Compressor unit, thermal plant comprising the compressor unit, and method for operating a compressor unit in a thermal plant

Abstract

Proposed is a compressor unit (10) in which a compressor and a drive motor are arranged in a housing that encloses the compressor and the drive motor. The housing further encloses a lubricant reservoir for a lubricant. At least one heat exchange device (18) is arranged within the lubricant reservoir. The compressor unit (10) is intended in particular for use in a thermal system, such as a heat pump system. The heat exchanger is arranged in the flow path of liquefied process fluid (203) and is interposed between the condenser (20) or a phase separation tank (21) arranged downstream of the condenser. Thus, during operation of the thermal system, a fluid flows through the heat exchange device whose temperature is higher than the boiling point of the low-pressure process fluid (201).In this way, the lubricant in the compressor unit's lubricant reservoir is heated upon contact with the heat exchanger to a temperature at which the process fluid dissolved in the lubricant boils and thus outgasses from the lubricant. This prevents the accumulation of process fluid in the lubricant.

Description

TECHNICAL FIELD

[0001] The present description relates to a compressor unit of the type characterized in the claims, a thermal plant comprising the compressor unit, and a method for operating a compressor unit in a thermal plant, as also outlined in the claims. TECHNOLOGICAL BACKGROUND

[0002] As part of the targeted reduction in fossil fuel consumption, heat pumps are a focus of attention, particularly but not exclusively in the building heating sector. Their emissions during operation are determined exclusively by the generation of the electrical energy used to operate the heat pumps.

[0003] In a heat pump, a process fluid is evaporated in a low-pressure heat exchanger acting as an evaporator using a low-temperature heat source. The heat source can be, for example, soil, water, brine, or ambient air. The process fluid, which is gaseous after evaporation, is compressed by a compressor, which also increases the temperature of the gaseous process fluid. The heated gaseous process fluid, which is usually present as superheated vapor downstream of the compressor, is then cooled and completely condensed in a high-pressure heat exchanger acting as a condenser for the process fluid in heat exchange with a secondary fluid, whereby the secondary fluid is heated from a return temperature to a supply temperature.The secondary fluid can, for example, be water from a heating circuit or domestic water, or another heat transfer fluid that transports the heat from the high-pressure heat exchanger, in which the heat transfer is transferred from the process fluid to the heat transfer fluid, to a heating circuit and / or a domestic water heater. The cooled and liquefied process fluid is then expanded back to low pressure in a pressure-reducing device, for example and usually a throttle device. In this process, the fluid is self-regulatingly cooled through the expansion and associated partial evaporation to a temperature lower than the temperature of the heat source and returned to the low-pressure heat exchanger, where the process fluid is evaporated again.

[0004] For the sake of completeness, it should be mentioned that, as is known to those skilled in the art, other thermal systems, such as air conditioning systems or refrigeration units, also operate with a fundamentally identical cycle. These may differ from a heat pump in certain cases due to the structural design, in particular of the heat exchangers, and the physical arrangement of the heat exchangers. However, as is easily understandable for those skilled in the art, the problems underlying the present invention and the proposed solutions, even if explained with reference to heat pumps as an example, can readily be transferred to any type of thermal system, such as air conditioning systems or refrigeration units, or applied in connection with them. The subject matter proposed and / or claimed here is therefore generally applicable in the field of thermal systems, in particular for heat pumps, air conditioning systems, and refrigeration units.

[0005] If an idealized Carnot cycle is used to compare heat pump operation, it becomes apparent that ideally—that is, if evaporation occurs at the heat source temperature, condensation occurs at the flow temperature, and compression and pressure reduction occur isentropically—the mass flow-specific compressor power to be supplied to the process fluid is almost proportional to the temperature difference between the flow temperature of the secondary fluid and the temperature of the heat source. Since all temperatures are to be specified as absolute temperatures above absolute zero, and a heat source whose temperature corresponds at least to an ambient temperature prevailing on Earth can generally be provided as the heat source, it becomes apparent that the power required to drive the compressor is significantly lower than the heat power transferred to the secondary fluid.

[0006] Even if these conditions are less favorable in real thermal processes than in an idealized Carnot cycle between a lower and an upper process temperature, this demonstrates the great benefit of heat pumps: namely, that heat present at a low temperature at the heat source is "raised" in the compressor to a higher temperature that can be used for heating purposes. This thermodynamic process is also used, for example, in air conditioning systems and refrigeration units, where the heat in the room or object to be cooled is raised above the temperature of an available heat sink, such as ambient air, and can thus be transferred to the heat sink, even if the temperature of the heat sink is higher than the temperature of the room or object to be cooled.The heating power available through cooling and condensation of the process fluid exceeds the drive power of the compressor, which is typically provided by an electric motor, and thus the electrical power required to drive the compressor, ideally several times over. This ratio of heating power to the absorbed mechanical or electrical power is referred to as the coefficient of performance, and the smaller the difference between the upper and lower process temperatures of the process fluid, the more efficient the heat pump operates.

[0007] Accordingly, even in the idealized Carnot process, the negative effects of lowering the lower process temperature—the temperature at which the process fluid evaporates—and / or increasing the upper process temperature—the temperature at which the process fluid liquefies—are immediately apparent. Even though these relationships are much more complex in real two-phase processes, it is immediately apparent that the lower process temperature should be as high as possible and the upper process temperature as low as possible.

[0008] It is also clear that selecting a process fluid with thermodynamic properties suited to the process is also crucial. The process fluid must exhibit a two-phase region at the expected process temperatures and at technically manageable pressures, thus being capable of evaporation and condensation, as well as exhibiting the highest possible mass-specific enthalpy of vaporization at the desired pressures. For a given mass-specific enthalpy of vaporization, operating the process at low pressures increases the flow cross-sections and / or velocities of the process fluid required to achieve a certain maximum heating output. The same effect results if the mass-specific enthalpy of vaporization is low at a given pressure, as this requires large mass flows per thermal output.High pressures, on the other hand, require thicker walls and pose a latent risk of leakage, making them undesirable for safety reasons. Other thermodynamic properties, such as the isentropic coefficient, also have an influence. For high-pressure refrigerants such as R410a, the limitation of the operating range on the high-pressure and high-temperature sides is very important, as the permissible discharge temperatures are reached much more quickly than with low-pressure refrigerants such as R134a.

[0009] Ammonia, or R-717, offers advantageous thermodynamic properties. While ammonia is also advantageous in terms of its lack of ozone depletion potential and global warming potential, it is, on the other hand, hazardous to water and highly toxic. It is also corrosive, and due to the large specific enthalpy difference, the low power-specific mass flow in low-capacity systems makes system control difficult. The use of ammonia as a process fluid is therefore limited in practice to large-scale plants. Dichlorodifluoromethane, R-12, and all other fluorinated refrigerants are no longer used due to their massive potential for damage to the ozone layer. Tetrafluoroethane is used in the isomer forms R-134 and R-134a. However, its use is increasingly restricted or prohibited by regulations due to its strong greenhouse gas impact.In addition, these substances, like numerous other chemicals used as refrigerants, belong to the so-called PFAS.

[0010] Propane, R290, is increasingly being used in heat pumps. This naturally occurring substance is characterized by its attractive thermodynamic properties, which translate into good efficiency in heat pump systems, as well as its environmental compatibility and low global warming potential. The potential fire hazard can be controlled with state-of-the-art safety measures; moreover, the fill volumes in heat pumps are low, especially in heat pumps for single-family and multi-family homes.

[0011] The minimum temperature to which the high-pressure process fluid must be heated for heat exchange with the secondary fluid or heat sink is the inlet temperature of the secondary fluid. However, since the heat exchange surfaces of the high-pressure heat exchanger cannot be infinitely large, the temperature of the high-pressure process fluid must always be several Kelvin higher than the inlet temperature. The lowest pressure at which this is theoretically possible therefore corresponds to the vapor pressure of the process fluid at the inlet temperature of the secondary fluid and, as indicated above, in reality corresponds to the vapor pressure of the process fluid at a temperature slightly higher than the inlet temperature of the secondary fluid. The maximum specific heat that the high-pressure process fluid can release at constant temperature and constant pressure is the specific latent heat of condensation at the pressure corresponding to the vapor pressure at this temperature.The pressure of the high-pressure process fluid therefore ideally corresponds to the vapor pressure of the high-pressure process fluid at a temperature a few degrees Kelvin higher than that of the secondary fluid feed. Depending on the design of the high-pressure heat exchanger, the process fluid is completely condensed, and the condensate may also be slightly subcooled.

[0012] The highest temperature that the low-pressure process fluid can theoretically reach during heat exchange with the heat source is the temperature of the heat source. Since the heat exchange surfaces of the low-pressure heat exchanger cannot be infinitely large, the temperature of the low-pressure process fluid must always be several degrees Celsius lower than the temperature of the heat source. The highest pressure at which this is theoretically possible therefore corresponds to the vapor pressure of the process fluid at the temperature of the heat source and, as indicated above, in reality corresponds to the vapor pressure of the process fluid at a temperature slightly lower than the temperature of the heat source.

[0013] Accordingly, the difference between the upper process temperature and the lower process temperature is greater than specified by the heat source and the consumer – i.e., the secondary fluid – for reasons of heat transfer alone – a driving temperature gradient must always be present. Furthermore, it must be prevented that liquid process fluid enters the compressor. Likewise, it must be ensured that the refrigerant mass flow can be controlled before entering the low-pressure heat exchanger without the risk of the process fluid in the low-pressure heat exchanger being incompletely evaporated due to the limited dynamics of this control. Therefore, the evaporated process fluid in the low-pressure heat exchanger is additionally superheated by the heat source.To ensure this, the pressure of the low-pressure process fluid is lowered even further below the vapor pressure at the heat source temperature, thus further reducing the lower process temperature of the low-pressure process fluid. This is, for example, 10 K below the heat source temperature. Just a simple consideration of the ideal Carnot cycle outlined above, and the consideration that this also reduces the pressure in the low-pressure region and the compressor has to overcome a larger pressure difference, immediately shows that this negatively impacts the heat pump's coefficient of performance. In addition, the reduction in density and the associated dilution of the suction-side mass flow negatively impact the volumetric cooling capacity and thus the efficiency.

[0014] According to the usual hermetic standard, the compressor unit of a heat pump typically consists of a compressor and an electric motor that drives the compressor, encapsulated within a housing. The housing is hermetically sealed except for an inlet for low-pressure process fluid and an outlet for the process fluid compressed by the compressor. The entire compressor drive is integrated into the compressor unit housing, so that no moving parts penetrate the housing. This prevents leaks from the compressor unit, either of environmentally harmful or toxic process fluids such as R717 or R134 or R134a, or of flammable process fluids such as R290, unless the compressor unit housing is forcibly destroyed. The compressor unit housing encloses a space in which low-pressure or, in other designs, high-pressure levels prevail.In the following, only hermetic compressors with low-pressure casings are considered. The actual compressor, with its suction inlet, is located in the gas chamber, into which the low-pressure process fluid is drawn via a connecting pipe from the evaporator. The compressor is connected to and driven by the drive motor. A lubricant reservoir is located in the lower area of ​​the casing. Lubricant is conveyed by suitable means to the rotor bearings, where it lubricates them. The lubricant then escapes from the bearings and is swirled by the rotation of the rotor. The lubricant droplets enter the gas chamber and / or are thrown onto the stator windings. A large portion of the lubricant droplets return to the lubricant reservoir under the influence of gravity and cohesive forces.A small portion of the lubricant droplets forms a mist and mixes with the suction gas stream, which freely enters the gas space. The compressor draws in the process fluid-lubricant mixture and compresses it. It then travels through the high-pressure side heat exchanger and dissolves in the now liquid process fluid. After passing through the pressure reducing agent, it enters the low-pressure side heat exchanger. The liquid process fluid evaporates, separating the lubricant and the process fluid. The now gaseous process fluid then flows back to the compressor. The liquid lubricant components are carried along by the gas stream and partially dissolve process fluid on their way to the inlet for the low-pressure process fluid, where it re-enters the gas space. In the example mentioned, a fluid connection is established between the gas space and the lubricant reservoir between the rotor and stator of the drive motor.The shaft of the drive motor rotor is often arranged vertically, with the compressor and gas chamber located above the drive motor, while the lubricant reservoir is located below the drive motor. The atmosphere of the process fluid lies above the liquid surface of the lubricant in the lubricant reservoir. Drops of lubricant fall into the lubricant reservoir under gravity. In the process, a portion of the lubricant hits the windings of the drive motor and cools them. When using R134 or R134a, for example, as the process fluid, the heat input from the windings of the drive motor into the lubricant, together with the superheating of the gaseous low-pressure process fluid, is sufficient to degas any process fluid dissolved in the lubricating oil from the lubricating oil, thus separating it from the process fluid.

[0015] However, propane, R290, is so soluble in commonly used lubricants that there is a risk that the process fluid will increasingly dilute the lubricant, which could result in a compressor unit failure. Compressor unit manufacturers specify minimum superheat values ​​for the process fluid. While these are often defined as minimum superheat values ​​at the compressor outlet, they can be converted into superheat values ​​for the gaseous low-pressure process fluid. This shows that when operating a heat pump with R290 as the process fluid and observing the compressor manufacturer's specifications, the gaseous low-pressure process fluid can easily be superheated by 15 K to 20 K above the evaporation or boiling point at the pressure of the low-pressure process fluid.the pressure of the low-pressure process fluid is reduced to such an extent that the temperature at the outlet from the low-pressure heat exchanger or evaporator is 15 K to 20 K above the evaporation or boiling temperature. If we further take into account that a driving temperature gradient of a few K, for example 2 K, must be present in the low-pressure heat exchanger, it is immediately clear that the evaporation temperature of the process fluid, i.e. the lower process temperature, is significantly below the temperature of the heat source. Here, too, the deterioration in the coefficient of performance due to the necessary superheating of the low-pressure process fluid is qualitatively immediately apparent from an examination of the ideal Carnot comparison process outlined above and the consideration that this reduces the pressure in the low-pressure region and the compressor has to overcome a larger pressure difference. DESCRIPTION OF THE SUBJECT OF THE PRESENT DESCRIPTION

[0016] The invention provides a compressor unit, a thermal system comprising the compressor unit, and a method for operating the compressor unit of the type described in the introduction. According to certain aspects of the subject matter described here, the above-mentioned adverse effects of the prior art are to be avoided or at least mitigated. One problem to be solved in specific aspects is to reduce the required superheating of the process fluid at the compressor inlet, particularly when using propane as the process fluid. According to further aspects, the dilution of the lubricant by process fluid dissolved in the lubricant is to be prevented or at least reduced, such that there is no risk of compressor failure due to the accumulation of dissolved process fluid in the lubricant and a resulting reduction in the viscosity of the lubricant.The aim is to ensure that, for example, a heat pump system can be operated with an improved coefficient of performance at given temperatures of the heat source and the heat sink and using a given process fluid.

[0017] The terms heat pump and heat pump system are used synonymously in this document.

[0018] Further effects and advantages of the objects described here, whether explicitly stated or not, will become apparent in the light of the present description.

[0019] This is achieved by means of the compressor unit described in claim 1.

[0020] Accordingly, a compressor unit is described, comprising a compressor, a drive motor, and a housing enclosing the compressor and the drive motor. The compressor is, in particular, a scroll compressor. The drive motor comprises a rotor and a stator. The rotor of the drive motor is operatively connected to the compressor to drive the compressor. For example, the rotor of the drive motor is operatively connected to a rotor of the compressor. Embodiments are conceivable in which the rotor of the drive motor and the rotor of the compressor have a common shaft. In certain embodiments of the compressor unit, the rotor of the drive motor defines an axial direction of the compressor unit, which is parallel to a rotational axis of the rotor, as well as a radial direction and a circumferential direction.The rotor comprises a shaft, for example with a rotationally symmetric permanent magnet package or with a rotationally symmetric asynchronous squirrel-cage rotor. The stator comprises stator windings, wherein in particular the rotor windings are arranged radially within the stator windings and furthermore in particular radially within the entirety of the stator windings. For example, the shaft is mounted on at least one bearing point and in particular on two bearing points within the housing. In embodiments with one bearing point, the motor shaft is additionally supported, for example, via a bearing of the compressor. The housing of the compressor unit encloses a gas chamber and a lubricant reservoir. The lubricant reservoir defines an area inside the housing that is provided to receive a lubricant and in particular to form a lubricant sump therein.Gaseous process fluid can also be present in the lubricant reservoir above a liquid surface of the lubricant sump. The gas space defines an area of ​​the housing interior that is not intended to form a lubricant sump in the gas space. These two areas are clearly distinguishable from one another by those skilled in the art, for example, due to the internal components, their design, and / or their position when properly arranged. Thus, the lubricant reservoir will generally be fluidly connected to a device suitable for conveying the lubricant to points to be lubricated, such as bearings and the like. Likewise, when properly arranged, the lubricant reservoir will be located at the bottom of the compressor unit so that the lubricant can collect there. An inlet opening of the compressor unit opens into the gas space.The compressor is configured and arranged within the gas space to suck fluid from the gas space, compress the fluid, and convey the compressed fluid through an outlet opening of the compressor unit. In certain embodiments, the drive motor is arranged between the gas space and the lubricant reservoir and divides the interior of the housing into the gas space and the lubricant reservoir. More specifically, the gas space and the lubricant reservoir can be arranged at opposite axial ends of the drive motor. The compressor unit is, in particular, intended to be arranged upright, with the lubricant reservoir located below or beneath the drive motor and the gas space defined above the drive motor. The drive motor can, in particular, be arranged with a vertical axis.If reference is made below to an upright installation or arrangement of the compressor unit, this is intended to describe the arrangement described above. In particular, it can further be provided that the axis of the rotor of the drive motor is arranged at least substantially vertically. In this way, the lubricant collects in the lubricant reservoir due to gravity and forms a lubricant sump there. A lubricant conveying device is provided which is arranged and configured to convey a lubricant from the lubricant reservoir, i.e. the storage volume for the lubricant, to the at least two bearing points, from where the lubricant is collected again in the lubricant reservoir by gravity. At least one heat exchanger device is arranged within the lubricant reservoir.This is particularly arranged and configured to effect a heat exchange between a heating medium flowing within the heat exchange device and a lubricant contained in the lubricant reservoir.

[0021] The lubricant droplets ejected from the drive motor rotor shaft have a large surface area, which is why, depending on the conditions, process fluid can easily dissolve in the lubricant droplets. This is particularly the case at relatively low process fluid temperatures. Propane, in particular, is highly soluble in common lubricants. In compressor unit designs, the lubricant droplets fall downward through air gaps in the drive motor, also impacting and cooling the drive motor windings. The ejected lubricant droplets contain process fluid dissolved in the lubricant and are ejected onto the windings and laminated cores, or onto the housing wall. Process fluid escapes from the lubricant when it reaches the boiling point of the process fluid at the current pressure. The evaporation process cools the components inside the housing.In particular, the engine's waste heat is absorbed. Depending on the circumstances, however, this may not be sufficient, and dissolved process fluid may accumulate in the lubricant, ultimately resulting in the reduced lubricant viscosity mentioned above and thus a risk of damage due to bearings. Therefore, especially when operating with propane as the process fluid, it is usually ensured that the temperature of the process fluid at the inlet into the gas space of the compressor unit is sufficiently high to keep the rate of dissolution of process fluid in the lubricant sufficiently low.However, this necessary superheating of the process fluid results, on the one hand, in increased specific compressor work and an increase in the final compression temperature due to the larger specific volume of the process fluid. On the other hand, for example, when operating the compressor unit in a heat pump system, it also requires a reduction in pressure in the low-pressure region of the heat pump system. Evaporation of the process fluid then necessarily occurs at a process fluid temperature that is at least the amount of superheating lower than the heat source temperature, resulting in a larger difference between the upper and lower process temperatures and thus in a loss of efficiency, as explained several times above.

[0022] According to the subject matter described here, a heat exchange device is arranged within the lubricant reservoir. This enables a heating medium to be passed through the heat exchange device, the inlet temperature of which is greater than the boiling point of the process fluid at the pressure prevailing within the compressor unit, for example, the low pressure of a heat pump system. During operation, the heat exchange device is located at least partially within the lubricant sump or is overflowed by lubricant. As a result, process fluid dissolved in the lubricant boils and is thus degassed from the lubricant and reenters the gas space. The process fluid degassed from the lubricant has the boiling point at the prevailing pressure above the lubricant sump formed in the lubricant reservoir and is thus cooler than the superheated process fluid entering the compressor unit.The result is that, despite the addition of heat to the lubricant sump, the temperature of the process fluid at the compressor inlet is lowered. This results in two beneficial effects: first, the necessary superheating of the process fluid at the compressor inlet can be reduced, which in turn increases the lower process temperature of a heat pump system for a given heat source temperature, and second, the specific compressor work relative to the process fluid is reduced. Both effects result in an increase in the efficiency of a heat pump system. In addition, the discharge temperature is lowered, which expands the operating limits.

[0023] In the context of this description, "Ein" or "eine" are to be understood as indefinite articles and not as numerals, unless another meaning is explicitly indicated, for example by the use of "genau ein" or "genau eine".

[0024] In certain embodiments of the compressor unit described here, the lubricant conveying device comprises the shaft of the rotor of the drive motor. The shaft is hollow at least along part of its axial extent, with the hollow interior of the shaft being in fluid communication with the lubricant reservoir. For example, the shaft is hollow starting from an end located adjacent to or in the lubricant reservoir. The hollow interior of the shaft leads in particular to all bearing points and is in fluid communication with them, such that a fluid connection is established from the lubricant reservoir to the bearing points of the shaft. A lubricant propulsion means is arranged in the hollow interior of the shaft and is configured such that, when the shaft rotates in the direction in which the compressor compresses, it conveys lubricant from the lubricant reservoir through the hollow interior of the shaft to the bearing points.

[0025] It can be provided that an intermediate floor is arranged within the lubricant reservoir, which divides the lubricant reservoir into a collecting space and a lubricant chamber. In more specific embodiments of the compressor unit, this division occurs transversely to the axis of the shaft of the rotor of the drive motor. The collecting space is located between the gas space and the lubricant chamber. It can be provided that the collecting space is formed between the windings of the motor and the intermediate floor, while the lubricant chamber is formed on a side of the intermediate floor opposite the collecting space, in particular between the intermediate floor and a floor of the housing. Furthermore, at least one overflow passage is arranged from the collecting space to the lubricant chamber, which forms a fluid connection between the collecting space and the lubricant chamber.A wall delimiting the overflow passage extends from the intermediate floor into the lubricant chamber. The delimiting wall extends in particular at least 50% and in particular at least 70% and furthermore in particular at least 80% or at least 90% of the distance from the intermediate floor to the floor of the housing into the lubricant chamber. The said distance is defined locally, at the point at which the wall is formed. Thus, in embodiments of the compressor unit, an opening of the overflow passage distal to the intermediate floor is located closer to the floor of the housing than to the intermediate floor. In particular, the aim is that the opening of the overflow passage distal to the intermediate floor is immersed in the lubricant sump within the lubricant chamber when the compressor unit is set up as intended during operation of the compressor unit.It is also provided that a suction end of the lubricant conveying device is in fluid communication with the lubricant chamber. In this way, lubricant is conveyed from the lubricant chamber to the bearings. In this context, it can be provided, in particular, that a suction end of the shaft of the rotor of the drive motor is in fluid communication with the lubricant chamber. Thus, the compressor unit is configured so that lubricant can be conveyed from the lubricant chamber through the hollow interior of the shaft to the bearings. In particular, it can further be provided that a riser pipe extends through the intermediate floor and the riser pipe extends from the intermediate floor into the lubricant chamber, and the suction end of the shaft is arranged inside the riser pipe.

[0026] In certain embodiments, the lubricant chamber is at least substantially sealed, except for the overflow passages and the fluid connection to the intake end of the lubricant delivery device, so that the supply of fluids into the lubricant chamber and the discharge of fluids from the lubricant chamber takes place at least substantially exclusively through the overflow passages and via the lubricant delivery device. The term "substantially" refers to leaks that are unavoidable in practice, depending on the design, for example, of the seal at the shaft passage through the intermediate floor. The aforementioned seal is designed in particular as a non-contact seal and furthermore, for example, as a gas gap seal or labyrinth seal, in order to ensure low-wear and maintenance-free operation of the compressor unit over several tens of thousands of operating hours.

[0027] If the temperature of the lubricant in the lubricant chamber is above the boiling point of the process fluid at the process fluid pressure at the inlet to the compressor unit, outgassing process fluid can build up a gas cushion within the lubricant chamber. The pressure of this gas cushion corresponds to the vapor pressure of the process fluid at the temperature of the lubricant in the lubricant chamber and is higher than the pressure in other areas of the compressor unit. This can support the delivery of lubricant to the bearings, particularly through the hollow shaft of the drive motor, and can also increase the tightness of the non-contact seal.

[0028] In non-limiting examples, at least a portion of the heat exchange device is arranged within the overflow passage. The overflow passage can be configured in the shape of an annular gap. It can be provided that at least a portion of the heat exchange device extends spirally and at least substantially coaxially with the annular gap-shaped overflow passage within the overflow passage. As a result, the lubricant, including any process fluid dissolved therein, flows past the heat exchange device as it flows from the collection space into the lubricant chamber and is heated. As a result, if the temperature of the heating medium within the heat exchange device is greater than the boiling point of the process fluid at the prevailing pressure, the dissolved process fluid outgasses from the lubricant.Gas bubbles rising countercurrent to the lubricant within the overflow passage cause strong turbulence within the overflow passage, which in turn results in intensive heat transfer between the heat exchanger and the fluid in the overflow passage. Heated lubricant enters the lubricant chamber, where additional dissolved process fluid may be released, forming a gas cushion in the manner described above. The pressure of this cushion may be higher than that in other areas of the compressor unit.

[0029] As indicated above, the compressor unit can be configured for an upright arrangement, with the lubricant reservoir located at the bottom and the gas chamber at the top. Furthermore, the shaft of the rotor of the drive motor is provided and configured for an at least substantially vertical arrangement.

[0030] According to further aspects of the subject matter described here, a thermal system is proposed which comprises a compressor unit of the type described above. The thermal system is, in particular, a heat pump. The compressor unit is arranged in a flow path between an outlet of a first heat exchanger and an inlet of a second heat exchanger in such a way as to draw in and compress a process fluid from the outlet of the first heat exchanger via the inlet opening of the compressor unit and to convey the compressed process fluid via the outlet opening of the compressor unit to the inlet of the second heat exchanger. For example, the first heat exchanger is an evaporator for the process fluid and the second heat exchanger is a condenser for the process fluid.Furthermore, an outlet of the second heat exchanger is in fluid communication with an inlet of the first heat exchanger, wherein a pressure reducing means is arranged in a flow path between the outlet of the second heat exchanger and the inlet of the first heat exchanger. The pressure reducing means can be a throttle element or a pressure reducing valve. A flow path from the outlet of the second heat exchanger to the pressure reducing means is configured such that, for at least a portion of a process fluid mass flow directed from the second heat exchanger to the pressure reducing means, the heat exchange device of the compressor unit is interposed in the flow path from the second heat exchanger to the pressure reducing means in the flow direction. In particular, it can be provided that the heat exchange device of the compressor unit is tubular and arranged within the compressor unit such that it is surrounded by the lubricant.Accordingly, in this case, at least a portion of the process fluid mass flow directed from the second heat exchanger to the pressure reduction means flows through the heat exchange device of the compressor unit. In exemplary embodiments, it is provided that the entire process fluid mass flow directed from the second heat exchanger to the pressure reduction means flows through the heat exchange device of the compressor unit. In this case, the process fluid guided downstream of the second heat exchanger or the condenser to the heat exchange device of the compressor unit is fluidically separated from the process fluid conveyed and compressed in the compressor unit and the lubricant of the compressor unit, but is in a heat exchange relationship with the lubricant of the compressor unit via the heat exchange device of the compressor unit, i.e.Heat can be transferred between the process fluid conducted downstream of the second heat exchanger to the heat exchange device of the compressor unit and the lubricant of the compressor unit. In particular, the process fluid conducted downstream of the second heat exchanger or the condenser to the heat exchange device of the compressor unit flows to a primary side of the heat exchange device, while the heat exchange device is surrounded on its secondary side by the lubricant of the compressor unit during operation.

[0031] In the aforementioned thermal system, it can further be provided that a phase separation tank is arranged in the flow path between the outlet of the second heat exchanger and the pressure-reducing means. A vapor chamber is arranged in an upper region of the phase separation tank, and a condensate collector is arranged in a lower region of the phase separation tank. A condensate outlet of the phase separation tank is in fluid communication with the condensate collector and opens, in particular, at the bottom of the phase separation tank. The line leading to the pressure-reducing means is connected to the condensate outlet of the phase separation tank.Both the pressure reduction device and the heat exchanger device of the compressor unit are arranged in fluidic series downstream of the condensate outlet of the phase separation tank, with the pressure reduction device being arranged downstream of the heat exchanger device of the compressor unit. This ensures that the heat exchanger device of the compressor unit is supplied on the primary side with liquid process fluid whose temperature corresponds at least to a good approximation to the temperature at which the process fluid condenses under pressure on the high-pressure side of the thermal system. This temperature is naturally higher than the boiling point on the low-pressure side of the thermal system. Thus, process fluid dissolved in the lubricant is degassed from the lubricant on the secondary side of the heat exchanger device of the compressor unit, and the lubricant is thus purified from the process fluid.

[0032] As mentioned above, for example, the first heat exchanger is an evaporator for the process fluid, and the second heat exchanger is a condenser or condenser for the process fluid, which is designed to at least partially or completely liquefy the process fluid. It should be noted that only liquid fluid should be supplied to the pressure-reducing means. Therefore, if only partial condensation occurs in the second heat exchanger, a phase separation vessel, as explained above, is preferably arranged between the second heat exchanger and the pressure-reducing means.

[0033] In further non-limiting examples, a fluid with which the process fluid in the first heat exchanger is in heat exchange relation, i.e. the heat source, is air.

[0034] In further non-limiting examples, a fluid with which the process fluid in the second heat exchanger is in heat exchange relation, i.e., the heat sink, is a heat transfer fluid of a heat consumer. The heat consumer may be, for example, a space heater and / or a domestic water heater, but is not limited to these.

[0035] Particularly in cases where the temperatures of the heat source at the low-pressure heat exchanger and / or the heat sink at the condenser can fluctuate significantly, it is useful to be able to vary the ratio between the high-pressure side pressure of the process fluid at the high-pressure heat exchanger and the low-pressure side pressure of the process fluid at the low-pressure heat exchanger. This is the case in heat pump applications, as the desired flow temperature of a heat consumer can fluctuate depending on the heat demand. It is desirable to keep the flow temperature of the heat consumer and the high-pressure side process fluid at the lowest possible temperature when the consumer's heat demand is low, as this has a positive effect on the coefficient of performance. However, when the consumer's heat demand is high, this temperature often needs to be increased, which requires a relatively high high-pressure side temperature and a relatively high high-pressure side pressure of the process fluid.Heat pumps, in particular, which use the ambient air as a heat source must, of course, be able to cope with strongly fluctuating temperatures at the heat source. Since the boiling point of the process fluid in the low-pressure section of the heat pump should be as little below the temperature of the air serving as the heat source as possible - apart from the superheating of the process fluid required for safe operation of the compressor unit and / or a temperature difference necessary for efficient heat transfer - the pressure of the process fluid in the low-pressure section of the thermal system must also be adjustable. Therefore, embodiments of the thermal system may prove advantageous in which the pressure reducing means is designed for an outlet pressure of the pressure reducing means or downstream of the pressure reducing means that is adjustable independently of the inlet pressure to the pressure reducing means.from the means for pressure reduction (where the above terms are to be understood as equivalent), and a temperature sensor for determining the temperature of the process fluid and a device for determining the pressure of the process fluid are arranged downstream of the first heat exchanger and upstream of the compressor unit. A controller is configured to adjust the outlet pressure of the first heat exchanger as a function of the measured temperature of the process fluid downstream of the first heat exchanger and upstream of the compressor unit such that the required superheating of the gaseous low-pressure process fluid is precisely achieved downstream of the first heat exchanger and upstream of the compressor unit.

[0036] Furthermore, a method for operating a compressor unit in a thermal plant is proposed, comprising using a compressor to draw in and compress a process fluid from a gas space within the compressor unit. The process fluid flows into the gas space via a fluid inlet with a total inlet pressure and a boiling point corresponding to the total inlet pressure. The total pressure corresponds to the static pressure plus the dynamic pressure of the fluid. It thus corresponds to the pressure that would be measured in a Pitot tube.The method further comprises driving the compressor by means of a drive motor, conveying lubricant to shaft bearings of a rotor of the drive motor and / or shaft bearings of a rotor of the compressor to lubricate the shaft bearings, with the lubricant exiting the shaft bearings, collecting the lubricant in a lubricant reservoir, and then conveying the lubricant back to the shaft bearings. Heat is supplied to the lubricant in the lubricant reservoir from a heat source whose temperature corresponds at least to the boiling point of the process fluid at the inlet total pressure, thus degassing the process fluid dissolved in the lubricant.

[0037] With the features described above, it is possible, even when the process fluid has a high solubility in the compressor unit's lubricant, to limit the superheating of the process fluid at the outlet of the low-pressure heat exchanger to a level necessary to enable the control process of the pressure reduction agent by creating a minimum hysteresis. Compared to the prior art, it is no longer necessary to configure a significantly increased superheat to limit the dissolution of process fluid in the lubricant. The reduction in the viscosity of the lubricant due to dissolved process fluid can be effectively prevented by degassing the process fluid from the lubricant by applying heat to the lubricant reservoir.

[0038] The specific embodiments mentioned above can be combined with one another. Further, non-specifically disclosed embodiments of the teachings of this document will be readily apparent to those skilled in the art. SHORT DESCRIPTION OF THE CHARACTERS

[0039] The facts presented here are explained in more detail below using selected exemplary embodiments shown in the drawings. Fig. 1 shows an exemplary state diagram of a process fluid in which the pressure is plotted logarithmically against the mass-specific enthalpy and wherein the illustrated region comprises a part of the wet vapor region of the process fluid; Fig. 2 shows an exemplary embodiment of a compressor unit proposed here; Fig. 3 shows a detail of the compressor unit from Figure 2 ; Fig. 4 an exemplary element for propelling lubricant in a lubricant conveying device in the compressor unit of Fig. 2; and Fig. 5 shows an example of a thermal plant in which a compressor of the type proposed here is integrated.

[0040] Some of the drawings are highly schematic. Details not necessary for understanding the described objects have been omitted. Furthermore, the drawings show only selected embodiments and should not be used to limit the scope of the claims. Embodiments not shown may well be covered by the claims. EXAMPLES OF IMPLEMENTATION

[0041] In Figure 1is a diagram in which the pressure p is plotted logarithmically against the mass-specific enthalpy h for an exemplary fluid that can be used as a refrigerant, hereinafter also referred to as the process fluid. The diagram includes the wet steam region designated 100, which lies below the shifted sugarloaf-shaped curve designated 110. The diagram will be familiar to those skilled in the art. To the left of the wet steam region 100, the fluid is in the liquid state, to the right of it in the gaseous state, and above the critical point 120 in a supercritical state. In the wet steam region, the liquid and gaseous phases exist side by side. Some isotherms and some isentropies are shown, which will also be familiar to those skilled in the art. In the wet steam region, the isotherms run horizontally, i.e. on lines of constant pressure. This pressure is the vapor pressure at the temperature represented by the isotherm.The isotherm T AMB represents, for example, the ambient air temperature of an air-to-water heat pump, or, in other words, the temperature of the heat source. The process fluid of a heat pump is to be evaporated using heat absorbed from the ambient air. Since a driving temperature gradient must exist for heat transfer from the heat source to the process fluid, the evaporation of the process fluid occurs at temperature Tv, which must be lower than the temperature T AMB of the heat source. The temperature difference is, for example, 2 K. For this to happen, the process fluid must be at pressure pv, which corresponds to the vapor pressure at temperature Tv. In an ideal process, the process fluid is exactly in the saturated vapor state at state 1 after evaporation. The process fluid is then compressed in a compressor to state 2. If the compression is idealized and loss-free, the change of state occurs along an isentropy.The pressure pc to which the gaseous working fluid is compressed is determined by the required hot gas temperature Tc, the minimum possible value of which is determined, for example, by the flow temperature of a heat consumer, i.e., the temperature of a heat transfer fluid of the heat consumer after heating by the heat pump. The heat is to be transferred to this heat transfer fluid during the transition from state 2 to state 3. During the change of state from state 2 to state 1, the compressed process fluid initially releases sensible heat to the heat transfer fluid while cooling. Once the two-phase region is reached, latent heat is released through condensation. State 3 ideally lies on the boiling point of the wet steam region. The process fluid is then expanded to state 4, whereby the pressure drops from pc to pv and the temperature from Tc to Tv.The process fluid can then be evaporated again from state 4 to state 1 by adding ambient heat. During compression from state 1 to state 2, the compressor applies the mass-specific mechanical work Δh C related to the process fluid, while the heat transferred to the heating circuit, corresponding to the difference between the mass-specific enthalpies of states 2 and 3 related to the process fluid, is significantly greater than the mass-specific mechanical work Δh C related to the process fluid by the compressor.

[0042] In reality, as explained above, it is necessary to superheat the gaseous low-pressure process fluid before compression, i.e. to add further heat after evaporation. The manufacturers of compressor units for heat pump systems specify minimum values ​​for the necessary superheating of the process fluid before and / or after the compressor. However, the maximum achievable temperature in the heat exchange with the heat source, e.g. the ambient air, remains Tv. This means that the evaporation temperature Tv,x in the process with superheating of the low-pressure process fluid must be lower than the ambient temperature T AMB by an amount that is the temperature difference necessary for heat transfer plus the value of the required superheat of the low-pressure process fluid. In the illustrated example, evaporation and superheating occur at temperature Tv,x and the vapor pressure pv,x corresponding to this temperature from state 4x to state 1x.The compression from state 1x to state 2x requires a mass-specific compressor work Δh C ,x related to the process fluid, which is greater than the mass-specific compressor work Δh C related to the process fluid from state 1 to state 2 in the process without superheating of the low-pressure process fluid. It is readily apparent that a larger proportion of the mass-flow-specific usable heat related to the process fluid comes from compressor work during the state change from state 2x to state 3 than during the state change from state 2 to state 3 in the process without superheating. In other words, in a process with superheating of the low-pressure process fluid, a larger proportion of the usable heat must be put into the process as compressor work than in a process without superheating or with less superheating of the low-pressure process fluid.The efficiency of the heat pump process, measured by the coefficient of performance, decreases with increasing superheat of the low-pressure process fluid. This effect is accentuated during real, lossy, and non-isentropic compression of the process fluids.

[0043] The statements made also apply analogously to cooling processes.

[0044] Another factor influencing the necessary superheating of the low-pressure process fluid is the solubility of the process fluid in the compressor unit's lubricant. There is a risk that the process fluid will dissolve and accumulate in the lubricant, thereby reducing the lubricant's viscosity, which could ultimately result in a compressor unit failure. The solubility of the process fluid in the lubricant decreases as the temperature of the process fluid rises above the vapor pressure. Therefore, especially when using R290, i.e., propane, the superheat must be increased further according to the state of the art. For example, when using propane as the process fluid, the evaporation temperature Tv,x may be required to be 20 K or even more below the temperature T AMB of the heat source, which, as is readily understandable, results in significant efficiency losses.Reducing the required superheating of the low-pressure process fluid, regardless of the process fluid used, is therefore a highly efficient measure to improve the coefficient of performance of heat pumps.

[0045] Therefore, the information provided in connection with the Figure 2 and 3The compressor unit 10 described above is proposed. The compressor unit 10 comprises a housing 1, which is closed except for an inlet opening 16 and an outlet opening 17. The actual compressor 12 and a drive motor 13 for driving the compressor are arranged within the housing. The drive motor 13 comprises a stator 131 with stator windings and a rotor 132, which is designed, for example, as a BLDC rotor or as a squirrel-cage rotor, with a shaft 133. The shaft 133 is mounted in plain bearings 137 and 138 and is coupled to a rotor of the compressor 12 to drive the compressor 12. A gas chamber 14 and a lubricant reservoir 15 are arranged adjacent to the end faces of the drive motor.The compressor unit 10 shown as an example is intended for an upright arrangement, wherein the shaft 133 of the rotor 132 of the drive motor, which defines an axis of the drive motor, is arranged at least substantially vertically and the gas chamber 14 is arranged above the drive motor 13 and the lubricant reservoir 15 is arranged below the drive motor 13. As can be seen, the lubricant reservoir is not completely filled with lubricant, but rather gaseous process fluid is also located above a liquid surface of the lubricant. The compressor 12 draws in process fluid 201a from the gas chamber 14 via its intake opening 121, with gaseous low-pressure process fluid 201 entering the gas chamber through the inlet opening 16 of the compressor unit. Compressed gaseous high-pressure process fluid 202 exits the compressor via the outlet opening 17.

[0046] A lubricant sump made of lubricant 301, for example, a suitable lubricating oil, is formed in the lubricant reservoir 15. The lubricant reservoir 15 is divided by an intermediate floor 112 into a collecting space 151 and a lubricant chamber 152. The collecting space 151 is formed above the intermediate floor 112 between the intermediate floor 112 and the drive motor 3, while the lubricant chamber 152 is formed below the intermediate floor 112 between the intermediate floor 112 and a floor 111 of the housing 11 of the compressor unit. Thus, the collecting space 151 lies between the gas space 14 and the lubricant chamber 152. At least one overflow passage 153, in the non-limiting example shown, an annular gap-shaped overflow passage 153, establishes a fluid connection between the collecting space 151 and the lubricant chamber 152. The overflow passage with the visible fittings is connected to the Figure 3 explained in more detail.

[0047] To supply the bearings 137 and 138 with lubricating oil, the shaft 133 of the drive motor is provided with a hollow shaft interior 134. The hollow shaft interior 134 is in fluid communication with the lubricant chamber 152 via a suction end 135 of the shaft. Adjacent to the suction end 135 of the shaft 133, a lubricant propulsion means 136 is arranged within the shaft 133. The lubricant propulsion means 136 is in Figure 4shown larger. The lubricant propulsion means 136 is essentially a twisted sheet metal element that forms a conveying spiral within the shaft 133 and is arranged such that, when the rotor 132 rotates in a direction of rotation in which the compressor 12 is driven to compress process fluid, it conveys lubricant 301 from the lubricant chamber 152 through the hollow shaft interior 134 to the bearings 137 and 138. Lubricant escaping from the bearings 137 and 138 is partially thrown off and forms lubricant droplets in an environment of process fluid. Due to the upright arrangement of the compressor unit 10 with the oil reservoir 15 located below, the lubricant droplets fall back into the collection chamber 151 and the lubricant sump. Due to the large surface-to-volume ratio of the lubricant droplets, these have great potential to absorb and dissolve process fluid.Lubricant droplets that fall back from the gas space 14 into the lubricant reservoir 15 come into contact with the motor windings of the drive motor 13 and cool them, whereby process fluid dissolved in the lubricant droplets is outgassed from the lubricant due to the heat absorption.

[0048] However, it has been shown that for process fluids whose solubility in the lubricant exceeds a certain level, this process is no longer sufficient to reliably prevent the accumulation of dissolved process fluid in the lubricant at temperatures of the gaseous low-pressure process fluid 201 that are, for example, very close to the saturated steam temperature of the low-pressure process fluid 201. Therefore, according to the state of the art, when using such process fluids, such as R290, the gaseous low-pressure process fluid 201 is superheated to such an extent that its solubility in the lubricant is sufficiently reduced, with the effects on the efficiency of a heat pump process described above.

[0049] With the proposed compressor unit, this is not required and superheating of the low-pressure process fluid is only necessary to the extent necessary to enable the control process of the pressure reduction means and to prevent liquid process fluid from entering the intake port 121. For example, when using R290, propane, as the process fluid, the superheating of the gaseous low-pressure process fluid can be reduced from 20K to 6K. When considering the Figure 1 From the above statements, it becomes clear that this results in a significant improvement in the efficiency of a thermal system, such as a heat pump. When used as intended, the proposed compressor unit does not pose a problem if the lubricant returning to the lubricant reservoir 15 initially contains comparatively high proportions of dissolved process fluid. Figure 3shows a view of the Figure 2Details marked III. As explained above, the intermediate floor 112 divides the lubricant reservoir 15? into the upper collecting space 151 and the lower lubricant chamber 152. The collecting space 151 and the lubricant chamber 152 are connected to one another by at least one overflow passage, in the exemplary embodiment shown, the annular gap-shaped overflow passage 153. The overflow passage 153 extends from the intermediate floor 112 downwards into the lubricant chamber 152. In the exemplary embodiment, the overflow passage is formed by the outer wall of the housing 11 of the compressor unit and an annular wall 154, which extends from the intermediate floor 112 downwards or into the lubricant chamber.The annular wall 154, which radially inwardly delimits the overflow passage 153, extends in particular at least 50% and in particular at least 70% and furthermore in particular at least 80% or at least 90% of the distance from the intermediate floor to the floor of the housing into the lubricant chamber. The opening of the overflow passage defined thereby should be located so far below the intermediate floor that, during operation of the compressor unit, it lies below the level of the lubricant 301 in the lubricant chamber 152. A heat exchange device 18 is arranged in the lubricant reservoir 15, and specifically within the overflow passage 153. In this embodiment, this consists of a heat exchange tube arranged spirally within the annular overflow passage 153. It is provided that a heating medium 203 flows through the heat exchange device 18, the temperature of which is greater than the boiling point of the low-pressure process fluid 201.As discussed below in connection with . Figure 5, liquefied process fluid is primarily used as the heating medium, which is passed through the heat exchange device. Lubricant 301 from the collection chamber 151 flows through the overflow passage 153 into the lubricant chamber 152. The lubricant comes into contact with the heat exchange device 18 and, in heat exchange with the heating medium 203 inside the heat exchange tube, is heated to a temperature that is at least locally above the boiling point of the low-pressure process fluid 201. As a result, the process fluid dissolved in the lubricant 301 boils and separates from the lubricant 301 or is outgassed from the lubricant 301. The gaseous process fluid thus separated from the lubricant 301 emerges from the surface of the lubricant 301 in the collection chamber 151 and flows through the air gaps of the drive motor 13 into the gas chamber 14. In the process, it mixes with the freshly flowing gaseous low-pressure process fluid 201.Since the inflowing gaseous low-pressure process fluid 201 is superheated, but the process fluid outgassed from the lubricant has at least approximately the boiling point of the low-pressure process fluid 201, cases can occur in which the process fluid 201a sucked in by the compressor has a lower temperature than the low-pressure process fluid 201 flowing into the compressor unit, which further reduces the specific compressor work required. Due to the limited boiling rate, a portion of the process fluid dissolved in the lubricant 301 only outgasses from the lubricant 301 within the lubricant chamber 152, where it forms a gas cushion. The pressure of this gas cushion is upper-limited at the latest by the vapor pressure of the process fluid at the temperature of the inflowing heating medium 203 and depends on the performance of the heat exchanger device 18.From the intermediate floor 112, a riser pipe 19 extends downward below the surface of the lubricant 301 in the lubricant chamber 152. The intake end 135 of the shaft 133 projects into the riser pipe from above. The gas cushion in the lubricant chamber 152 is at a higher pressure than the pressure in the collecting chamber 151 and in the gas chamber 14 of the compressor unit, and thus also at a higher pressure than the hollow shaft interior 134. As a result, the gas cushion pushes a column of lubricant upward into the riser pipe 19, thus assisting the delivery of lubricant to the bearing points 137 and 138. Cooled heating medium 203a flows out of the heat exchange device 18.

[0050] In the Figure 5A heat pump system is shown as an example of a thermal system that integrates the compressor unit 10 described above. Those skilled in the art will readily be able to transfer this example to other thermal systems, such as air conditioning systems or refrigeration units. The compressor unit 10 compresses gaseous low-pressure process fluid 201 and thereby generates gaseous high-pressure process fluid 202 downstream of the compressor unit 10. The gaseous high-pressure process fluid 202 is fed into the high-pressure heat exchanger 20 and releases heat to a heating circuit, indicated at 20a. The cooled high-pressure process fluid exiting the high-pressure heat exchanger 20 is at least partially condensed. The high-pressure heat exchanger 20 is therefore also referred to as a condenser. The cooled high-pressure process fluid 203' flows into a phase separation tank 21.There, the liquid phase settles at the bottom, while any remaining gas phase remains in the upper part or is returned to the high-pressure heat exchanger. At the bottom of the phase separation vessel 21, or at least below the surface of the liquid phase, liquefied high-pressure process fluid 203 flows out. Its temperature corresponds at least essentially to the condensation temperature of the process fluid at high pressure, but may also be slightly subcooled. In principle, the condensation temperature of a fluid and its boiling temperature at a specific pressure are identical and equivalent. In any case, the process is designed so that the temperature of the liquefied high-pressure process fluid 203 upon exiting the high-pressure heat exchanger 20 or the phase separation vessel 21 is higher than the boiling temperature of the low-pressure process fluid.The liquefied high-pressure process fluid 203 is introduced as a heating medium into the heat exchange device 18 of the compressor unit 10. At this point, it can in principle also be provided that only a partial mass flow of the liquefied high-pressure process fluid 203 is fed to the heat exchange device 18 of the compressor unit 10, for example, if an excessively large, uncontrolled throttling effect in the heat exchange device 18 is to be feared. In the heat exchange device 18 of the compressor unit 10, the liquid high-pressure process fluid cools in heat exchange with the lubricant in the compressor unit 10. As mentioned, the temperature of the liquefied high-pressure process fluid 203 is higher than the boiling point of the low-pressure process fluid.Therefore, in the heat exchange with the liquefied high-pressure process fluid 203 in the lubricant of the compressor unit 10, process fluid dissolved is heated above its boiling point and thus outgassed from the lubricant of the compressor unit, as described in connection with the . Figure 2 and 3explained. This prevents dilution of the compressor unit's lubricant due to an accumulation of process fluid dissolved therein. Cooled liquid high-pressure process fluid 203a flows from the heat exchange device 18 of the compressor unit 10 to a controllable throttle element 30, which acts as a pressure-reducing means. The cooled low-pressure process fluid 203a expands in the controllable throttle element 30. This expansion generally occurs in the wet vapor region of the process fluid. Downstream of the controllable throttle element 30, the process fluid is present as a low-pressure process fluid in the wet vapor state 204. This is introduced into the low-pressure heat exchanger 40, also referred to as an evaporator, and flows through a finned heat exchanger 40a suitable for heat exchange with air.In the presently shown embodiment, an air flow is directed over the finned heat exchanger by means of a fan 40b, from which the heat is extracted for the evaporation of the low-pressure process fluid. Examples are also known in which a fan is omitted and convection currents of the ambient air are utilized. Solar collectors, soil, brine, or groundwater could also be used as heat sources, although this list is by no means exhaustive.

[0051] Sensors 43 and 44 measure the temperature and pressure of the process fluid downstream of the low-pressure heat exchanger 40. For clarity, sensors 43 and 44 are shown offset in the flow direction of the process fluid, but this may not necessarily reflect the actual arrangement. With a proper design of the finned heat exchanger and sufficient air flow over the finned heat exchanger, the temperature of the process fluid is slightly below the temperature of the overflowing air serving as the heat source. The measured values ​​from sensors 43 and 44 are fed to the adjustable throttle device 30.The adjustable throttle device 30 adjusts the mass flow through the adjustable throttle device 30 and, in connection with this, the pressure downstream of the adjustable throttle device such that these values ​​are just large enough to evaporate all of the moisture in the two-phase low-pressure process fluid 204 in the low-pressure heat exchanger 40, and the gaseous low-pressure process fluid 201 with the desired superheat is present downstream of the low-pressure heat exchanger 40. Control interventions on the adjustable throttle device 30 are only registered downstream of the low-pressure heat exchanger 40 with a certain delay. Therefore, a certain superheating of the gaseous low-pressure process fluid 201 downstream of the low-pressure heat exchanger 40 is also desired in order to ensure, despite this limited dynamics of the control loop, that downstream of the low-pressure heat exchanger 40 there is actually only gaseous low-pressure process fluid.In comparison to supplying the liquefied high-pressure process fluid directly to the controllable throttle device 30, in the present embodiment, the humidity of the two-phase low-pressure process fluid 204 is increased, or the proportion of gaseous process fluid after expansion in the throttle device is reduced. The enthalpy of the process fluid, reduced due to the comparatively higher humidity, can be supplied to the process fluid "free of charge" through evaporation in the low-pressure heat exchanger 40 in heat exchange with the ambient air; if necessary, appropriate dimensioning of the heat exchange surfaces of the plate heat exchanger 40a is required. However, the heat transfer in the low-pressure heat exchanger 40 is also improved due to the higher humidity, and more heat from the heat source can be supplied to the process fluid in the low-pressure heat exchanger 40.

[0052] Although the subject matter of the present description has been explained using selected exemplary embodiments, these are not intended to limit the claimed invention. The claims encompass embodiments not explicitly illustrated, and embodiments that deviate from the examples shown are nevertheless covered by the claims.

Claims

1. Compressor unit (10), comprising a compressor (12), a drive motor (13) and a housing (11) enclosing the compressor and the drive motor, wherein the drive motor comprises a rotor (132) and a stator (131), wherein the rotor of the drive motor is operatively connected to the compressor (12) for driving the compressor, wherein the housing (11) of the compressor unit (10) encloses a gas space (14) and a lubricant reservoir (15), wherein an inlet opening (16) of the compressor unit opens into the gas space (14), and the compressor (12) is configured and arranged within the gas space to suck in fluid (201a) from the gas space, compress the fluid, and convey the compressed fluid (202) through an outlet opening (17) of the compressor unit, wherein a lubricant conveying device is further provided, which is arranged and configured is to convey a lubricant (301) from the lubricant reservoir (15),and wherein at least one heat exchange device (18) is arranged within the lubricant reservoir., 2. Compressor unit according to the preceding claim, wherein the drive motor (13) is arranged between the gas space (14) and the lubricant reservoir (15) and divides the interior of the housing (11) into the gas space (14) and the lubricant reservoir (15).

3. Compressor unit according to one of the preceding claims, wherein the lubricant conveying device comprises the shaft (133), wherein the shaft is hollow along at least part of its axial extent and the hollow interior (134) of the shaft (133) is in fluid communication with the lubricant reservoir (15), and wherein a lubricant propulsion means (136) is arranged in the hollow interior of the shaft.

4. Compressor unit according to one of the preceding claims, wherein an intermediate floor (112) is arranged within the lubricant reservoir (15), which divides the lubricant reservoir into a collecting space (151) and a lubricant chamber (152), wherein the collecting space (151) lies between the gas space (14) and the lubricant chamber (152) and wherein at least one overflow passage (153) is arranged from the collecting space to the lubricant chamber, which forms a fluid connection between the collecting space and the lubricant chamber, wherein a wall (154) delimiting the overflow passage extends from the intermediate floor (112) into the lubricant chamber (152) and wherein a suction end (135) of the lubricant conveying device is in fluid connection with the lubricant chamber (152).

5. Compressor unit according to claim 4 in conjunction with claim 3, wherein a suction end (135) of the shaft (133) is in fluid communication with the lubricant chamber (152) and in particular a riser pipe (19) extends through the intermediate floor (112) and the riser pipe (19) extends from the intermediate floor (112) into the lubricant chamber (152), wherein the suction end (135) of the shaft (133) is arranged inside the riser pipe.

6. Compressor unit according to one of claims 4 or 5, wherein at least a part of the heat exchange device (18) is arranged within the overflow passage (153).

7. Compressor unit according to one of claims 4 to 6, wherein the overflow passage (153) is designed in an annular manner.

8. Compressor unit according to the preceding claim, wherein at least a part of the heat exchange device (18) is spiral-shaped and arranged in the annular overflow passage (153).

9. Compressor unit according to one of the preceding claims, wherein the compressor unit is configured for an upright arrangement, wherein the lubricant reservoir (15) is arranged below and the gas space (14) is arranged above.

10. A thermal plant comprising a process fluid circuit, in which process fluid circuit a compressor unit (10) according to one of the preceding claims is arranged in a flow path between an outlet of a first heat exchanger (40) and an inlet of a second heat exchanger (20) in such a way as to suck in and compress a process fluid (201) from the outlet of the first heat exchanger (40) via the inlet opening (16) of the compressor unit (10) and to convey the compressed process fluid (202) via the outlet opening (17) of the compressor unit (10) to the inlet of the second heat exchanger (20), and wherein an outlet of the second heat exchanger is further fluidically connected to an inlet of the first heat exchanger, wherein a pressure reducing means (30) is arranged in a flow path between the outlet of the second heat exchanger (20) and the inlet of the first heat exchanger (40),wherein a flow path from the outlet of the second heat exchanger (20) to the pressure reducing means (30) is configured such that, for at least a portion of a process fluid mass flow directed from the second heat exchanger to the pressure reducing means, the heat exchange device (18) of the compressor unit (10) is interposed in the flow path from the second heat exchanger to the pressure reducing means in the flow direction.

11. Thermal plant according to the preceding claim, wherein a phase separation tank (21) is arranged in the flow path between the outlet of the second heat exchanger (20) and the means for pressure reduction (30), wherein a vapor space is arranged in an upper region of the phase separation tank and a condensate collector is arranged in a lower region of the phase separation tank, wherein both the means for pressure reduction (30) and the heat exchange device (18) of the compressor unit are arranged downstream of the condensate outlet of the phase separation tank (21).

12. Thermal plant according to one of claims 10 or 11, wherein the first heat exchanger (40) is an evaporator for the process fluid and the second heat exchanger (20) is a condenser for the process fluid.

13. Thermal plant according to one of claims 10 to 12, wherein a fluid with which the process fluid in the first heat exchanger (40) is in heat exchange relation is air and a fluid with which the process fluid in the second heat exchanger (20) is in heat exchange relation is a heat transport fluid (20a) of a heat consumer.

14. Thermal plant according to one of claims 10 to 13, wherein the pressure reducing means (30) is configured for an adjustable outlet pressure of the pressure reducing means, wherein a temperature sensor (43) for determining the temperature of the process fluid (201) and a device (44) for determining the pressure of the process fluid (201) are arranged downstream of the first heat exchanger (40) and upstream of the compressor unit (10), wherein a controller is configured to adjust the outlet pressure of the first heat exchanger (40) as a function of the measured temperature of the process fluid downstream of the first heat exchanger and upstream of the compressor unit.

15. A method for operating a compressor unit (10) in a thermal plant, comprising, by means of a compressor (12), sucking in and compressing a process fluid (201a) from a gas space (14) within the compressor unit, wherein the process fluid (201) flows into the gas space (14) via a fluid inlet (16) with an inlet total pressure and a boiling temperature corresponding to the inlet total pressure, wherein the method comprises driving the compressor (12) by means of a drive motor (13), conveying lubricant (301) to shaft bearings (137, 138) of a rotor (132) of the drive motor and / or shaft bearings of a rotor of the compressor in order to lubricate the shaft bearings, wherein the lubricant exits the shaft bearings, and collecting the lubricant again in a lubricant reservoir (15) in order to subsequently convey the lubricant back to the shaft bearings, wherein the procedure also includesto supply heat to the lubricant in the lubricant reservoir from a heat source whose temperature corresponds at least to the boiling point of the process fluid at the inlet total pressure and thus to degas the process fluid dissolved in the lubricant.

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