Abstract
A system for reducing the refrigerant pressure in an oil sump or in a cavity of a housing. The invention is particularly useful for reducing pressure in a compressor for heat pump applications that has been validated for water chiller operations or in turbine and generator systems in ORC systems generating electricity using refrigerant, the ORC systems essentially being a heat pump application operating in reverse. An auxiliary compressor, an auxiliary condenser or an ejector pump may be used to reduce pressure in the oil sump, to separate refrigerant from oil. The auxiliary compressor, the auxiliary condenser or the ejector pump may also be used to reduce the pressure of refrigerant in the housing of a compressor in heat pump applications at temperatures and pressures at which the compressor was validated for water chiller applications and of the turbine and generator in ORC applications.
Claims
1. A system for cooling a semi-hermetic compressor motor in a refrigeration or heat pump system using a refrigerant, the system including a refrigeration circuit comprising: a compressor configured to raise a pressure of a refrigerant gas, a main condenser in fluid communication with the compressor and configured to condense the refrigerant gas into a high pressure liquid, an expansion valve in fluid communication with the condenser, the expansion valve configured to convert the high pressure liquid into a mist of liquid entrained in gas, an evaporator in communication with the expansion valve and with the compressor, the evaporator configured to change a state of liquid refrigerant to refrigerant gas, the compressor further including a compressor motor, the compressor motor further including shaft, a housing for the motor, the housing having a cavity, the motor housed in the housing, the motor having a stator configured to alternate an electric field, and a rotor attached to the shaft, the rotor and the shaft configured to rotate with the alternating electrical field; a refrigerant inlet in the housing; a refrigerant outlet from the housing; and a refrigerant pressure reducing device in communication with the housing and with a low pressure region of the system that is downstream of the expansion valve and upstream of a compressor inlet, wherein the refrigerant pressure reducing device is configured to draw refrigerant from the housing, reduce refrigerant pressure to a pressure lower than that of the low pressure region of the system, and direct refrigerant toward the low pressure region of the system.
2. The system of claim 1, comprising a magnetic bearing system configured to the shaft when the system is operational.
3. The system of claim 1, wherein the refrigerant pressure reducing device is an auxiliary compressor.
4. The system of claim 1, wherein the refrigerant pressure reducing device is an ejector pump.
5. The system of claim 1, wherein the refrigerant pressure reducing device comprises a circuit in communication with the housing and the low pressure region of the system, the circuit comprising an auxiliary condenser configured to cool and condense refrigerant gas from the housing, a conduit extending between the housing and the auxiliary condenser to transport refrigerant to the auxiliary condenser, a fluid storage space in fluid communication with the auxiliary condenser, the fluid storage space configured to store condensed refrigerant after cooling in the auxiliary condenser, a liquid pump configured to pump refrigerant from the fluid storage space to the low pressure region of the system, and a liquid level sensor configured to an amount of liquid in the fluid storage space.
6. The system of claim 1, wherein the refrigerant pressure reducing device comprises a circuit in communication with the housing and the low pressure region of the system, the circuit comprising an auxiliary condenser configured to cool and condense refrigerant gas, a conduit extending between the housing and the auxiliary condenser to transport refrigerant gas from the housing to the auxiliary condenser, a fluid storage space in fluid communication with the auxiliary condenser, the fluid storage space configured to store condensed refrigerant after cooling in the auxiliary condenser, and a valve configured to regulate refrigerant flow from the fluid storage space to the low pressure region of the system.
Description
BRIEF DESCRIPTION OF THE DRAWINGS
(1) FIG. 1 is a schematic of a typical well-known refrigeration system, but specifically depicting the oil sump.
(2) FIG. 2 is a cross-sectional view of a prior art compressor depicting the associated sump system.
(3) FIG. 3 is a simplified schematic of a prior art compressor lubrication circuit.
(4) FIG. 4 is a simplified schematic of the compressor lubrication circuit of the present invention.
(5) FIG. 5 a simplified schematic of an embodiment of the compressor lubrication circuit of the present invention utilizing an auxiliary compressor.
(6) FIG. 6 is a simplified schematic of an embodiment of the compressor lubrication circuit of the present invention utilizing an ejector pump.
(7) FIG. 7 is a simplified schematic of an embodiment of the compressor lubrication circuit of the present invention utilizing an auxiliary condenser and liquid pump.
(8) FIG. 8 is a cross-sectional view of a prior art cooling scheme utilized for cooling a compressor motor having a centrifugal compressor attached at either end of the rotor shaft.
(9) FIG. 9 is a simplified schematic of the motor and compressor depicted in FIG. 8.
(10) FIG. 10 is a simplified schematic for the motor depicted in FIG. 8 of an embodiment of the present invention using a motor cooling arrangement having a pressure reducing device in communication with the motor cavity and intermediate a low pressure point in the refrigeration system.
(11) FIG. 11 is a simplified schematic of an embodiment of FIG. 10 for the motor cooling arrangement of the present invention utilizing an ejector pump.
(12) FIG. 12 is a simplified schematic of an embodiment of FIG. 10 for the motor cooling arrangement of the present invention utilizing an auxiliary condenser.
(13) FIG. 13 is a modification of the motor cooling arrangement of FIG. 12 utilizing a pair of vessels connected to the main condenser to return fluid from the auxiliary condenser to the evaporator.
(14) FIG. 14 is a modification of the motor cooling arrangement of FIG. 10 utilizing an auxiliary compressor in conjunction with a thermal expansion valve instead of a fixed orifice.
(15) FIG. 15 is a further embodiment of the motor cooling arrangement of FIG. 10.
(16) FIG. 16 is a prior art schematic of an organic Rankine Cycle system, depicting operation in reverse to the system depicted in FIG. 1.
DETAILED DESCRIPTION OF THE INVENTION
(17) FIG. 1 is a schematic of a typical refrigeration system depicting a motor/compressor 23 in fluid communication with a condenser 25 which is in fluid communication with an evaporator 27. Refrigerant gas is compressed to a higher pressure in compressor 23. The high pressure refrigerant gas, after flowing to condenser 25 is condensed to a high pressure liquid via heat exchange, not shown. The high pressure refrigerant liquid is then sent to evaporator 27. An expansion valve 31 intermediate condenser 25 and evaporator 27 expands the high pressure refrigerant liquid to a mist, the mist being a mixture of gas and liquid at a lower temperature. In evaporator 27, the liquid refrigerant is evaporated, absorbing heat from a heat exchange fluid, as liquid refrigerant mist changes phase from liquid to gas. The cooled heat exchange fluid may be sent directly to a building environment or indirectly to an intermediate medium, such as a chiller for storage of chilled water until required. Refrigerant gas from evaporator 27, having undergone a phase change, is at a low pressure and serves as a refrigerant gas source for compressor 23. Also depicted in FIG. 1 is a sump 10, which collects the oil from operation of compressor 23 and is fundamental to proper functioning of compressor 23. Sump 10, as shown, is below compressor so that lubricating oil flows to sump 10 by gravity.
(18) FIG. 2 is a cross-sectional view of a prior art centrifugal compressor and associated sump system. FIG. 2 depicts compressor 23 and oil sump 10. Some lubricating oil is retained in an auxiliary oil reserve 32, intended to keep some oil supply during coast-down in the event of a power failure. Compressor 23 includes an inlet 34 which receives refrigerant gas from a low pressure source, typically an evaporator (shown in FIG. 1). The refrigerant gas is compressed by an impeller 36 before being delivered to volute 38. Lubrication is provided to lubricate shaft seal 40, main journal and thrust bearing 42, thrust collar 44, double bellows shaft seal 46, low speed gear rear bearing 48, pinion gear shaft bearing 50, thrust collar bearing 52 and low speed gear 54. Lubricant and refrigerant are in contact with one another as a small amount of pressurized refrigerant gas invariably leaks from impeller 36 into the various lubricated components described above. After lubricating the compressor components, the lubricant/refrigerant mixture drains by gravity through conduit 56 into sump 10. While settling in oil sump 10 before being re-circulated, refrigerant gas is released from the mixture in excess of the steady-state solubility, dependent upon the pressure and temperature conditions in the sump. Although the exact amount of refrigerant that may collect in sump 10 at any one instant of time is difficult to measure, it is estimated that the refrigerant that is absorbed by the oil and which should be separated in sump 10 is about 1-3% of the total flow of the compressor. To avoid an undesired oil viscosity as the oil cools once the compressor is stopped, an oil heater 57 is provided, heating or maintaining the lubricant within a predetermined temperature range so that it has the proper viscosity as soon as compressor 23 starts. Fluid is pumped from sump 10 by submersible pump 60 and sent to oil cooler 62, which is activated only when the oil is above its predetermined operating temperature. The refrigerant gas that is separated from the oil in the sump is sent to compressor inlet 34 through a vent line 102 (see FIG. 3), while oil, which still may include miscible refrigerant gas, is sent to oil reserve 32 wherein it is metered to the compressor for lubrication purposes, after which the lubrication cycle repeats.
(19) In heat pump systems in which the evaporation pressure and temperature tend to be substantially higher than in water chillers, the oil temperature also should to be set to a higher value in order to keep the oil dilution at an acceptable value. As a result of this higher temperature, the oil viscosity will be reduced if the same grade oil is used as in water chiller systems. An oil grade with higher viscosity can be used to compensate for the higher temperatures experienced in heat pump systems. But even with this compensation for the viscosity, the temperature elevation of the oil in such heat pump systems raises other issues. Among these is a risk of failure of the shaft seals and bearings if the oil temperature should become too high. The present invention provides a system that compensates for some of the differences between operation of standard chillers and higher temperature heat pumps due to the temperature difference of operation that also affects oil temperature. This invention should extend the range of application of current standard compressor systems used in chiller applications to heat pump applications, with minor, inexpensive modifications.
(20) FIG. 3 is a simplified version of the cross sectional representation of prior art FIG. 2 which shows a simplified lubrication cycle schematic (for illustration purposes), with lubricant and miscible refrigerant being drained from compressor 23 through conduit 56 to sump 10, and then refrigerant gas at sump pressure returned to the compressor inlet along gas conduit 102, while lubricant with miscible refrigerant is returned to compressor 23 along conduit 104.
(21) Although FIGS. 3 through 7 are simplified schematics (for illustration purposes) that depict the prior art and the improvement provided by the present invention, the features required for operation of lubrication circuit depicted in FIG. 2 are also present in the circuits represented in FIGS. 4-7, although with the addition of the innovative pressure reducing device 409, as set forth herein.
(22) FIG. 4 provides a simplified version of the present invention, again using a simplified schematic. In FIG. 4, a pressure reducing device 409 is positioned between sump 10 and compressor inlet 34 to draw refrigerant gas from the sump while reducing the pressure of refrigerant gas in the sump. Although pressure reducing device 409 is shown as connected to the inlet of compressor 34 through connection 411, it is not so restricted, and, as will be recognized by one of skill in the art, pressure reducing device 409 can be connected to any low pressure point of the refrigeration circuit. Most often this low pressure point is the evaporator 27, but may be by any connection to the system between the evaporator 27 or an evaporator inlet and compressor inlet 34, including compressor inlet 34. Pressure reducing device 409 enables lowering of the pressure (and temperature) of the refrigerant gas in the oil sump. As previously set forth, the lowering of the pressure of refrigerant gas in oil sump 10 has the beneficial effect of reducing the dilution of refrigerant in the oil, thereby mitigating the reduction of oil viscosity while providing proper lubrication of shaft seals and bearings. Lowering the refrigerant pressure in the oil sump initiates a virtuous cycle combining several combined benefits, one of which is the ability of refrigeration system 21 to operate at higher evaporation temperatures and pressures such as encountered in heat pump conditions. When operating at such heat pump conditions, the target for pressure reduction is to set the oil sump gas pressure at a value consistent with the validated range of the same compressor when operating as a water chiller. Thus, if a given type of compressor is validated, for example, for an evaporation temperature of 20 C. (68 F.) with a given refrigerant, the target will be to set the sump pressure corresponding to a 20 C. saturation temperature in heat pump operation, in order to set all the lubrication parameters at the same standard value as for chillers. Of course, this is not enough to guarantee that the machine will be reliable. While this course of action will not solve all of the problems in converting a standard compressor for chiller applications for use in high temperature heat pump applications, as other parameters such as design pressure, shaft power, bearing loads etc. must be validated, problems associated with lubrication should be solved. Although all of the detail of the system as shown in FIG. 2 is not shown in the simplified version of FIG. 4, it will be understood that all of the detail of the system shown in FIG. 2 also may be in the simplified system of FIG. 4, except that the novel pressure reducing device 409 is included between sump and a low pressure point of the refrigeration system 21.
(23) The pressure reduction in the oil sump can be achieved in different ways. FIG. 5 depicts a simplified version of an embodiment of the present invention, again using a simplified schematic for illustration of the invention. Although all of the detail of the system as shown in FIG. 2 is not shown in the simplified version of FIG. 5, it will be understood that all of the detail of the system shown in FIG. 2 also may be in the simplified system of FIG. 5, except that a pressure reducing device 509 is included between sump and a low pressure point of the refrigeration system 21. In FIG. 5, the pressure reducing device is a small additional auxiliary compressor 509 positioned between sump 10 and compressor inlet 34 to draw refrigerant gas from sump 10 while reducing the pressure of refrigerant gas in the sump. Auxiliary compressor 509 has its suction side connected to the gas volume of oil sump 10 and its discharge side connected, for example, to compressor inlet 34 of main compressor 23. In this implementation, the capacity of auxiliary compressor 509 is controlled in such a way that it keeps the refrigerant pressure in oil sump 10 at a pre-selected value as described above (e.g. corresponding to the saturated pressure of the refrigerant fluid at 20 C. in the above example). As discussed above and recognized by those skilled in the art, the discharge of auxiliary compressor 509 can also be connected to any lower pressure point in refrigeration system 21, such as evaporator 27 or any point between evaporator 27 and compressor inlet 34 as shown in FIG. 1.
(24) While the use of auxiliary compressor 509 is conceptually simple, it also has some drawbacks. Besides its additional manufacturing and operational cost, auxiliary compressor 509 is also a mechanical component with possible reliability and maintenance issues. In addition, its operational costs, specifically energy consumption, may be significant. Furthermore, in circumstances of variable operating conditions, the capacity control related to the use of such an auxiliary compressor 509 may be problematic. However, the use of auxiliary compressor 509 in refrigeration system 21 is a viable option to reduce refrigerant in sump 10.
(25) In another embodiment depicted in FIG. 6, a simplified schematic of an embodiment of the present invention, an ejector pump 609, also referred to as a jet pump, is depicted as the pressure reducing device associated with sump 10. Again, all of the detail of the system as shown in FIG. 2 is not shown in the simplified version of FIG. 6, and it will be understood that all of the detail of the system shown in FIG. 2 also may be in the simplified system of FIG. 6, except that ejector pump 609 is positioned between sump 10 and a low pressure point of the refrigeration system. In FIG. 6, high pressure gas from conduit 615, which is in fluid communication with condenser 25, after passing through an expansion valve (not shown), if required, is used to provide the energy to operate ejector pump 609. At the ejector outlet, the mixture of this high pressure refrigerant fluid from condenser 25 and the low pressure gas pumped from oil sump 10 is sent to a low pressure point in the refrigeration system, preferably the evaporator. Although shown in FIG. 6 as in direct fluid communication with compressor inlet 34 via conduit 611 (for consistency with FIGS. 4 and 5), the low pressure point may be at any intermediate location between compressor 23 and evaporator 27 that is at a low pressure, as previously discussed. The advantage of this embodiment, using an ejector pump, is that it avoids moving parts such as found with the use of auxiliary compressor 509 of FIG. 5. This embodiment does suffer from a drawback, because ejector pumps 609 usually have a relatively poor efficiency, and thus penalize the energy efficiency of the refrigeration system. Nevertheless, the use of ejector pump 609 in refrigeration system 21 is a viable option to reduce refrigerant in sump 10, while allowing the lubrication system to operate with higher temperature systems seen in heat pump applications.
(26) In a preferred embodiment of the present invention depicted in FIG. 7, a simplified schematic of an embodiment of the present invention, an auxiliary condenser 709 is depicted as the pressure reducing device associated with sump 10. Again, all of the detail of the system as shown in FIG. 2 is not shown in the simplified version of FIG. 7, and it will be understood that all of the detail of the system shown in FIG. 2 also may be in the simplified system of FIG. 7, except that auxiliary condenser 709 is included between sump 10 and a low pressure point of the refrigeration system. In FIG. 7, refrigerant gas from sump 10 is in fluid communication with auxiliary condenser 709 via conduit 713. Gas from sump 10 enters auxiliary condenser 709 where it is in heat exchange relationship with a cooling fluid flowing through cooling circuit 715. Cooling fluid in cooling circuit 715 cools the refrigerant gas, condensing it from a gas to a liquid, the liquid refrigerant being sent to liquid storage space 717 via conduit 730.
(27) The auxiliary condenser 709 is selected to provide a condensing pressure equal to the desired refrigerant pressure in oil sump 10. This requires the refrigerant gas in auxiliary condenser 709 to be cooled by a cooling fluid at a temperature lower than the cold source of the heat pump. For example, if the desired condensing pressure in the auxiliary condenser 709 corresponds to a 20 C. (68 F.) saturation temperature, auxiliary condenser 709 preferably is cooled with water having an entering temperature of about 12 C. (about 54 F.) and a leaving temperature of about 18 C. (about 64 F.). The cooling water may be provided from any available chilled water source as well as from ground water within the desired temperature range. The condensing pressure in auxiliary condenser 709 may be controlled by varying the flow and/or temperature of the cooling fluid through cooling circuit 715 of auxiliary condenser 709 to maintain the desired gas pressure in oil sump 10. As depicted in FIG. 7, liquid storage space 717 for condensed refrigerant may be a separate vessel as shown, or may be a separate storage space integral to auxiliary condenser 709.
(28) Per the principle of the system, liquid storage space 717 is at a lower pressure than the compressor inlet and the evaporator in the main refrigerant circuit. To avoid accumulation of liquid refrigerant in liquid storage space 717, refrigerant must be pumped from storage space 717 back to refrigerant system 21 by pump 719 that is controlled by liquid level sensor 721. This pump 719 has its suction side connected to fluid storage space 717 and its discharge side in fluid communication with refrigerant system 21. To reduce the head and the absorbed power of the pump, it is preferred to set the pump discharge to a low pressure portion of the main refrigerant circuit 21. While this low pressure region may be compressor inlet 34, as previously discussed with regard to FIGS. 3-6, FIG. 7 depicts the low pressure region as the conduit between expansion valve 31 and evaporator 27, although refrigerant may be sent to the low pressure region at any convenient point, such as between expansion valve 31 and compressor suction 34. It is also normally desired to avoid sending refrigerant liquid directly into compressor suction 34 (inlet) from liquid storage space 717 to avoid liquid flooding of compressor 23. Therefore, a location along the conduit between expansion valve 31 and evaporator 27 is a desirable and preferred refrigerant input, as is supplying this liquid refrigerant to evaporator 27, such as at the liquid inlet of evaporator 27. More specifically, if evaporator 27 is of the dry-expansion technology (either shell and tube or plate heat exchanger), then it is desirable to discharge the liquid refrigerant into the main liquid line at the evaporator inlet. If evaporator 27 is of the flooded type, falling film or hybrid falling film, an alternative is to discharge the liquid directly in the evaporator shell, at a location away from the suction pipe to avoid liquid carry-over to compressor inlet 34.
(29) Means also is provided to control the operation of liquid pump 719, depicted in FIG. 7 as liquid level sensor 721. A desired arrangement is to have fluid storage space 717 located at the outlet of auxiliary condenser 709, allowing liquid refrigerant to flow by gravity from auxiliary condenser 709 into storage space 717. This volume can either be included in the same shell as the auxiliary condenser 709, or as a separate vessel. The liquid level in this storage space is sensed by a liquid level sensor which includes a control loop, depicted simply as liquid level sensor 721. This control loop portion of liquid level sensor 721 manages the operation of liquid pump 719 in order to keep the liquid level in the fluid storage space 717 within predetermined, pre-set acceptable limits. Liquid pump 719 can either have a variable speed drive, with the speed being controlled by the control loop of liquid level sensor 721, or it may simply have an ON/OFF operation sequence, also under control of the same control loop.
(30) In another embodiment, a conventional mechanical pump 719 may be replaced by a purely static pumping system. In a variation to this embodiment, the static pumping system may utilize an ejector pump 609 powered by high pressure gas from main condenser 25. A mixture of pumped liquid from fluid storage space 717 and of high pressure gas from main condenser 25 is returned to evaporator 27. In still another variation to this embodiment, two fluid storage vessels 717 may be located below auxiliary condenser 715, each having an inlet (A) connected to the discharge port of auxiliary condenser 709 to receive condensed refrigerant liquid, an inlet (B) connected to receive gas from evaporator or main condenser 25, and each having outlet (C) connected to evaporator 27. Each of these connections has an automatic valve that can be opened or closed. The system is operated in batches, being activated by a control circuit using principles known to those skilled in the art. This system also is represented in FIG. 13, as associated with the cooling of a semi-hermetic motor.
(31) Any of these embodiments enable removal of refrigerant from oil in a lubricated compressor, and is not limited to use with a centrifugal compressor. The present invention may also find use with reciprocating compressors, scroll compressors and turbines as used in ORC systems, each of which requires lubrication. An auxiliary compressor 509 or ejector pump 609 may advantageously be used to remove refrigerant from oil in these units, as described above. These components may require significant power consumption or otherwise penalize system efficiency. An auxiliary condenser 709 has the further advantage of not requiring power to operate, assuming that water at the desired temperature is available. But it also requires a liquid pump 719 to transfer condensed refrigerant liquid to refrigerant system 21 at or near evaporating pressure. Although this does require a small amount of power, it is significantly less than the power required for operation of an auxiliary compressor 509, and there is no penalty to overall system efficiency such as with operation of ejector pump 609.
(32) The basic pressure reducing devices described above with reference to FIGS. 4-7 to separate refrigerant from lubrication systems may also be adapted for use in refrigeration circuits to extend the operational limits of refrigerant fluid for cooling semi-hermetic motors. These pressure reducing devices 409 can advantageously be utilized in heat pump systems which typically operate at higher temperatures than chiller systems. These pressure reducing devices 409 extend the motor cooling capability of the refrigerant, permitting the use of chiller system equipment for heat pump applications. In these systems, refrigerant is utilized to cool the motor and the motor cavity from heat generated by operation of the motor. The pressure in the motor housing and in the coil surrounding the motor stator without such pressure reducing devices is nearly equal to or slightly higher than the pressure in the evaporator. But, pressure reducing devices are controlled to maintain the pressure in the motor cavity at a preset value below that of the compressor inlet and preferably lower than that of the evaporator so that refrigerant gas can be drawn through the housing. For a system operating in heat pump applications, it is desired to maintain the pressure in the motor cavity at a preset value below the pressure at the compressor inlet, for example, at a saturation temperature of 20 C. corresponding to the desired pressure for a given refrigerant. These values typically correspond to the temperatures at which the compressor is validated when the system operates as a water chiller system.
(33) FIG. 8 depicts a prior art cooling scheme utilized for cooling a semi-hermetic motor 350 driving a compressor, as set forth in prior art patent application WO 2012/082592 A1 assigned to the assignee of the present invention. In the cross sectional representation of motor of FIG. 8, a centrifugal compressor 376 is shown with an impeller 91 attached to either end of motor shaft 128 in a preferred embodiment, but the invention is not so limited, as the motor cooling scheme may be utilized with any type of compressor driven by a semi-hermetic motor in a refrigeration circuit and does not require a compressor attachment at both ends of shaft 128 as depicted in FIG. 8. In FIG. 8, liquid refrigerant from the condenser is provided via a line 78 to an expansion device 80 which reduces the pressure and temperature of the liquid refrigerant, preferably converting it to a mist, as previously defined, a mixture of refrigerant liquid droplets and gas. The refrigerant mixture then enters motor inlet 81 passing into motor housing 382, which is hermetically sealed to prevent gas (refrigerant) leakage across its boundaries.
(34) The operation of motor 350, which comprises a motor stator 88 and motor rotor 129, generates heat. Motor stator 88, motor rotor 129 and shaft 128 are positioned in a cavity 352 within motor housing 382. Rotor 129 is attached to shaft 128, and an alternating electrical field in motor stator 88 rotates rotor 129 and shaft 128. Also depicted in FIG. 8 are bearings 90 at either end of motor shaft 128, which support rotor 129 during operation. In FIG. 8, these bearings 90 are depicted as mechanical bearings, but, as recognized by those skilled in the art, also may be magnetic bearings. Like motor 350, magnetic bearings are operated by strong magnetic fields and also generate heat. Thus, heat is generated within motor housing 382 whether bearings 90 are magnetic bearings or mechanical bearings. The refrigerant introduced into motor housing 382 through motor inlet 81 is used to remove heat from both motor 350 and bearings 90.
(35) In this particular embodiment, after entering motor housing 382 through motor inlet 81, refrigerant passes into a coil that surrounds motor stator, the refrigerant removing heat from motor stator 88. The refrigerant then passes into a line 378 that conveys the refrigerant to a secondary cavity 380. The refrigerant entering secondary cavity 380 may be a mist, that is, it is refrigerant in two phases. The liquid phase 384 separates by gravity to the bottom of secondary cavity 380 and is sent to evaporator 27 through a first motor housing outlet 386 via line 388. Line 388 may include restriction 390, such as a fixed orifice or control valve to control the flow of refrigerant liquid. Restriction 390 prevents refrigerant gas from passing out of the motor via this path together with the liquid phase. The remaining refrigerant entering secondary cavity 380 passes through apertures 108 as a gas and reenters motor cavity 352 wherein it passes between stator 88 and rotor/shaft 128/129, as depicted by the arrows in FIG. 8, removing heat from these components. Some of the refrigerant also passes over bearings 90 removing heat and cooling them. The refrigerant traverses the gap between stator 88 and motor/rotor 129/128 as it removes heat from them. The refrigerant gas then is cycled back to evaporator 27 through a second motor housing outlet 387 via conduit 392 either directly or after passing through and around bearings 90. This is one of the many possible ways to circulate refrigerant in a motor to cool its various components, using a combination of liquid, gas, or two-phase refrigerant. While a variety of configurations is possible, prior art systems have in common that the pressure in the motor housing is close to the evaporating pressure of the refrigeration circuit. In the prior art cooling arrangement, the pressure in motor cavity 352 and in the coil surrounding stator 88 is nearly equal to the pressure in evaporator 27. One source of heat in the motor is the gas friction power generated by the speed of the rotating parts. This power increases with gas density. Thus, a higher gas pressure in the motor 350 generates higher friction losses that contribute to further heating of the motor. Also the gas temperature in the motor housing is equal to or greater than the saturated temperature and pressure of the refrigerant within the motor housing. Finally, the evaporation temperature of the refrigerant in the coil surrounding the stator is at least equal to the saturated pressure in the motor housing. The result is that when the temperature and the pressure increase in the evaporator, the temperature and pressure in the motor also increases. For this reason, the prior art cooling arrangement, although useful in semi-hermetic compressor applications used for water chillers, is not utilized in high temperature heat pump applications because required cooling cannot be provided by maintaining these temperature and pressure settings.
(36) A cooling arrangement using refrigerant can be successful when the pressure of the refrigerant in the motor cavity is lower than the pressure at compressor inlet 34 or the pressure of evaporator 27. Lowering the pressure of the refrigerant in the motor cavity 352 reduces the gas friction losses and improves motor cooling. When operating at heat pump conditions, an ideal target for pressure reduction is to set the pressure of the refrigerant from the motor cavity at a value consistent with the validated range of the same standard machine when operating as a water chiller. For instance, if a given type of compressor and associated semi-hermetic motor is validated in chiller applications for a maximum evaporation temperature of 20 C. with a given refrigerant, the target will be to set the motor cavity to 20 C. saturation temperature in heat pump operation. Of course, it is not enough to guarantee that the motor cooling will be acceptable. Many other parameters must be checked and resolved, such as design pressure, shaft power, bearing loads, etc; but a solution to motor cooling problems is provided.
(37) The pressure reduction of refrigerant in the motor cavity 352 may be achieved in different ways. This pressure reduction may be achieved using the same equipment that was utilized for pressure reduction in oil sump 10, described above.
(38) FIG. 9 is a simplified version of FIG. 8 showing the circuitry from motor inlet 81 for the refrigerant fluid through motor 350. Liquid refrigerant in line 388 passes through restriction 390 to conduit 392 which channels the refrigerant to evaporator 27.
(39) FIG. 10 depicts an embodiment of the present invention, again using a simplified schematic. Although all of the detail of the system as shown in FIG. 8 is not shown in the simplified version of FIG. 10, it will be understood by one skilled in the art that all of the detail of the system shown in FIG. 8 with regard to the motor 350 also may be included in the embodiment of the invention depicted in FIG. 10. This omitted detail is not required to understand the improvement depicted in FIG. 10. Generically, FIG. 10 depicts a pressure reducing device 409 in communication with motor cavity 352, pressure reducing device 409 being intermediate a low pressure point in the refrigeration system and the motor cavity. In FIG. 10, this low pressure point in refrigeration system 10 may be evaporator 27 as shown, but it also may be the compressor suction (i.e. inlet 34) or other low pressure point. In FIG. 14, pressure reducing device 409 is a small additional auxiliary compressor 509 positioned between motor 350 and the evaporator 27 or compressor inlet 34 to draw refrigerant from motor cavity 352. In the arrangement depicted in FIG. 14, a schematic diagram in accordance with FIG. 10 desirably should not be adopted, as the arrangement of FIG. 10 contemplates some liquid flowing though orifice 390 into the inlet of the pressure reducing device 409, which is not acceptable when this device is an auxiliary compressor such as contemplated in FIG. 14, with associated potential of compressor flooding. To avoid this, means must be provided to avoid sending an excessive amount of liquid through the orifice at motor inlet 81. An example of such implementation is set forth in FIGS. 14 and 15, FIGS. 14 and 15, differing in how the fluid entering motor cavity through expansion valve 802 is controlled. In FIG. 14, the circuit of FIG. 10 is modified as follows: the fixed orifice at motor inlet 81 set forth in FIG. 10 includes a thermostatic expansion valve 802 used to reduce the refrigerant flow to the stator coil. The fixed orifice 390 set forth in FIG. 10 is replaced by the thermostatic expansion valve 802 used to reduce the refrigerant flow to the stator 88. The sensor 804, which may be a temperature sensor, associated with expansion valve 802 may be located on line 378, or at any convenient location on the motor housing. With this arrangement, only some gas exits from motor housing 382 and enters cavity 380 through line 378. The liquid phase 384 is eliminated and liquid line 388 may be removed as liquid in secondary cavity 350 is eliminated, as shown in FIG. 14. Since a reduced amount of refrigerant enters housing 382 through expansion valve 802, a reduced amount or refrigerant gas exits from compressor housing 382 through line 392, ensuring that there are no liquid droplets at the suction of the auxiliary compressor, as desired.
(40) In this implementation, the capacity of pressure reducing device 409 (auxiliary compressor 509 in FIG. 15) is controlled in such a way that it keeps the pressure in motor cavity 352 at a pre-selected value. This preselected value may correspond to a maximum evaporation temperature for a given refrigerant, which may be the same temperature for a compressor operating under heat pump conditions as a standard compressor when operating as a water chiller. For example, the pressure may be set to correspond to a temperature of 20 C. As discussed above and recognized by those skilled in the art, the discharge of pressure reducing device 409 such as an auxiliary compressor 509 can also be connected to any lower pressure point in refrigeration system 21, such as evaporator 27 as shown in FIG. 1. In the schematic of FIG. 15, liquid does pool in secondary cavity 380, but the level is monitored by level control 805 which in turn controls thermostatic expansion valve 802 which controls the refrigerant entering motor housing 382.
(41) While the use of the auxiliary compressor is conceptually simple, it also has some drawbacks. Besides its additional manufacturing and operational cost, the auxiliary compressor is also a mechanical component with possible reliability and maintenance issues. In addition, its operational costs, specifically energy consumption, may be significant. Furthermore, in circumstances of variable operating conditions, the capacity control related to the use of such an auxiliary compressor may be problematic. However, the use of auxiliary compressor in refrigeration system 21 is a viable option to reduce refrigerant pressure in the motor cavity 352.
(42) In another embodiment depicted in FIG. 11, a simplified schematic of an embodiment of the present invention, an ejector pump 609, also referred to as a jet pump, is depicted as pressure reducing device 409 associated with motor 350. Again, all of the detail of the system as shown in FIG. 8 is not shown in the simplified version of FIG. 11, and it will be understood that all of the detail of the system shown in FIG. 8 also may be in the simplified schematic shown in FIG. 11, except that ejector pump 609 is positioned between motor 350 and motor cavity 352 and a low pressure point of the refrigeration system. In FIG. 11, high pressure gas from conduit 615, which is in fluid communication with condenser 25, after passing through an expansion valve, if required, is used to provide the energy to operate ejector pump 609. At the ejector pump outlet, the mixture of this high pressure refrigerant fluid from condenser 25 and low pressure refrigerant pumped from motor 350 is sent to a low pressure point in the refrigeration system, preferably evaporator 27. The refrigerant may be in direct fluid communication with compressor inlet 34 via conduit 611 as shown in FIG. 11, or the low pressure point may be at any intermediate location between evaporator inlet and compressor inlet 34. The advantage of this embodiment is that it avoids moving parts such as found with the use of auxiliary compressor 509 discussed above. The embodiment utilizing an ejector pump 609 such as depicted in FIG. 11 does suffer from a drawback, as ejector pumps 609 usually have a relatively poor efficiency, and thus penalize the energy efficiency of the refrigeration system. Nevertheless, the use of ejector pump 609 in refrigeration system 21 is a viable option to lower refrigerant pressure in motor 350 and return the refrigerant to the refrigerant circuit, while allowing the refrigerant to cool the motor as it operates with higher temperature systems seen in heat pump applications.
(43) In a preferred embodiment of the present invention depicted in FIG. 12, a simplified schematic of an embodiment of the present invention, a small auxiliary condenser 709 is depicted as the pressure reducing device associated with motor 350 and motor cavity 352. Again, all of the detail of the system as shown in FIG. 8 is not shown in the simplified schematic of FIG. 12, and it will be understood that all of the detail of the system shown in FIG. 8 also may be in the simplified system of FIG. 12, except that auxiliary condenser 709 is included between motor 350 and a low pressure point of refrigeration system 21. In FIG. 12, refrigerant from motor 350 is in fluid communication with auxiliary condenser 709 through line 388 and restriction 390 as well as through conduit 392. Refrigerant from motor 350 enters auxiliary condenser 709 where it is in heat exchange relationship with a cooling fluid flowing through cooling circuit 715 of auxiliary condenser 709. Cooling fluid in cooling circuit 715 cools the refrigerant gas, condensing it from a gas to a liquid that is sent to liquid storage space 717.
(44) The auxiliary condenser 709 is selected to provide a condensing pressure equal to the desired refrigerant pressure in the cavity of motor 350. This requires the refrigerant gas in auxiliary condenser 709 to be cooled by a cooling fluid at a temperature lower than the cold source of the heat pump. For example, if the desired condensing pressure corresponds to a 20 C. (68 F.) saturation temperature, auxiliary condenser 709 preferably is cooled with water having an entering temperature of about 12 C. (about 54 F.) and a leaving temperature of about 18 C. (about 64 F.). The cooling water may be provided from any available chilled water source as well as from ground water within the desired temperature range. The condensing pressure may be controlled by varying the flow and/or temperature of the cooling fluid through cooling circuit 715 of auxiliary condenser 709 to maintain the desired gas pressure in the cavity of motor 350. As depicted in FIG. 12, fluid storage space 717 may be a separate unit as shown, or may be a separate storage space integral to auxiliary condenser 709. Regardless of the location of fluid storage space 717, liquid refrigerant in fluid storage space may be pumped conveniently from storage space 717 by pump 719 that is activated by liquid level sensor 721.
(45) Once refrigerant from the cavity of motor 350 has been condensed and sent to fluid storage space 717, it may be pumped back to refrigerant system 21 by liquid refrigerant pump 719 having its suction side connected to fluid storage space 717 and its discharge side in communication with a low pressure region in refrigerant system 21 to reduce the head and the absorbed power of the pump. While this low pressure region may be the compressor inlet, as previously discussed with regard to FIGS. 10 and 11, it is not desirable to send liquid to the compressor inlet, as this could flood the compressor with liquid refrigerant. Thus, refrigerant pump desirably should cycle to a low pressure region of the system such as to the conduit between expansion valve 31 and evaporator 27, (see FIG. 1) or to evaporator 27, such as at the liquid inlet of evaporator 27, although refrigerant may be sent to the low pressure region at any convenient point. As previously noted, this reduces the head and the absorbed power of the pump, as it is supplying this liquid refrigerant to evaporator 27. More specifically, if evaporator 27 is of the dry-expansion technology type (either shell and tube or plate heat exchanger), then it is desirable to discharge the liquid refrigerant into the main liquid line at the evaporator inlet. If evaporator 27 is of the flooded type, falling film or hybrid falling film, an alternative is to discharge the liquid directly in the evaporator shell, at a location away from the suction pipe to avoid liquid carry-over.
(46) Means also is provided to control the operation of liquid pump 719, depicted in FIG. 12, means being identified as liquid level sensor 721. A desired arrangement is to have fluid storage space 717 located at the outlet of auxiliary condenser 709, allowing liquid refrigerant to flow by gravity to fluid storage space 717. This volume can either be included in the same shell as the auxiliary condenser 709, or as a separate vessel as depicted in FIG. 12. The liquid level in fluid storage space 717 is sensed by a liquid level sensor 721 which includes a control loop, depicted simply as liquid level sensor 721. This control loop portion of liquid level sensor 721 manages the operation of liquid pump 719 in order to keep the liquid level in the fluid storage space 717 within pre-set acceptable limits. Liquid pump 719 can either have a variable speed drive, with the speed being controlled by the control loop of liquid level sensor 721, or it may simply have an ON/OFF operation sequence, also under control of the same control loop. Pump 719 returns refrigerant liquid back to refrigeration system 21. In order not to flood compressor inlet 34 with liquid, refrigerant may be returned to refrigeration system anywhere between expansion device 31 and evaporator 27 as shown in FIG. 12, including evaporator 27. In FIG. 12, the centrifugal compressor is a two-stage compressor, so that low pressure gas refrigerant is input into the first stage compressor inlet and high pressure gas is discharged into condenser 25 from the second stage compressor.
(47) In another embodiment, a conventional mechanical pump may be replaced by a purely static pumping system. In a variation to this embodiment, the static pumping system may utilize an ejector pump powered by high pressure gas from main condenser 25. A mixture of pumped refrigerant liquid from fluid storage space 717 and of high pressure refrigerant gas from main condenser 25 is returned to evaporator 27 as a mist. Alternatively, this refrigerant may be returned to compressor inlet 34.
(48) In still another variation of this embodiment, as depicted in FIG. 13, two vessels may be located below auxiliary condenser 709, each having an inlet connected to the liquid outlet from auxiliary condenser 709 to receive condensed refrigerant liquid via conduit 730, a high pressure gas inlet 723 connected to receive high pressure gas, from main condenser 25 as shown in FIG. 13, and each having outlet 725 connected to evaporator 27. Condenser 25 is a convenient source for the high pressure gas in FIG. 13, but any other high pressure gas source may be utilized. High pressure gas inlet 723 provides the power to empty the fluid storage vessels or spaces 717, forcing the liquid from the fluid storage vessels 717 into the evaporator. The valves, depicted as valves 17, 18 and 19 in FIG. 13, are actuated to perform the function of alternatively emptying and filling each fluid storage vessel 717. Their operation is straightforward to those skilled in the art, having been used in some ice skating rinks to replace the liquid pump with the two receivers used alternatively: one being filled with the liquid draining from the auxiliary condenser, while the other is emptied by high pressure gas from the condenser. Each of these connections has an automatic valve that can be opened or closed. The system is operated in batches, being activated by a control circuit using principles known to those skilled in the art. Liquid pump 719 is not required in this arrangement.
(49) FIG. 15 is an alternative arrangement to that shown in FIG. 14. Both FIGS. 14 and 15 illustrate a pressure reducing device which is an auxiliary compressor. FIG. 15 provides another mode of active control for motor cooling by controlling the refrigerant introduced into motor 350 in order to avoid the intake of refrigerant liquid into auxiliary compressor 509. In FIG. 14, expansion valve 802 controls the flow of refrigerant into and from the coil that surrounds stator 88. Liquid refrigerant is introduced from condenser 25 (or subcooler if utilized) into the coil(s) that surrounds stator 88 through expansion valve 802 situated in line or conduit 378, see FIG. 8. Expansion valve 802 is controlled by a level sensor 805 that monitors the height of the liquid fluid column in secondary cavity 380. Refrigerant flowing through expansion valve 802 expands while having its pressure lowered. On entering secondary cavity 380, the liquid from the two-phase flow will fall by gravity to the bottom of secondary cavity 380. The amount of liquid refrigerant in secondary cavity 380 is determined by sensor 805 that detects fluid height in secondary cavity 380. Once the liquid height achieves a preselected level as determined by sensor 805, expansion valve 802 may be activated to reduce the flow of refrigerant fluid into secondary cavity. No liquid line is required between secondary cavity 380 and pressure reducing device 409. Only refrigerant gas will flow between rotor 129 and stator 88 and through line 392 to device 409. The increase of liquid refrigerant height as detected by sensor 805 in secondary cavity 380 indicates that no more refrigerant liquid should be sent into the motor, and expansion valve 802 will reduce the flow of refrigerant from stator 88. When the liquid refrigerant height in secondary cavity 380 has fallen below a preselected level as detected by sensor 805, a signal may be transmitted to expansion valve 802 to open and resume feeding refrigerant through conduit 378 to secondary cavity 380.
(50) In FIGS. 14 and 15, device 409 may be any of the aforementioned devices. Thus it may be an auxiliary compressor 509 as set forth in FIG. 5, ejector pump 609 as set forth in FIG. 6, auxiliary condenser as set forth in FIG. 7 or any combination thereof, such as a compressor/condenser system of a condenser/pumping system.
(51) Any of the embodiments allow for refrigerant to be used to cool the motor while removing refrigerant from the cavity of the motor, and the embodiments are not limited to a centrifugal compressor, which is exemplary in the Figures. Thus, the present invention may also find use with reciprocating compressors and scroll compressors, each of which requires motor cooling, and particularly when such compressors are adapted for use in heat pump systems. The system also provides cooling for bearings, particularly in systems utilizing magnetic bearings. The use of an auxiliary compressor 509 or ejector pump 609 may advantageously be used to remove refrigerant from the motor cavity. However, these components may require significant power consumption or otherwise penalize system efficiency. An auxiliary condenser 709 has the further advantage of not requiring power to operate, assuming that water at the desired temperature is available for heat exchange. But a system utilizing the auxiliary condenser also requires a liquid pump 719 to transfer condensed liquid to refrigerant system 21 at or near evaporating pressure. Although this does require a small amount of power, it is significantly less than the power required from operation of an auxiliary compressor 509, and there is no penalty to overall system efficiency when the liquid pump is replaced, such as with an ejector pump 609.
(52) The basic pressure reducing devices described above with reference to FIGS. 10-13 effectively remove refrigerant from the cavity of the motor while allowing the refrigerant to remove heat from motor operation as well as magnetic bearings, when the system is so equipped. These pressure reducing devices can advantageously be utilized in heat pump applications systems which typically operate at higher temperatures than chiller systems. These pressure reducing devices extend the motor cooling capability of the refrigerant, permitting the use of chiller system equipment for heat pump applications and enable refrigerant to be circulated through the motor housing.
(53) The description of the present invention provided above is with respect to a circuit having a compressor, such as a heat pump system or refrigeration system, where the condenser is on the higher pressure side of the refrigeration circuit and the evaporator is on the lower pressure side of the refrigeration circuit providing cooling to a motor, separation of refrigeration from lubricant or both. It will be understood that the present invention operates identically to an ORC system, which operates in reverse to the heat pump system as previously described, but where the evaporator is on the high pressure side of the circuit and the condenser is on the low pressure side of the circuit. The present invention serves to provide cooling to a generator, separation of refrigeration from lubricant or both.
(54) While the invention has been described with reference to a preferred embodiment, it will be understood by those skilled in the art that various changes may be made and equivalents may be substituted for elements thereof without departing from the scope of the invention. In addition, many modifications may be made to adapt a particular situation or material to the teachings of the invention without departing from the essential scope thereof. Therefore, it is intended that the invention not be limited to the particular embodiment disclosed as the best mode contemplated for carrying out this invention, but that the invention will include all embodiments falling within the scope of the appended claims.
