Energy efficiency waste heat utilization with screw expanders

Transcription

1 Energy efficiency waste heat utilization with screw expanders Prof Dr-Ing Andreas Brümmer In a vehicle about one third of the energy stored in the fossil fuel is used for propulsion The remaining two thirds are lost as waste heat Regarding the increase in efficiency of combustion engines, the future use of this thermal energy plays a decisive role An advantageous thermal power process is the Rankine cycle which is also commonly used for electricity generation in stationary power plants In the simplest case, this cycle process consists of a pump, a preheater, an evaporator and, where appropriate, of a superheater, the expansion machine and a condenser After the pressure increase through the pump, the working fluid is preheated by the heat source (eg exhaust tract of a combustion engine) up to the boiling curve where it is then evaporated Afterwards, the vapor is expanded in the expansion machine while producing power before it is liquefied again and returned to the pump In order to maximize thermal efficiency (useful power output referred to the thermal power transferred to the working fluid) for this cyclic process, the averaged temperature of the heat supply must be as high as possible and that of the heat dissipation as low as possible The averaged heat supply temperature can be realized through high operating pressure and maximum superheating The temperature of heat dissipation is predetermined by the heat sink which exists for cooling purposes At the same time, the counter pressure of the expansion machines results from this temperature, which except for slight pressure losses corresponds to the vapour pressure of the working fluid at this temperature However, the maximization of the above-mentioned thermal efficien cy should not be equated with the maximization of the useful power output The reason for this is the magnitude of the thermal output transferred to the working fluid during the cyclic process, which in case of maximum superheating can be smaller than for instance without superheating This is the case when the mass flow of the working fluid is limited by the available heat flow of the waste gas between pinch point and the point of maximum superheating, for example due to the required temperature gradient towards the working fluid Exemplary studies on the utilization of waste heat on commercial vehicles confirm that an increased operating pressure without superheating is advantageous for a large-scale transmission of the thermal output existing in the exhaust gas flow, for example, to the working fluid Ethanol, as well as for a maximization of the useful power output At the expansion machine inlet saturated steam exists which is prone to condensation during the expansion process 120 Pumps, Compressors and Process Components 2012

2 Due to the liquid to be expected inside the expansion machine, screw expanders the functional principle and describing parameters of which will be presented in the following, are particularly suitable for these applications With the help of basic design variants, the properties of these fluid energy machines are demonstrated and the systematic and targeted procedure for their design will be described Functioning and parameters s are displacement machines without clearance volume, the working chambers of which are formed by the tooth gaps of two helically toothed gears, the rotors The volume of a working chamber depending on the rotation angle is exemplarily demonstrated in Fig 1 If a tooth of the female rotor on the high-pressure side (HP-side) unscrews from the tooth space of the male rotor, a working chamber arises With progressing rotation of the rotor, the volume of this working chamber increases up to a maximum before it is reduced again by a repeated tooth engagement or by the low-pressure side at the front edge, and the working chamber finally disappears The rotors are tightly enclosed by a casing which provides the working chambers with external sealing and sealing towards the front edge The casing houses the inlet and outlet ports, the limitations of which are called control edges The outlet port in the front edge casing of the low-pressure side enables the working fluid to be discharged usually from the maximum chamber volume V max up to the disappearance of the working chamber In the range in which the chamber is formed the inlet port represents a connection to the highpressure side of the screw expander This connection is maintained up to a selectable chamber volume V Ex,th (Fig1) The ratio is called internal volume ratio v i and is one of the essential parameters of a screw machine The ideal comparison process for a screw expander consists of an isobaric chamber filling with the inlet pressure p E and the inlet density ρ E (highpressure side) up to V Ex,th, an isentropic expansion up to the discharge pressure p A (low-pressure side) and an isobaric discharge (Fig 2) [2] The isentropic work W s this process delivers per working cycle together with the actually performed inner work per working cycle forms the compression work ratio W i /W s The volumetric efficiency λ L with describes the ratio of system mass flow m A to theoretical delivery mass flow m th, with the product of male rotor speed n MR and male rotor number of teeth z MR being the working cycle frequency Another important parameter of fluid energy machines is the inner isentropic efficiency η is with In this case, the isentropic power P s corresponds to the product of specific isentropic work w s and system mass flow m A and P i of the indicated or internal power of the screw expander In connection with the mechanical efficiency η m the effective output power P e and the effective isentropic efficiency η es results in Fig 1: Volume curve as well as inlet and discharge areas of a screw-type expander over the rotational angle of the male rotor (asymmetric SRM profile; v i = 8; z MR /z FR = 4/6; ϕ MR = 200 ; L/D = 14) Fig 2: Indicator diagram of a screw expander optimized with regard to the output power and ideal comparison process (not represented from V max ) (asymmetric SRM profile; v i = 8 ; z MR /z FR = 4/6; ϕ MR = 200 ; L/D = 14) Pumps, Compressors and Process Components

3 The volumetric efficiency of a screw expander is influenced by two opposing effects On the one hand, inlet throttling occurs during the chamber filling, since it is only for a short time that the surface of the inlet port is completely available for the filling of the chamber (Fig 1) In particular, the decreasing inlet surface in connection with the increasing chamber volume, immediately before the inlet control edge for the closure of the chamber is reached, results in a chamber pressure below the inlet pressure p E at the point of the theoretical start of expansion at V Ex,th This effect reduces the volumetric efficiency λ L On the other hand, due to the design-related gaps of the screw expander, the gap mass flows cause an increase in volu metric efficien cy In this regard, the profile gap, which partly provides a direct connection of the working chambers towards the outlet side, and the blow hole, the dimensions of which are independent of the rotation angle, dominate the front clearance at the high-pressure side as well as the casing clearance during the chamber filling and at the beginning of the expansion Depending on the weighting of these two effects, the volu metric efficiency of a screw expander can be greater than, equal to or less than one so that the informative value of the volumetric efficiency with regard to an evaluation of the expansion machine is limited In the event of an external pressure ratio Π a adapted to the inner volume ratio the volumetric efficiency describes the utilization of the working chamber volume and thus indirectly also that of the design volume with regard to the system mass flow In this situation, the compression work ratio analogously describes the energetic utilization of the given working chamber volume If in addition, the delivered mass flow corresponds to the theoretical mass flow (λ L =1), the compression work ratio also describes the energetic utilization of the isentropic power - provided by the system mass flow It is only in this special case that the compression work ratio is identical to the inner isentropic efficiency The inner isentropic efficiency is suitable for the evaluation of the energy conversion of the screw expander It quantifies the proportion of the isentropic power the system mass flow provides to the expander which is transferred to the rotors in a mechanically useful way Accordingly, the inner isentropic efficiency des cribes the energetic losses, for example, through the inflow and discharge of the working fluid, the gap flows and the heat flows In this connection, it is initially confusing that the inner isentropic efficiency of a geometrically defined screw expander can be increased through a decline in chamber filling thus through the decrease of the volu metric efficiency The physical reason for this is the reduction of the isentropic power provided due to the reduced system mass flow, which in this case outweighs the reduction of the mechanically useful internal power which is transferred to the rotors Consequently, the internal power of the expander decreases with an increasing inner isentropic efficiency of the energy conversion Therefore, it is not the energetic efficiency of the expansion machine, but the output power in general that needs to be maximized to achieve an energy-efficient design of a screw expander for a cyclic process to be able to use the wasteheat flow of defined thermal power Design variants The by far predominant part (approx 90 %) of the screw compressors sold is unsynchronized To ensure adequate lubrication of the rotors and bearings as well as other advantages (e g cooling, increase in volumetric general efficiency, reduction of acoustic emission), an auxiliary fluid (usually oil) is injected into the working chambers of these machines Afterwards, the oil is separated from the working fluid by means of oil separators and is fed back to the machine in a separate cycle With this method however there is always a small amount of oil remain ing in the working fluid due to the limited separation rates of the separators Even though in (still) rather rare cases, water is used alternatively to oil, which then requires special rotor materials and/or coatings Also, a dry-running unsynchronized screwtype supercharger has already been realized and successfully operated for research purposes (< 1000 operating hours) Here, however, the tribological stress of the rotors had to be limited by special profiling and operation at low to medium compression ratios In case of synchronized screw machines, as an alternative to the unsynchronized design, it is usually a timing gear that prevents rotor contact In general, in these machines there is no liquid working fluid to be found in the working chamber so the result is an increased gap mass flow In order to reduce these mass and energy losses, synchronized screw compressors are operated with significantly increased circumferential speeds at the crown circle With air being used as working fluid, the common circumferential speeds of the so-called dry runners range between m/s; in case of wet runners, they range between m/s due to the splashing losses These basic differences make it necessary to decide upon the variant that will later be realized before the screw expander is designed In this respect, it needs to be clarified whether a timing gear is required and a relevant liquid content in the working chambers is to be expected The answer to the first question depends substantially on the lubrication properties of the fluid in the working chambers As long as this fluid, consisting of the working fluid and for example of the liquid injected into the machine, ensures adequate lubrication of the rotors from the screw expander point of view the unsynchronized design is to be preferred In this case, the lighter and smaller constructional size, the simpler assembly and the energetically more favorable behavior of the unsynchronized screw expanders provide great advantages The latter is explained in particular by the minimization of the profile gap and thus of the gap losses Regarding the answer to the second question, it needs to be considered that a certain part of liquid in the screw expander has generally pro 122 Pumps, Compressors and Process Components 2012

4 ved to be advantageous The liquid can be injected directly into the working chamber or can arise during expansion through condensation of the working fluid Since a forecast regarding the chronological course of the condensation of the working fluid within the screw expander seems to be only insufficiently possible and so far hardly controllable at a local level, from the screw expander s point of view the targeted injection of liquids is desirable For this purpose, either oil or the working fluid in liquid form can be used Consequently, the design of a screw expander is oriented towards an unsynchronized, liquid injection machine with the rotors and bearings being lubricated by the working fluid and the injected liquid Despite the comprehensive experience gained with such constructions in the field of compressors (compressed-air and refrigeration technology) the design of a suitable screw expander for waste heat utilization on combustion engines represents a challenge Compared to the application fields of the compressors, the maximum temperatures of the working fluid are significantly higher The exhaust gas temperatures of Diesel motors for example are still clearly over 250 C even behind the exhaust gas aftertreatment system At these temperatures, there is a risk that oil existing in the Rankine cycle is destroyed in the heat exchanger Due to the limited separation rates mentioned before, even an oil separator would not basically change this problem However, the oil separator would have the advantage of making the separated oil directly available for the lubrication of bearings and for injection Moreover, the amount of oil transported in the cycle would not be permanently heated in the heat exchanger and cooled down in the condenser without being relevantly involved in the useful energy conversion Exemplary calculations for heat utilization on a commercial vehicle, as mentioned in the introduction, show that without an oil separator about 5 % of the available thermal output would get lost due to the cyclic heating and cooling of an oil amount favorable for the screw expander [2] that is contained in the working fluid In general, it is also possible to refrain from oil in the cycle In this case, it should be examined whether the working fluid, for example injected in liquid form, ensures the adequate lubri cation of the rotors, the movable bearing at the low-pressure side (LP side) and the fixed bearing at the highpressure side (HP side) Here, the tribological demands are usually rising in the specified order Depending on the result, one of the design variants mentioned in Table 1 is to be chosen In Fig 3, the SK1D variant is shown as an example ENERGYEFFICIENCY IT S A MATTER OF THE OPERATING POINT As one of the world s leading manufacturers in the industry, BECKER sets the pace in boosting energy efficiency in using industrial vacuum and compressed air Discover our innovative VARIAIR technology that - improves the efficiency of products and their components - provides the right solution for optimising installations between air user and air generator - controls air generators in line with actual air demand VASF series helping to keep you ahead W W W B E C K E R - I N T E R N AT I O N A L C O M

5 Design timing gearbox overhung bearing shaft sealing opposite to working chamber HP side LP side UK UK1D UK2D UF1D SK1D SK2D SF1D Table 1: Design variants U or S K or F 1D or 2D UK UK1D UK2D UF1D SK1D unsynchronized or synchronized classical fixed/movable bearing or overhung bearing sealings between working and bearing chamber at the HP side or HP and LP side desirable variant, with bearings and rotor being lubricated by the working fluid and/or by an injected auxiliary fluid (e g oil) Compared to UK, the bearings at the HP side require special lubrication, mixing the bearing lubrication with the working fluid is to be avoided Compared to UK1D, also the bearings at the LP side require special lubrication and sealing towards the working chamber Compared to UK2D, it is possible to renounce special lubrication and, in part, costly sealing of the bearings at the LP side due to an overhung bearing Compared to UK1D, rotor contact is prevented by means of a timing gear The lubrication of this gear and of the bearings at the HP side needs to be sealed towards the working chamber The bearings at the LP side are either encapsulated (grease lubrication) or lubricated by the working fluid (media lubrication) (cf Fig 2) SK2D/SF1D Analogous to UK2D/UF1D, however rotor contact is prevented by a timing gear Exhaust gas/flue gas Rankine-Cycle Composition Working fluid Mass flow max operating pressure/temperature Temperature and pressure Condensation temperature/pressure min temperature differences in the heat exchangers Table 2 : Data required for the design of a screw expander for heat utilization on a commercial vehicle (exemplary extraction) x MR (FR), y MR (FR) tooth profile for male (MR) and female rotor (FR) z MR (FR) tooth number of male and female rotor v i internal volume ratio L/D length to diameter ratio ϕ MR wrap angle male rotor h Gap gap heights of different gaps Table 3: Geometric parameters of a screw expander The design of a screw expander for the utilization of waste heat with predefined thermal output and temperature consists of two phases In phase I, the machine geometry along with the system conditions are optimized according to the previously defined objectives [4] For the application mentioned, the preferred target figure is the effective output power instead of efficien cy The selection of a geometry is followed by phase II, by the calculation of the characteristic maps and by the detailed engineering The data given in Table 2 should be available for phase I These data enable the calculation of the possible system mass flow from the thermodynamic simulation of the vapor generator (preheater, evaporator, superheater) for a selected combination of inlet pressure and temperature at the screw expander In the following step, an initially arbitrarily dimensioned screw expander is generated for one of the data sets given in Table 3 For this machine, the flow- and thermodynamic simulation for a selected constant circumferential speed and/or constant rotational speed at the crown circle provides the absolute dimensions and the indicated power of the expander For this purpose, the arbitrarily dimensioned machine is iteratively scaled in order to convey the predefined system mass flow Except for the predetermined absolute gap heights, this scaling occurs in a geometrically similar way Due to the variation of the system parameters (inlet pressure and temperature, circumferential speed of the male rotor) and the geometric parameters, this finally results in the curve of the indicated power via circumferential speed and/or rotational speed of the male rotor for a large number of screw expanders with different geometries Considering the mechanical power loss, the combination of system parameters and machine geometry results which lets expect a maximum effective power output Here, the effects of a liquid phase within the expander, for example due to the selection of the gap heights, must be taken into account when calculating the mechanical losses [5] Such an approach was realized as an example for waste heat recovery on a commercial vehicle 1 Here, for reasons of simplicity, it was assumed that there was no relevant amount of liquid in the working chambers (dry runner) A schematic representation of the selected design variant (SK1D) is to be found in Figure 3 and the related indicator diagram in Figure Pumps, Compressors and Process Components 2012

6 Fig 3: Exemplary presentation of the design variant SK1D Regarding the system parameters, it is evident that, for the reasons initially mentioned, a superheating of the working fluid Ethanol is disadvantageous Furthermore, an increased inlet pressure (here 40 bar) in connection with an increased internal volume ratio (here v i = 8) turns out to be an advantage Due to the large internal volume ratio, the inlet area of the screw expander is small (Fig 1) In order to never theless realize a sufficient chamber filling, a supercritical influx is advantageous Then how ever the maximum chamber pressure is considerably below the applied inlet pressure The compression work ratio as well as the volumetric efficiency is correspondingly low At the same time, the dimensions of the machine are increasing in relation to the inlet pressure so that the internal power reaches its maximum With regard to other geometrical parameters it is to be noticed that a relatively small wrap angle of e g ϕ = 200 is advantageous and the combination of teeth numbers at male and female rotor as well as the L/D-ratio play a rather minor role concerning the maximum possible internal power [4] The effective output power of this screw expander calculated in design phase II is shown in Figure 4 over the system mass flow for different MR speeds Here, a reduced system mass flow at fixed speed can only be realized via reduced inlet pressure Accordingly, the output power falls linearly with the mass flow Due to the fixed internal volume ratios, overexpansion inside the screw expander occurs with decreasing inlet pressure Apart from friction losses, the compres sion work required to compensate overexpansion offsets the effective area in the indicator diagram in case of diminishing Fig 4: Effective output power over system mass flow for different MR speeds (inlet pressures) Pumps, Compressors and Process Components

7 output power All in all, this results in an optimum combination of inlet pressure and speed for each sys tem mass flow with regard to the effective output power for the machine with fixed geometry Conclusion In order to increase energy efficiency, the focus is more and more on the utilization of waste heat, for example, on combustion engines For these applications, the Rankine cycle turned out to be advantageous In view of the maximization of the thermal power transferred to the working fluid, in some applications superheating of the working fluid can and should be renounced The expansion of the vapour inside the expansion machine then begins at the saturated vapour line so that liquid is to be expected in the expander The screw expander is particularly suitable for these applications When designing this machine type however some special features must be considered A generally advan tageous design for the screw machine is that with two rotors that roll off on each other and that are lubricated by the working fluid, and, if necessary, by an auxiliary fluid (e g oil) along with the bearings Moreover, it is not the efficien cy of the screw expander, but the output power in connection with variable system parameters and the geometry of the expander that should be optimized For predefined system mass flows, this results in screw expanders which provide maximum output power with small compression work ratios and volumetric efficiencies References [1] Konka, K-H: Schraubenkompressoren: Technik und Praxis VDI-Verlag GmbH, 1988 (ISBN ) [2] Zellermann, R: Optimierung von Schraubenmotoren mit Flüssigkeitseinspritzung Dissertation FG Fluidenergiemaschinen, Universität Dortmund, 1996 [3] Brümmer, A; Hütker, J: Influence of geometric parameters on inletlosses during the filling process of screw-type motors Developments in mechanical engineering, vol 4, 2011, pp [4] Hütker, J; Brümmer, A: A comparative examination of steam-powered screw motors for specific installation conditions 8 VDI-Fachtagung Schraubenmaschinen 2010 (2010), VDI Bericht 2101, S [5] Deipenwisch, R; Kauder, K: Oil as a design parameter in screw-type compressors: oil distribution and power losses caused by oil in the working chamber of a screw-type compressor, IMechE 1999, p The study was carried out within the framework of the Project FVV No Expansion Machines, Phase I (Preliminary Study) of the Forschungsvereinigung Verbrennungskraftmaschinen (FVV) [Research Association for Combustion Engines] We thank the FVV for its support Author: Prof Dr-Ing Andreas Brümmer, FG Fluidtechnik, TU Dortmund 126 Pumps, Compressors and Process Components 2012