed loop circulation in effect when only the primary compressor  24  is operating, or to the suction or low side of booster compressor  22  if both compressors are operating. 
     Meanwhile, the refrigerant bled via line  38  which vaporizes within the economizer to perform the cooling effect in the economizer, passes via conduit segment  50  to point  52  in conduit  48  connected to the inlet of the primary compressor  24 . 
     The heating mode of operating of the heat pump system is shown in  FIG. 1 , and the cooling mode of operation is shown in FIG.  2 . In the heat mode operation of  FIG. 1 , the refrigerant flows through the closed loop conduit in the direction shown by the arrows in the conduit. To change from the heating mode of  FIG. 1  to the cooling mode of  FIG. 2 , four way valve  80  is operated, as by a mode selection signal from thermoset  62  or microprocessor  54 , whereby the direction of refrigerant flow in the closed loop conduit is reversed, as indicated by the arrows in FIG.  2 . In the cooling mode, indoor coil  26  functions as an evaporator, and outdoor coil  30  functions as a condenser. 
     Lubricant, e.g., oil, management is an important aspect of the present invention. With two compressors connected to series, a potential exists for most or all of the lubricant in the system to accumulate in the sump of one of the compressors, and for the other compressor to become starved for lubricant. That, of course, can lead to failure of the lubricant-starved compressor. The present invention addresses and solves this problem. 
     Reference is made to  FIG. 4 and 5 , which are side schematic views of a compressor module housing the booster and primary compressors  22 ,  24 . Parts in  FIGS. 4 and 5  are numbered as in FIG.  1 . 
     As shown in  FIG. 4 , only the primary compressor is operational. Each of compressors  22  and  24  has a reservoir of oil, respectively  104  and  106  in the sump of each compressor. The compressors also have aspiration tubes  108 ,  110 , respectively, from the sump to the cylinder intake. The tubes  108 ,  110  operate to prevent accumulation of lubricant above the lower level of the tubes when each compressor is operating. When the lubricant level rises above the lower level of a tube, the tube sucks lubricant from the sump into the cylinder intake when a compressor is operating. The lubricant is then entrained as liquid droplets in the circulating refrigerant for circulation through the system, and the lubricant droplets then return and drop into the compressor sump when the refrigerant enters the compressor intake. 
     The refrigerant and entrained lubricant, as indicated by the arrows, flows into the inlet to the interior of compressor shell or can  150 . The lubricant droplets fall into sump  106 , and the refrigerant gas flows through the holes  150  of the compressor motor to cool the motor, and the refrigerant gas then pass flows to the intake manifold to the cylinders  154  of primary compressor  24 . The lubricant in the sump  106  is at the pressure of the refrigerant entering can  150 , which is slightly higher than the pressure of the refrigerant entering the intake  154  to the cylinders. Aspiration tube  110  extends from the lubricant sump to the point of lowest pressure in the intake to the cylinders of the primary compressor. 
     In the system as shown in  FIG. 4 , since booster  22  is not operating, it is important to prevent the lubricant from entering into and accumulating in the sump of the booster. If the lubricant were permitted to enter into booster  22 , the lubricant would merely accumulate in the sump of booster  22 , since compressor  22  is inoperative and, therefore, no aspiration occurs through tube  108 . This would eventually result in adequate lubrication for primary compressor  24 . 
     Entry of the lubricant into the booster  22  when the booster is inoperative is prevented by a pair of traps  112  and  114 . Trap  112  in conduit section  100  prevents entry of lubricant from conduit segment  46 b, and trap  114  in conduit section  48  prevents entry of lubricant from conduit segment  48 . Accordingly, all circulating lubricant is directed to the operating primary compressor  24 . 
     If the level of lubricant  106  in the sump of primary compressor  24  rises above the bottom of siphon tube  110 , the excess accumulation will be aspirated into the intake of the primary compressor and then circulates with the refrigerant vapor leaving the compressor. 
     Referring to  FIG. 5 , which shows both compressors operational, the refrigerant and entrained oil flow via conduit  100  to the inlet to booster can or shell  110 . The lubricant droplets fall into the lubricant sump  104 , and the refrigerant gas flows through the holes  162  of the compressor motor to cool the motor, and the refrigerant gas then flows to the intake manifold to the cylinders  164  of booster compressor  22 . The lubricant in sump  104  is at the pressure of the refrigerant entering can  160 , which is slightly higher than the pressure of the refrigerant entering the intake  164  to the cylinders. Aspiration tube  108  extends from the lubricant sump to the point of lowest pressure in the intake to the cylinders of the booster compressor. 
     When both compressors  22  and  24  are operating (see FIG.  5 ), the circulating refrigerants and any entrained lubricant, are drawn from conduit segment  46 b through trap  112  and conduit segment  100  into the shell of booster  22 . The refrigerant discharge from compressor  22  then flows through conduit segment  48  and trap  114  to the intake manifold of compressor  24 . If the level of lubricant  104  in the sump of compressor  22  rises above the bottom of tube  108 , the excess lubricant is aspirated into the cylinder of compressor  22  and then entrained in the refrigerant fluid delivered through conduit  48  and trap  114  to the intake to compressor  24 . Similarly, if the level of sump lubricant  106  rises above the bottom of the tube  110 , the lubricant is aspirated into the cylinder of compressor  24  and is then circulated as before. In this way, the sump lubricant levels in both compressors are maintained at desired levels, and both compressors are lubricated. 
     While the lubricant management has been described for a pair of compressors connected in series, the same system of traps and siphon tubes can be used with compressors connected in parallel. The tubes maintain desired levels of lubricant when in each operating compressor, and the traps prevent delivery and build-up of lubricant in compressor  22  when it is not operating. 
     EXAMPLE 1 
     Referring now to  FIGS. 6 and 7 , a heating capacity chart and an exemplary operating sequence are shown. It will be recalled that thermostat  62  preferably has three stages in the heat mode. The thermostat stages signal microprocessor  54 , which, in turn, sends signals to allow (i.e., control) the operation of the compressors, and/or one or both cylinders of the compressors, and/or the economizer. 
     Typically, the heating cycle starts when the first stage or step of indoor thermostat  62  calls for heat. When this occurs somewhere between 75° F. and above 50° F. outdoor ambient temperature, as sensed by sensor  56 , one piston of primary compressor  24  is allowed to operate, i.e., is activated by a signal from microprocessor  54 . Depending on the configuration, this provides 40% or 50% of the displacement of primary compressor  24 . This mode of operation is identified in  FIGS. 6 and 7  as 010, signifying 0 cylinder operation of the booster, 1 cylinder operation of the primary and no operation of the economizer. 
     When the ambient temperature drops to about 50° F., operation of the second piston of primary compressor  24  is allowed by microprocessor  54 , but only if called for by the second stage of thermostat  62 . This mode of operation is indicated by the 020 lines on  FIGS. 6 and 7 . 
     No additional heating capacity can be brought on line until the outdoor ambient further drops to about 43° F. or so, even if the third step of the indoor thermostat calls for more heat. This is designed to prevent the system from supplying more capacity than is really needed, as, if it were to be supplied, it would come about at a low efficiency level because the condenser would operate at an unnecessarily high pressure and the evaporator would operate at an unnecessary low pressure. 
     When the outdoor ambient temperature reaches about 43° F., microprocessor  54  allows operation of both cylinders of booster compressor  22  (100% booster operation), but with operation of only one cylinder (40% or 50% displacement) of the primary compressor, and without operation of the economizer. This mode is indicated at the  210  lines in  FIGS. 6 and 7 . This mode becomes the maximum capacity heat capacity allowed until the outdoor ambient temperature drops to about 33° F. Then, microprocessor  54  also allows operation of the economizer  28  by sending a signal to open valve  40 , but this signal is sent only if the third stage or step of indoor thermostat  62  calls for more heat. This mode of operation is indicated at the  211  line in  FIGS. 6 and 7 . 
     When outdoor ambient temperature reaches about 25° F. a signal from microprocessor  54  allows operation of both cylinders of primarily compressors  24 , along with both booster cylinders and the economizer. This mode of operation is indicated at the  221  lines in  FIGS. 6 and 7 . This is the maximum capacity heat pump mode, and it continues in operation until the outdoor ambient temperature reaches about 15° F. 
     At the 15° F. outdoor ambient temperature level, and if the third step of the indoor thermostat is calling for more heat, microprocessor  54  sends a signal to allow operation of back-up electric resistance heating. 
     As seen from  FIG. 6 , the BSHP (boosted source heat pump) of this invention meets the heating requirement without the need for back-up resistance heating all the way down to an outdoor ambient temperature of about 10° F. This is far superior to a typical prior art heat pump, the capacity line of which is labeled “HP-TODAY” in  FIG. 6 , where back-up resistance heat is required at about 30° F. outdoor ambient temperature. Bearing in mind that the high cost of resistance back-up heat is one of the main disadvantages of typical prior art heat pumps, the significant advantages of the present invention are apparent. 
     EXAMPLE 2 
     While incorporation of economizer  28  in the system is preferred, the economizer can be omitted. In that case, the displacement ratio of the booster compressor  22  to primary compressor  24  would be increased sufficiently to realize a system capacity about that of the system with the economizer, understanding that a system efficiency loss would occur due to the absence of the economizer. In this case, at conditions of 0° F. outdoor ambient and 70° F. indoor heated space temperature, the heating coefficient of performance (C.O.P.) will be at least 1.5 and may approach 2. By way of example, the displacement ratio of booster compressor  22  to primary compressor  24  could be increased to about 1.4:1 to about 1.7:1. Referring to  FIGS. 6 and 7 , with the economizer eliminated, the  211  line would be eliminated, and the  211  line is replaced by a  220  line, with the allowance point being the outdoor ambient temperature at which the  221  line was previously allowed, i.e., between about 15° F.-25° F. in Example 1. The  220  line is shown as a dashed line in FIG.  6 . 
       FIG. 8  shows cooling performance for the heat pump of the present invention, and  FIG. 9  shows a typical cooling operating sequence. 
     With outdoor ambient temperature of about 80° F., and with the first stage of thermostat  62  calling for cooling, microprocessor  62  allows operation of only one piston (40%-50% capacity) of the primary compressor  24 . It is expected that this will handle most of normal cooling requirements. This is the  010  line in FIG.  8 . 
     At about 85° F. outside ambient temperature, microprocessor  54  allows operation of only both pistons of primary compressor  24  (100% primary capacity) if called for by the thermostat. This is indicated at the  020  line in  FIGS. 8 and 9 . 
     At about 105° F. outdoor ambient, microprocessor  54  allows operation of both pistons of booster compressor  22  and both pistons of primary compressor  24 . (100% capacity for both compressors). This is indicated at the  220  line of  FIGS. 8 and 9 . This will be effective to meet cooling needs up to about 115° F. outdoor ambient. 
     In addition to the foregoing, operation of both primary pistons, both booster pistons, and the economizer can be manually selected for special requirements, e.g., quick cool down, or to handle large numbers of people in a room, or high humidity conditions, etc. This is indicated at the  221  line in  FIGS. 8 and 9 . 
     It will be noted in  FIGS. 6 ,  7 ,  8  and  9 , that whenever operation of the booster compressor is allowed, both cylinders (100% capacity) are utilized. This means that variable capacity is not required for booster  22 . Accordingly, booster  22  can be a single speed compressor (or any type). This will reduce the manufacturing cost of the system, since single speed compressors can be obtained less expensively than the Bristol or similar type TS compressor. 
     It will be understood that the operating modes and sequences illustrated in  FIGS. 6-8  are only be way of example. Other operating modes and sequences can be employed within the scope and interest of this invention. 
     While this invention has been described in terms of a system for both heating and cooling, the invention can be applied for a heating system alone or a cooling system alone. In that event, the four-way valve  80  would be eliminated, and the refrigerant would always flow in one direction only. 
     If the booster compressor fails to operate when called for (for example because of an electrical contractor problem) the microprocessor is programmed to sense the non-operation of the booster and to proceed to a single TS mode of operation (for the primary compressor). The heating operating sequence of this single TS mode is as follows, with reference to FIG.  7 :
         1. From 50° F.-75° F. outdoor ambient temperature, only one cylinder of the primary compressor is allows ( 010  operation).   2. From 43° F.-50° F., operation of both cylinders of the primary compressor only are allowed ( 020  operation).   3. From any temperature below and up to 43° F., backup heat will be allowed, along with operation of both cylinders of the primary compressor. It will be noted that in this mode of operation, the  210 ,  211  and  221  steps of operation are eliminated, because the booster is inoperative.       

     In this mode of operation, where only the TS primary compressor is greater, the cooling sequence is also varied by the microprocessor. 
     Since the system is sized to deliver its rated cooling capacity at normal AR 1  operating conditions with a 0-2-0 combination, it would just operate as a typical system of today does with the exception that 0-2-0 would not be allowed until 85° F. outdoor ambient or thereabout whereas today, it is allowed whenever the indoor thermostat would call for it. In this case as mentioned above (booster failing to operate for some reason), the second step of the cooling thermostat could call for 0-2-0 at any outdoor ambient about 65° F. or so. The first step of the cooling thermostat would still call for 0-1-0 as long as the outdoor ambient temperature is above 60° F. or so. 
     Accordingly, the one TS sequence of operation for cooling is as follows:
         1. When the indoor thermostat calls for step  1  of cooling, operation of only one cylinder of the primary compressor will be allowed (the 010 mode) by the microprocessor as long as the sensed outdoor ambient temperature is in the range of about 60° F.-85° F.   2. When the indoor thermostat calls for second stage cooling, and the outdoor ambient temperature rises to about 85° F., operation of both stages of the TS primary compressor will occur (the 020 mode).   3. The operation of both cylinders of the primary TS compressor will be allowed on manual selection of the second step of the indoor thermostat and as long as outdoor ambient temperature is above about 60° F. (the 020 mode).       

     While preferred embodiments have been shown and described, various modifications and substitutions may be made thereto without departing from the spirit and scope of the invention. Accordingly, it is to be understood that the present invention has been described by way of illustrations and not limitations.